Sound Insulation
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Sound Insulation
Carl Hopkins
AMSTERDAM • BOSTON • HEIDELBERG • LONDON • NEW YORK • OXFORD
PARIS • SAN DIEGO • SAN FRANCISCO • SINGAPORE • SYDNEY • TOKYO
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First edition 2007
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Preface
The most effective approach to sound insulation design involves the use of measured data
along with statistical and/or analytical models, blended with a combination of empiricism, experience, and pragmatism. Engineering design is predominantly experiential in nature; applying
past experience to new problems. This often impedes rapid progress when applying a general
knowledge of acoustics to a specific area such as sound insulation. This book is intended for
students, engineers, consultants, building designers, researchers and those involved in the
manufacture and design of building products. It uses theory and measurements to explain
concepts that are important for the application, interpretation and understanding of guidance
documents, test reports, product data sheets, published papers, regulations, and Standards.
The intention is to enable the reader to tackle many different aspects of sound insulation by
providing a textbook and a handbook within a single cover. Readers with a background in
acoustics can jump straight to the topic of interest in later chapters and, if needed, return to
earlier chapters for fundamental aspects of the theory. This book draws on a wealth of published literature that is relevant to sound insulation, but it does not document a historical review
of every incremental step in its development, or cover all possible approaches to the prediction
of sound transmission. The references will provide many starting points from which the reader
can dive into the vast pool of literature themselves.
All prediction models and measurement methods have their limitations, but with knowledge of
their strengths and weaknesses it becomes much easier to make design decisions and to find
solutions to sound insulation problems. A model provides more than just a procedure for calculating a numerical result. The inherent assumptions should not simply be viewed as limitations
from which use of the model is quickly dismissed; the assumptions may well shed light on a
solution to the problem at hand. The fact that we need to deal with a relatively wide frequency
range in building acoustics means that there is no single theoretical approach that is suitable
for all problems. We rarely know all the variables; but the simple models often identify the ones
which are most important. A model provides no more than the word implies; in this sense, every
model is correct within the confines of its assumptions. The models and theories described in
this book have been chosen because of the insight they give into the sound transmission process. There are many different approaches to the prediction of sound transmission. In general,
the more practical theories are included and these will provide the necessary background for
the reader to pursue more detailed and complex models when required. Occasionally a more
complex theory is introduced but usually with the intention of showing that a simpler method
may be adequate. Choosing the most accurate and complex model for every aspect of sound
transmission is unnecessary. There are so many transmission paths between two rooms in a
building that decisions can often be made whilst accepting relatively high levels of uncertainty
in the less important transmission paths. Before doing any calculations it is worth pausing for a
second to picture the scene at the point of construction. The uncertainty in describing a building
construction tends to differ between walls, floors, or modular housing units that are produced
on a factory assembly line compared with a building site that is exposed to the weather and
a wide range in the quality of workmanship. Uncertainty is often viewed rather negatively as
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being at the crux of all design problems, but uncertainty in the form of statistics and probability
is part of many theoretical solutions for the transmission of sound and vibration in built-up
structures. To avoid sound insulation problems in a completed building, uncertainty needs to
be considered at an early stage in the design process. In the words of Francis Bacon (Philosopher, 1561–1626) “If a man will begin with certainties, he shall end in doubts, but if he will be
content to begin with doubts, he shall end in certainties’’.
Sound not only travels via direct transmission across the separating wall or floor, but via
the many other walls and floors, as well as via other building elements such as cavities,
ceiling voids, and beams; we refer to this indirect transmission as flanking transmission. To
predict both direct and flanking transmission, statistical models based upon Statistical Energy
Analysis (SEA) are particularly practical because they tend to make gross assumptions about
the building elements. This is important because specific details on the material properties and
dimensions are not always available in the early (and sometimes late) stages of the design. In
fact, during the construction phase a variety of similar building products are often substituted
for the one that was originally specified. In addition the quality of workmanship can be highly
variable within a single building, let alone between different buildings. SEA or SEA-based
models allow an assessment of the different sound transmission paths to determine which
paths are likely to be of most importance. These models are also attractive because laboratory sound insulation measurements of complex wall or floor elements can be incorporated
into the models. Some construction elements or junction details are not well suited to SEA
or SEA-based models. Analytical models and finite element methods therefore have a role to
play, although they tend to be more orientated to research work due to the time involved in
creating and validating each model. Engineers involved in laboratory measurements become
painfully aware that significant changes in the sound insulation can sometimes be produced
by small changes to the test element (such as extra screws, different layouts for the framework, or different positions for the porous material in the cavity). Hypersensitivity to certain
changes in the construction can often be explained using statistical or deterministic models;
but the latter may be needed to help gain an insight into the performance of one specific test
element.
Sound insulation tends to be led by regulations, where the required performance is almost
always described using a single integer number in decibels. It is common to draw a ‘line in the
sand’ for an acceptable level of sound insulation, such that if the construction fails to achieve
this by one decibel, the construction is deemed to have failed. This needs to be considered in
the context of a specific pair of rooms in one specific building; we can rarely predict the sound
insulation in a specific situation to plus or minus one decibel. However, we can often make
reasonable estimates of the average sound insulation for a large number of nominally identical
constructions. The design process must therefore consider the sound insulation that can be
provided on average, as well as the performance of individual constructions on a particular site.
The fact that we need to design and predict the sound insulation in the field on a statistical
basis does not mean that we can accept low precision measurements; quite the opposite.
Every decibel is important to the builder having the sound insulation measured to check that
the building satisfies the regulations; to the manufacturer who wants to claim an advantage
over a competitor’s product; to the engineer trying to assess sound transmission mechanisms
with laboratory measurements; to the designer trying to choose between two different building
products, and to the house builder trying to reduce costs by avoiding over specification in the
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design. It is important to keep in perspective those situations where the highest level of accuracy
is necessary, along with those situations where a rough estimate is more than sufficient. The
intention is to provide the reader with a background from which they can decide the appropriate
level for the problem at hand. In an engineering context the words ‘reasonable’ and ‘adequate’
will quite often be used to describe equations, prediction models, assumptions, and rules of
thumb.
Overview of contents
Chapters 1 and 2 deal with theoretical aspects of sound fields in spaces and vibration fields
on structures. Sound transmission in buildings is fundamentally concerned with the coupling
between these fields. For the reader who is relatively new to acoustics it should be sufficient
to start these chapters with a basic background in acoustics terminology, wave theory, and
room acoustics. It is assumed that the reader has more experience in room acoustics, or is
perhaps more comfortable with these concepts. Sound and vibration are discussed in a similar
style so that the reader can see the similarities, and the many differences, between them. The
layout of these chapters is intended to simplify its use as a handbook when solving problems
that are specific to room acoustics, vibration in buildings, and sound insulation. Sound and
vibration fields are described in terms of both waves and modes. It is useful to be able to think
in terms of waves and modes interchangeably, taking the most convenient approach to solve
the problem at hand.
Chapter 3 looks at sound and vibration measurements relating to sound insulation and material
properties. This chapter deals with the underlying theory behind the measurements, and the
reasons for adopting different measurement methods. This chapter forms a bridge between
the sound and vibration theory in Chapters 1 and 2 and the prediction of sound insulation in
Chapters 4 and 5. However, it is not possible to explain all aspects of measurements without
referring to some of the theory in Chapters 4 and 5. Some readers may choose to start the
book in Chapter 3 and it will sometimes be necessary to refer forward as well as back.
Chapter 4 looks at direct sound transmission across individual building elements. Sound and
vibration theory from Chapters 1 and 2 is combined with material property measurements from
Chapter 3 to look at prediction models for different sound transmission mechanisms. There is no
single theoretical model that can deal with all aspects of sound insulation. Many constructions
are so complex that reliance is ultimately placed on measurements. The aim of this chapter is
to give insight and understanding into sound transmission for relatively simple constructions.
These form a basis from which measurement, prediction, and design decisions can be tackled
on more complex constructions. The chapter is based around prediction using SEA augmented
by classical theories based on infinite plates. Some aspects of sound transmission are not
suitable for SEA models, but the SEA framework can conveniently be used to highlight these
areas such that other models can be sought.
Chapter 5 concerns sound insulation in situ where there is both direct and flanking transmission.
Prediction of vibration transmission across idealized plate junctions is used to illustrate issues
that are relevant to measurement and prediction with other types of plates and more complex
junction connections. Following on from Chapter 4 the application of SEA is extended to the
prediction of direct and flanking transmission. In addition, a simplification of SEA results in an
SEA-based model that facilitates the inclusion of laboratory sound insulation measurements.
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Generalizations
In building acoustics the main frequency range used to assess sound insulation lays between
the 100 and 3150 Hz one-third-octave-bands; an optional extended frequency range is defined
between the 50 and 5000 Hz one-third-octave-bands. In this book the range between 50 and
5000 Hz will be referred to as the building acoustics frequency range. For sound and vibration
in buildings, it is possible to describe many general trends by defining the low-, mid-, and
high-frequency ranges using one-third-octave-band centre frequencies as follows:
• Low-frequency range : 50–200 Hz.
• Mid-frequency range : 250–1000 Hz.
• High-frequency range : 1250–5000 Hz.
The only exact boundaries in these ranges correspond to the 50 and 5000 Hz bands; the
intermediate boundaries need to be considered with a degree of flexibility; usually within plus
or minus one-third-octave-band.
It is also useful to try and define a range of room volumes that are typically encountered in
buildings. For the purpose of making general statements it will be assumed that ‘typical rooms’
have volumes between 20 and 200 m3 ; this covers the majority of practical situations.
Constructions throughout the world are primarily built with concrete, masonry, timber, steel,
glass and plasterboard. In a very general sense, the term ‘heavyweight’ is used for concrete, masonry, and heavy-steel elements and ‘lightweight’ is used for timber, glass,
plasterboard, and light-steel elements. There are also combinations that form a separate
lightweight/heavyweight category, such as timber floors with a surface layer of concrete screed.
For generic and proprietary materials that are commonly used in lightweight and heavyweight
constructions, material properties, and sound insulation values are included in this book to
help the reader get a feel for realistic values and to assess general trends.
viii
Contents
Acknowledgements
xix
List of symbols xxi
1. Sound fields
1.1 Introduction 1
1.2 Rooms 1
1.2.1 Sound in air 1
1.2.1.1 Complex notation 3
1.2.1.2 Plane waves 4
1.2.1.3 Spherical waves 6
1.2.1.4 Acoustic surface impedance and admittance 7
1.2.1.5 Decibels and reference quantities 8
1.2.1.6 A-weighting 9
1.2.2 Impulse response 10
1.2.3 Diffuse field 11
1.2.3.1 Mean free path 12
1.2.4 Image sources 15
1.2.4.1 Temporal density of reflections 15
1.2.5 Local modes 17
1.2.5.1 Modal density 19
1.2.5.2 Mode count 22
1.2.5.3 Mode spacing 23
1.2.5.4 Equivalent angles 24
1.2.5.5 Irregularly shaped rooms and scattering objects 24
1.2.6 Damping 26
1.2.6.1 Reflection and absorption coefficients 27
1.2.6.2 Absorption area 29
1.2.6.3 Reverberation time 29
1.2.6.3.1 Diffuse field 31
1.2.6.3.2 Non-diffuse field: normal mode theory 34
1.2.6.3.3 Non-diffuse field: non-uniform distribution of absorption 41
1.2.6.4 Internal loss factor 42
1.2.6.5 Coupling loss factor 42
1.2.6.6 Total loss factor 42
1.2.6.7 Modal overlap factor 42
1.2.7 Spatial variation in sound pressure levels 44
1.2.7.1 Sound fields near room boundaries 45
1.2.7.1.1 Perfectly reflecting rigid boundaries 45
1.2.7.1.2 Other boundary conditions 51
1.2.7.2 Sound field associated with a single mode 52
1.2.7.3 Excitation of room modes 56
1.2.7.4 Diffuse and reverberant fields 57
1.2.7.5 Energy density 59
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1.2.7.6
1.2.7.7
1.2.7.8
1.2.7.5.1 Diffuse field 59
1.2.7.5.2 Reverberant sound fields with non-exponential
decays 61
Direct sound field 62
Decrease in sound pressure level with distance 62
Sound fields in frequency bands 65
1.2.7.8.1 Below the lowest mode frequency 65
1.2.7.8.2 Reverberant field: below the Schroeder cut-off
frequency 66
1.2.7.8.3 Reverberant field: at and above the Schroeder cut-off
frequency 70
Statistical description of the spatial variation 71
75
Energy density near room boundaries: Waterhouse correction 76
1.2.7.9
1.2.8 Energy
1.2.8.1
1.3 Cavities 77
1.3.1 Sound in gases 77
1.3.2 Sound in porous materials 78
1.3.2.1 Characterizing porous materials 79
1.3.2.1.1 Porosity 79
1.3.2.1.2 Airflow resistance 80
1.3.2.1.3 Fibrous materials 81
1.3.2.2 Propagation theory for an equivalent gas 82
1.3.3 Local modes 87
1.3.3.1 Modal density 87
1.3.3.2 Equivalent angles 89
1.3.4 Diffuse field 90
1.3.4.1 Mean free path 92
1.3.5 Damping 92
1.3.5.1 Reverberation time 92
1.3.5.2 Internal losses 92
1.3.5.2.1 Sound absorption coefficient: Locally reacting porous
materials 94
1.3.5.3 Coupling losses 96
1.3.5.4 Total loss factor 96
1.3.5.5 Modal overlap factor 96
1.3.6 Energy 97
1.4 External sound fields near building façades 97
1.4.1 Point sources and semi-infinite façades 97
1.4.1.1 Effect of finite reflector size on sound pressure levels near the façade
1.4.1.2 Spatial variation of the surface sound pressure level 102
1.4.2 Line sources 104
References 107
2. Vibration fields
2.1 Introduction 111
2.2 Vibration 111
2.2.1 Decibels and reference quantities 112
2.3 Wave types 112
2.3.1 Quasi-longitudinal waves 114
2.3.1.1 Thick plate theory 117
x
101
C o n t e n t s
2.4
2.5
2.6
2.7
2.8
2.9
2.3.2 Transverse waves 118
2.3.2.1 Beams: torsional waves 118
2.3.2.2 Plates: transverse shear waves 121
2.3.3 Bending waves 123
2.3.3.1 Thick beam/plate theory 133
2.3.3.2 Orthotropic plates 135
2.3.3.2.1 Profiled plates 136
2.3.3.2.2 Corrugated plates 139
2.3.3.2.3 Ribbed plates 139
Diffuse field 140
2.4.1 Mean free path 140
Local modes 141
2.5.1 Beams 141
2.5.1.1 Bending waves 142
2.5.1.2 Torsional waves 144
2.5.1.3 Quasi-longitudinal waves 145
2.5.1.4 Modal density 145
2.5.2 Plates 148
2.5.2.1 Bending waves 149
2.5.2.2 Transverse shear waves 149
2.5.2.3 Quasi-longitudinal waves 150
2.5.2.4 Modal density 150
2.5.3 Equivalent angles 152
Damping 154
2.6.1 Structural reverberation time 154
2.6.2 Absorption length 156
2.6.3 Internal loss factor 157
2.6.4 Coupling loss factor 158
2.6.5 Total loss factor 159
2.6.6 Modal overlap factor 159
Spatial variation in vibration level: bending waves on plates 160
2.7.1 Vibration field associated with a single mode 160
2.7.2 Nearfields near the plate boundaries 161
2.7.3 Diffuse and reverberant fields 167
2.7.4 Reverberant field 167
2.7.5 Direct vibration field 168
2.7.6 Statistical description of the spatial variation 172
2.7.7 Decrease in vibration level with distance 173
Driving-point impedance and mobility 178
2.8.1 Finite plates (uncoupled): Excitation of local modes 181
2.8.2 Finite plates (coupled): Excitation of global modes 182
2.8.3 Infinite beams and plates 185
2.8.3.1 Excitation in the central part 185
2.8.3.2 Excitation at the edge 189
2.8.3.3 Finite beams and plates with more complex cross-sections
Sound radiation from bending waves on plates 197
2.9.1 Critical frequency 197
2.9.2 Infinite plate theory 198
2.9.3 Finite plate theory: Radiation from individual bending modes 202
2.9.4 Finite plate theory: Frequency-average radiation efficiency 209
2.9.4.1 Method No. 1 209
2.9.4.2 Method No. 2 210
189
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C o n t e n t s
2.9.4.3 Method No. 3 (masonry/concrete plates)
2.9.4.4 Method No. 4 (masonry/concrete plates)
2.9.4.5 Plates connected to a frame 213
2.9.5 Radiation into a porous material 213
2.9.6 Radiation into the soil 214
2.9.7 Nearfield radiation from point excitation 214
2.10 Energy 217
References 217
212
212
3. Measurement
3.1 Introduction 221
3.2 Transducers 221
3.2.1 Microphones 221
3.2.2 Accelerometers 222
3.2.2.1 Mounting 223
3.2.2.2 Mass loading 223
3.3 Signal processing 224
3.3.1 Signals 224
3.3.2 Filters 227
3.3.2.1 Bandwidth 227
3.3.2.2 Response time 230
3.3.3 Detector 230
3.3.3.1 Temporal averaging 230
3.3.3.2 Statistical description of the temporal variation 232
3.4 Spatial averaging 234
3.4.1 Spatial sampling of sound fields 235
3.4.1.1 Stationary microphone positions 235
3.4.1.2 Continuously moving microphones 237
3.4.2 Measurement uncertainty 238
3.5 Airborne sound insulation 239
3.5.1 Laboratory measurements 239
3.5.1.1 Sound intensity 242
3.5.1.1.1 Low-frequency range 243
3.5.1.2 Improvement of airborne sound insulation due to wall linings,
floor coverings, and ceilings 245
3.5.1.2.1 Airborne excitation 245
3.5.1.2.2 Mechanical excitation 245
3.5.1.3 Transmission suites 246
3.5.1.3.1 Suppressed flanking transmission 247
3.5.1.3.2 Total loss factor 249
3.5.1.3.3 Niche effect 253
3.5.2 Field measurements within buildings 258
3.5.2.1 Reverberation time 260
3.5.2.2 Sound intensity 261
3.5.3 Field measurements of building façades 262
3.5.3.1 Sound insulation of building elements 262
3.5.3.1.1 Loudspeaker method 262
3.5.3.1.2 Sound intensity 264
3.5.3.1.3 Road traffic method 266
3.5.3.1.4 Aircraft and railway noise 266
3.5.3.2 Sound insulation of façades 267
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3.6
3.7
3.8
3.9
3.5.4 Other measurement issues 268
3.5.4.1 Background noise correction 268
3.5.4.2 Converting to octave-bands 270
3.5.4.3 Comparing the airborne sound insulation measured using
sound pressure and sound intensity 270
3.5.4.4 Variation in the sound insulation of an element due to moisture
content and drying time 271
3.5.4.5 Identifying sound leaks and airpaths 271
Impact sound insulation (floors and stairs) 272
3.6.1 Laboratory measurements 273
3.6.1.1 Improvement of impact sound insulation due to floor
coverings 273
3.6.1.1.1 Heavyweight base floor (ISO) 274
3.6.1.1.2 Lightweight base floors (ISO) 274
3.6.2 Field measurements 275
3.6.3 ISO tapping machine 275
3.6.3.1 Force 276
3.6.3.2 Power input 283
3.6.3.3 Issues arising from the effect of the ISO tapping machine
hammers 284
3.6.3.4 Modifying the ISO tapping machine 291
3.6.3.5 Rating systems for impact sound insulation 292
3.6.3.6 Concluding discussion 297
3.6.4 Heavy impact sources 298
3.6.5 Other measurement issues 301
3.6.5.1 Background noise correction 301
3.6.5.2 Converting to octave-bands 301
3.6.5.3 Time dependency 301
3.6.5.4 Dust, dirt, and drying time 303
3.6.5.5 Size of test specimen 303
3.6.5.6 Static load 303
3.6.5.7 Excitation positions 305
Rain noise 305
3.7.1 Power input 305
3.7.2 Radiated sound 311
3.7.3 Other measurement issues 311
Reverberation time 313
3.8.1 Interrupted noise method 313
3.8.2 Integrated impulse response method 314
3.8.3 Influence of the signal processing on the decay curve 318
3.8.3.1 Effect of the detector 318
3.8.3.2 Effect of the filters 321
3.8.3.2.1 Forward-filter analysis 322
3.8.3.2.2 Reverse-filter analysis 326
3.8.4 Evaluation of the decay curve 327
3.8.5 Statistical variation of reverberation times in rooms 329
Maximum Length Sequence (MLS) measurements 333
3.9.1 Overview 334
3.9.2 Limitations 339
3.9.2.1 Temperature 340
3.9.2.2 Air movement 341
3.9.2.3 Moving microphones 341
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C o n t e n t s
3.10 Sound intensity 342
3.10.1 p–p sound intensity probe 344
3.10.1.1 Sound power measurement 346
3.10.1.1.1 Measurement surfaces 346
3.10.1.1.2 Discrete point and scanning measurements 347
3.10.1.2 Error analysis 348
3.11 Properties of materials and building elements 353
3.11.1 Airflow resistance 353
3.11.2 Sound absorption 354
3.11.2.1 Standing wave tube 354
3.11.2.2 Reverberation room 354
3.11.3 Dynamic stiffness 355
3.11.3.1 Resilient materials used under floating floors 356
3.11.3.1.1 Measurement 357
3.11.3.1.2 Calculation of dynamic stiffness 359
3.11.3.2 Wall ties 361
3.11.3.2.1 Measurement 361
3.11.3.2.2 Calculation of dynamic stiffness 364
3.11.3.3 Structural reverberation time 364
3.11.3.4 Internal loss factor 365
3.11.3.5 Quasi-longitudinal phase velocity 367
3.11.3.6 Bending phase velocity 370
3.11.3.7 Bending stiffness 371
3.11.3.8 Driving-point mobility 371
3.11.3.9 Radiation efficiency 373
3.12 Flanking transmission 375
3.12.1 Flanking laboratories 376
3.12.1.1 Suspended ceilings and access floors 376
3.12.1.2 Other flanking constructions and test junctions 377
3.12.2 Ranking the sound power radiated from different surfaces 380
3.12.2.1 Vibration measurements 380
3.12.2.2 Sound intensity 380
3.12.3 Vibration transmission 384
3.12.3.1 Structural intensity 384
3.12.3.1.1 a–a structural intensity probe 388
3.12.3.1.2 Structural power measurement 389
3.12.3.1.3 Error analysis 390
3.12.3.1.4 Visualizing net energy flow 391
3.12.3.1.5 Identifying construction defects 395
3.12.3.2 Velocity level difference 396
3.12.3.2.1 Stationary excitation signal and fixed power input 397
3.12.3.2.2 Impulse excitation 398
3.12.3.2.3 Excitation and accelerometer positions 400
3.12.3.3 Coupling Loss Factor, ηij 400
3.12.3.4 Vibration Reduction Index, Kij 402
References 402
4. Direct sound transmission
4.1
4.2
xiv
Introduction 409
Statistical energy analysis
409
C o n t e n t s
4.3
4.2.1 Subsystem definition 411
4.2.2 Subsystem response 412
4.2.3 General matrix solution 414
4.2.4 Converting energy to sound pressures and velocities 415
4.2.5 Path analysis 416
Airborne sound insulation 418
4.3.1 Solid homogeneous isotropic plates 418
4.3.1.1 Resonant transmission 419
4.3.1.2 Non-resonant transmission (mass law) 421
4.3.1.2.1 Infinite plate theory 422
4.3.1.2.2 Finite plate theory 424
4.3.1.3 Examples 427
4.3.1.3.1 Glass 427
4.3.1.3.2 Plasterboard 429
4.3.1.3.3 Masonry wall (A) 429
4.3.1.3.4 Masonry wall (B) 431
4.3.1.3.5 Masonry wall (C) 432
4.3.1.4 Thin/thick plates and thickness resonances 433
4.3.1.5 Infinite plates 435
4.3.1.6 Closely connected plates 440
4.3.2 Orthotropic plates 442
4.3.2.1 Infinite plate theory 443
4.3.2.2 Masonry/concrete plates 446
4.3.2.3 Masonry/concrete plates containing hollows 448
4.3.2.4 Profiled plates 449
4.3.3 Low-frequency range 451
4.3.4 Membranes 454
4.3.5 Plate–cavity–plate systems 454
4.3.5.1 Mass–spring–mass resonance 457
4.3.5.1.1 Helmholtz resonators 460
4.3.5.2 Using the five-subsystem SEA model 461
4.3.5.2.1 Windows: secondary glazing 462
4.3.5.2.2 Masonry cavity wall 467
4.3.5.2.3 Timber joist floor 469
4.3.5.3 Sound transmission into and out of cavities 471
4.3.5.4 Structural coupling 473
4.3.5.4.1 Point connections between plates and/or beams 473
4.3.5.4.2 Line connections 475
4.3.5.4.3 Masonry/concrete walls: foundations 476
4.3.5.4.4 Lightweight cavity walls 478
4.3.5.5 Plate–cavity–plate–cavity–plate systems 479
4.3.6 Sandwich panels 480
4.3.7 Composite sound reduction index for several elements 482
4.3.8 Surface finishes and linings 483
4.3.8.1 Bonded surface finishes 483
4.3.8.2 Linings 485
4.3.9 Porous materials (non-resonant transmission) 488
4.3.9.1 Fibrous sheet materials 490
4.3.9.2 Porous plates 491
4.3.9.3 Coupling loss factor 493
4.3.10 Air paths through gaps, holes, and slits (non-resonant transmission) 493
4.3.10.1 Slit-shaped apertures (straight-edged) 494
xv
C o n t e n t s
4.4
4.5
4.3.10.2 Circular aperture 497
4.3.10.3 More complex air paths 499
4.3.10.4 Using the transmission coefficients 499
4.3.11 Ventilators and HVAC 502
4.3.12 Windows 502
4.3.12.1 Single pane 502
4.3.12.2 Laminated glass 503
4.3.12.3 Insulating glass unit (IGU) 504
4.3.12.4 Secondary/multiple glazing 506
4.3.13 Doors 507
4.3.14 Empirical mass laws 507
Impact sound insulation 509
4.4.1 Heavyweight base floors 509
4.4.2 Lightweight base floors 512
4.4.3 Soft floor coverings 513
4.4.3.1 Heavyweight base floors 513
4.4.3.2 Lightweight base floors 515
4.4.4 Floating floors 516
4.4.4.1 Heavyweight base floors 517
4.4.4.1.1 Resilient material as point connections 517
4.4.4.1.2 Resilient material over entire surface 520
4.4.4.1.3 Resilient material along lines 523
4.4.4.2 Lightweight base floors 524
Rain noise 526
References 527
5. Combining direct and flanking transmission
5.1
5.2
xvi
Introduction 535
Vibration transmission across plate junctions 536
5.2.1 Wave approach: bending waves only 538
5.2.1.1 Angular averaging 539
5.2.1.2 Angles of incidence and transmission 541
5.2.1.3 Rigid X, T, L, and in-line junctions 541
5.2.1.3.1 Junctions of beams 543
5.2.2 Wave approach: bending and in-plane waves 543
5.2.2.1 Bending waves 545
5.2.2.1.1 Incident bending wave at the junction 546
5.2.2.1.2 Transmitted bending wave at the junction 546
5.2.2.2 In-plane waves 547
5.2.2.2.1 Incident quasi-longitudinal wave at the junction 547
5.2.2.2.2 Incident transverse shear wave at the junction 548
5.2.2.2.3 Transmitted in-plane waves at the junction 548
5.2.2.2.4 Conditions at the junction beam 549
5.2.2.2.5 Transmission coefficients 550
5.2.2.2.6 Application to SEA models 550
5.2.2.3 Example: Comparison of wave approaches 552
5.2.2.4 Other plate junctions modelled using a wave approach 556
5.2.2.4.1 Junctions of angled plates 556
5.2.2.4.2 Resilient junctions 557
5.2.2.4.3 Junctions at beams/columns 559
5.2.2.4.4 Hinged junctions 559
C o n t e n t s
5.2.3
5.3
5.4
Finite element method 560
5.2.3.1 Introducing uncertainty 563
5.2.3.2 Example: Comparison of FEM with measurements 563
5.2.3.3 Example: Comparison of FEM with SEA (wave approaches)
for isolated junctions 565
5.2.3.4 Example: Statistical distributions of coupling parameters 567
5.2.3.5 Example: Walls with openings (e.g. windows, doors) 572
5.2.3.6 Example: Using FEM, ESEA, and SEA with combinations of
junctions 576
5.2.4 Foundation coupling (Wave approach: bending waves only) 578
Statistical energy analysis 580
5.3.1 Inclusion of measured data 581
5.3.1.1 Airborne sound insulation 581
5.3.1.1.1 Example 582
5.3.1.2 Coupling loss factors 583
5.3.1.3 Total loss factors 583
5.3.2 Models for direct and flanking transmission 584
5.3.2.1 Example: SEA model of adjacent rooms 585
5.3.2.2 Example: Comparison of SEA with measurements 588
SEA-based model 592
5.4.1 Airborne sound insulation 592
5.4.1.1 Generalizing the model for in situ 595
5.4.2 Impact sound insulation 597
5.4.3 Application 597
5.4.4 Example: Flanking transmission past non-homogeneous separating
walls or floors 600
References 603
Appendix: Material properties
Standards
Index
607
613
619
xvii
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Acknowledgements
I would like to thank Les Fothergill (Office of the Deputy Prime Minister, ODPM), Martin Wyatt
(Building Research Establishment Ltd, BRE), and John Burdett (BRE Trust) for making the
arrangements that allowed me to write this book. I’m very grateful to Les Fothergill (ODPM),
Richard Daniels (Department for Education and Skills, DfES), and John Seller (BRE) who
provided me with many opportunities to work in research, consultancy, regulations, and standardization over the years. I have also been fortunate enough to have met experts working in
the field of building acoustics from all around the world; many of their insights into the subject
are referenced within the pages of this book. I extend my thanks to Yiu Wai Lam, Werner Scholl,
Ole-Herman Bjor, and Michael Vorländer who made time available to read and comment on
various sections.
xix
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List of symbols
a
c
c0
cg
cg(B)
cg(L)
cg(T)
cm
cpm
cB
cD
cL
cT
d
dmfp
e
f
fc
fco
fd
fms
fmsm
fS
fl and fu
fp , fp,q , fp,q,r
fB(thin)
fL(thin)
g
h
h(t)
i
k
kpm
kB
m
n(f )
n2D (f )
n3D (f )
n(ω)
p
r
Acceleration (m/s2 ), equivalent absorption length (m)
Phase velocity (m/s)
Phase velocity of sound in air (m/s)
Group velocity (m/s)
Group velocity for bending waves (m/s)
Group velocity for quasi-longitudinal waves (m/s)
Group velocity for torsional waves on beams, or transverse shear waves on
plates (m/s)
Phase velocity for a membrane (m/s)
Complex phase velocity for sound in a porous material
Bending phase velocity (m/s)
Dilatational wave phase velocity (m/s)
Quasi-longitudinal phase velocity (m/s)
Torsional phase velocity (m/s), transverse shear phase velocity (m/s)
Distance (m), thickness of porous material (m)
Mean free path (m)
Normalized error
Frequency (Hz)
Critical frequency (Hz)
Cut-off frequency (Hz)
Dilatational resonance frequency (Hz)
Mass–spring resonance frequency (Hz)
Mass–spring–mass resonance frequency (Hz)
Schroeder cut-off frequency (Hz)
Lower and upper limits of a frequency band (Hz)
Mode frequency/eigenfrequency (Hz)
Thin plate limit for bending waves (Hz)
Thin plate limit for quasi-longitudinal waves (Hz)
Acceleration due to gravity (m/s2 )
Plate thickness (m), height (m)
impulse response
√
−1
Wavenumber (radians/m), spring stiffness (N/m or N/m3 )
Complex wavenumber for an equivalent gas
Bending wavenumber (radians/m)
Mass (kg), attenuation coefficient in air (Neper/m)
Modal density (modes per Hz)
Modal density for a cavity – two-dimensional space (modes per Hz)
Modal density for a cavity – three-dimensional space (modes per Hz)
Modal density (modes per radians/s)
Sound pressure (Pa)
Radius (m), airflow resistivity (Pa.s/m2 )
xxi
L i s t
o f
s y m b o l s
rrd
s
s′
sa′ , sg′
st′
′
ssoil
sX mm
t
u
v
w
Reverberation distance (m)
Sample standard deviation (–)
Dynamic stiffness per unit area for an installed resilient material (N/m3 )
Dynamic stiffness per unit area for enclosed air(a) or other gas(g) (N/m3 )
Apparent dynamic stiffness per unit area for a resilient material (N/m3 )
Compression stiffness per unit area for soil (N/m3 )
Dynamic stiffness for a wall tie in a cavity of width, X mm (N/m)
Time (s)
Sound particle velocity (m/s)
Velocity (m/s)
Energy density (J/m3 )
A
AT
B
Bb
Bp
Bp,eff
CW
D
DI,n
DI,n,e
Dn
Dn,e
Dn,f
DnT
Dv,ij
Absorption area (m2 )
Total absorption area (m2 )
Bandwidth (Hz), filter bandwidth (Hz)
Bending stiffness for a beam (Nm2 )
Bending stiffness per unit width for a plate (Nm)
Effective bending stiffness per unit width for an orthotropic plate (Nm)
Waterhouse correction (dB)
Sound pressure level difference (dB), damping factor
Intensity normalized level difference (dB)
Intensity element-normalized level difference (dB)
Normalized level difference (dB)
Element-normalized level difference (dB)
Normalized flanking level difference (dB)
Standardized level difference (dB)
Velocity level difference between source element, i, and receiving element,
j (dB)
Direction-averaged velocity level difference (dB)
Energy (J), Young’s modulus (N/m2 )
Force (N), shear force (N)
Surface pressure-intensity indicator (dB)
Shear modulus (N/m2 )
Shear stiffness per unit area for soil (N/m3 )
Sound intensity (W/m2 ), structural intensity (W/m), moment of inertia of the
cross-sectional-area about the y- or z-axis (m4 )
Normal intensity component (W/m2 )
Intensity in the x-, y-, and z-directions
Polar moment of inertia about the longitudinal axis of a beam (m4 )
Torsional moment of rigidity for a beam (m4 )
Stiffness (N/m), contact stiffness (N/m), bulk compression modulus of a gas
(Pa)
Vibration reduction index (dB)
Dynamic capability index (dB)
Junction length between elements i and j (m)
Temporal and spatial average sound intensity level over the measurement
surface (dB)
Normalized impact sound pressure level (dB)
Dv,ij
E
F
FpI
G
Gsoil
I
In
Ix , Iy , Iz
Iθ
J
K
Kij
Ld
Lij
LIn
Ln
xxii
L i s t
Ln,f
L′n
L′nT
Lp (t)
Lp
Lp,s
Lp,A
Lx , Ly , Lz
LT
LW
M
Mav
N(k)
N
Ns
P
P0
Q
R
R0
R′
RI
RI′
Rij
Rs
′
′
′
Rtr,s
, Rrt,s
, Rat,s
S
SM
ST
T
Tint
Ts
TX
U
V
W
Ydp
Z0
Z0,pm
Za,n
o f
s y m b o l s
Normalized flanking impact sound pressure level (dB)
Normalized impact sound pressure level – field measurement (dB)
Standardized impact sound pressure level (dB)
Instantaneous sound pressure level in a space (dB)
Temporal and spatial average sound pressure level in a space (dB)
Temporal and spatial average sound pressure level next to a surface (dB)
A-weighted sound pressure level (dB)
x-, y-, z- dimensions (m)
Total length of all room edges (m)
Temporal and spatial average sound power level radiated by a surface (dB)
Moment (Nm), moment per unit width (N), modal overlap factor (–), molar
mass (kg/mol)
Geometric mean of the modal overlap factors for subsystems i and j (–)
Number of modes below wavenumber, k (–)
Mode count in a frequency band (–), number of reflections, positions,
samples, etc.
Statistical mode count in a frequency band (–)
Static pressure (Pa)
Static pressure for air at atmospheric pressure (Pa)
Shear force per unit width (N)
Reflection coefficient (–), sound reduction index (dB), universal gas constant
(J/mol.K), airflow resistance (Pa.s/m3 ), auto-correlation functions, damping
constant (–)
Normal incidence sound reduction index (dB)
Apparent sound reduction index (dB)
Intensity sound reduction index (dB)
Apparent intensity sound reduction index (dB)
Flanking sound reduction index (dB)
Specific airflow resistance (Pa.s/m)
Apparent sound reduction index for road (tr), railway (rt), and aircraft (at)
traffic (dB)
Area (m2 )
Area of the measurement surface (m2 )
Total area of the room surfaces (m2 )
Period (s), averaging time (s), reverberation time (s), temperature (◦ C), torsional stiffness (Nm2 ), tension per unit length around the edge of a membrane
(N/m)
Integration time (s)
Structural reverberation time (s)
Reverberation time determined from linear regression over a range of
X dB (s)
Perimeter (m)
Volume (m3 )
Power (W)
Driving-point mobility (m/Ns)
Characteristic impedance of air (Pa.s/m)
Characteristic impedance for an equivalent gas (Pa.s/m)
Normal acoustic surface impedance (Pa.s/m)
xxiii
L i s t
o f
Za,s
Zdp
Zp
Specific acoustic impedance (–)
Driving-point impedance (Ns/m)
Surface impedance of a plate (Pa.s/m)
α0
αθ
αst
αs
βa,s
δ(t)
δf
δpI0
ε
ε2
φ
γ
η
ηint
ηii
ηij
ηi
κ
κeff
λ
μ
ν
θ
ρ
ρbulk
ρeff
ρ0
ρs
ρl
σ
τ
ω
ξ
ψ
ζ
ζcdr
f
f3dB
L
Normal incidence sound absorption coefficient (–)
Angle-dependent sound absorption coefficient (–)
Statistical sound absorption coefficient (–)
Sound absorption coefficient (–)
Specific acoustic admittance (–)
Dirac delta function
Average frequency spacing between modes (Hz)
Pressure-residual intensity index (dB)
Strain (–), normalized standard deviation (–), absolute error
Normalized variance (–)
Porosity (–)
Shear strain (–), ratio of specific heats (–)
Displacement (m), loss factor (–)
Internal loss factor (–)
Internal loss factor for subsystem i (–)
Coupling loss factor from subsystem i to subsystem j (–)
Total loss factor for subsystem i (–)
Gas compressibility (Pa−1 )
Effective gas compressibility for an equivalent gas (Pa−1 )
Wavelength (m), Lamé constant
Mean (–), Lamé constant
Poisson’s ratio (–)
Angular torsional displacement (radians)
Density (kg/m3 )
Bulk density (kg/m3 )
Effective gas density for an equivalent gas (kg/m3 )
Density of air (kg/m3 )
Mass per unit area/surface density (kg/m2 )
Mass per unit length (kg/m)
Standard deviation (–), stress (N/m2 ), radiation efficiency (–)
Transmission coefficient (–), shear stress (N/m2 ), time constant (s)
Angular frequency ω = 2πf (radians/s), angular velocity (radians/s)
Displacement (m)
Mode shape/eigenfunction
Displacement (m)
Constant damping ratio (–)
Frequency spacing between eigenfrequencies (Hz), frequency bandwidth (Hz)
3 dB bandwidth (or half-power bandwidth) (Hz)
Decrease in sound pressure level along a corridor (dB), improvement of impact
sound insulation (dB)
Sound reduction improvement index (dB)
Resonant sound reduction improvement index (dB)
Potential function (m2 )
Complex propagation constant
R
RResonant
Ŵ
xxiv
s y m b o l s
L i s t
<>
< >t
< >s
< >t,s
< >f
X̂
o f
s y m b o l s
Stream function (m2 )
Mean value
Temporal average
Spatial average
Temporal and spatial average
Frequency average
Denotes peak value of variable X
xxv
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Chapter 1
Sound fields
1.1 Introduction
T
his opening chapter looks at aspects of sound fields that are particularly relevant
to sound insulation; the reader will also find that it has general applications to room
acoustics.
The audible frequency range for human hearing is typically 20 to 20 000 Hz, but we generally
consider the building acoustics frequency range to be defined by one-third-octave-bands from
50 to 5000 Hz. Airborne sound insulation tends to be lowest in the low-frequency range and
highest in the high-frequency range. Hence significant transmission of airborne sound above
5000 Hz is not usually an issue. However, low-frequency airborne sound insulation is of particular importance because domestic audio equipment is often capable of generating high levels
below 100 Hz. In addition, there are issues with low-frequency impact sound insulation from
footsteps and other impacts on floors. Low frequencies are also relevant to façade sound insulation because road traffic is often the dominant external noise source in the urban environment.
Despite the importance of sound insulation in the low-frequency range it is harder to achieve
the desired measurement repeatability and reproducibility. In addition, the statistical assumptions used in some measurements and prediction models are no longer valid. There are some
situations such as in recording studios or industrial buildings where it is necessary to consider
frequencies below 50 Hz and/or above 5000 Hz. In most cases it should be clear from the text
what will need to be considered at frequencies outside the building acoustics frequency range.
1.2 Rooms
Sound fields in rooms are of primary importance in the study of sound insulation. This section
starts with the basic principles needed to discuss the more detailed aspects of sound fields that
are relevant to measurement and prediction. In the laboratory there is some degree of control
over the sound field in rooms due to the validation procedures that are used to commission
them. Hence for at least part of the building acoustics frequency range, the sound field in
laboratories can often be considered as a diffuse sound field; a very useful idealized model.
Outside of the laboratory there are a wide variety of rooms with different sound fields. These
can usually be interpreted with reference to two idealized models: the modal sound field and
the diffuse sound field.
1.2.1 Sound in air
Sound in air can be described as compressional in character due to the compressions and
rarefactions that the air undergoes during wave propagation (see Fig. 1.1). Air particles move to
and fro in the direction of propagation, hence sound waves are referred to as longitudinal waves.
The compressions and rarefactions cause temporal variation of the air density compared to the
1
S o u n d
I n s u l a t i o n
Propagation direction
Wavelength, λ
Figure 1.1
Longitudinal wave – compression and rarefaction of air particles.
density at equilibrium. The result is temporal variation of the air pressure compared to the static
air pressure. Sound pressure is therefore defined by the difference between the instantaneous
pressure and the static pressure.
The phase velocity for sound in air, c0 (or as it is more commonly referred to, the speed of
sound) is dependent upon the temperature, T , in ◦ C and for most practical purposes can be
calculated using
c0 = 331 + 0.6T
(1.1)
for temperatures between 15◦ C and 30◦ C and at atmospheric pressure.
The density of air at equilibrium, ρ0 , is also temperature dependent and can be calculated from
ρ0 =
353.2
273 + T
(1.2)
For calculations in buildings it is often assumed that the temperature is 20◦ C, for which the
speed of sound is 343 m/s and the density of air is 1.21 kg/m3 . This will be assumed throughout
the book.
The fundamental relationship between the phase velocity, the frequency, f , and the wavelength, λ, is
c0 = f λ
(1.3)
The wavelength is the distance from peak to peak (or trough to trough) of a sinusoidal wave;
this equals the distance between identical points of compression (see Fig. 1.1) or rarefaction.
For the building acoustics frequency range, the wavelength in air at 20◦ C is shown in Fig. 1.2.
If we consider these wavelengths relative to typical room dimensions, it is clear that we are
dealing with a very wide range. For this reason it is useful to describe various aspects of sound
fields by referring to low-, mid-, and high-frequency ranges; corresponding to 50 to 200 Hz,
250 to 1000 Hz, and 1250 to 5000 Hz respectively.
2
Chapter 1
7
6
Wavelength, λ (m)
5
4
3
2
1
0
50
80
125
200
315
500
800
1250 2000
One-third-octave-band centre frequency (Hz)
3150
5000
Figure 1.2
Wavelength of sound in air at 20◦ C.
As we need to describe the spatial variation of sound pressure as well as the temporal variation,
it is necessary to use the wavenumber, k, which is defined as
k=
2π
ω
=
c0
λ
(1.4)
where the angular frequency, ω, is
ω = 2πf
(1.5)
and the period, T , of the wave is
T =
1
2π
=
ω
f
(1.6)
The wavenumber is useful for discussing aspects relating to spatial variation in both sound and
vibration fields in terms of kd, where d is the distance between two points.
Two types of wave need to be considered both inside and outside of buildings: plane waves
and spherical waves. Before reviewing these waves we will briefly review the use of complex
notation that simplifies many derivations for sound and structure-borne sound waves.
1.2.1.1
Complex notation
For both sound and vibration, it is useful to look at wave motion or signals at single frequencies; these are defined using harmonic sine and cosine functions, e.g. p(x, t) = cos (ωt − kx).
It is usually more convenient to describe these simple harmonic waves using complex
exponential notation, where
exp(iX ) = cos X + i sin X
(1.7)
Equations using complex notation are often easier to manipulate than sines and cosines, and
can be written in a more compact form. A brief review of complex notation is given here as this
is covered in general acoustic textbooks (e.g. see Fahy, 2001).
3
S o u n d
I n s u l a t i o n
The most commonly used complex exponentials are those that describe the temporal and
spatial variation of harmonic waves. For the convention used in this book, these are
exp(iωt) = cos(ωt) + i sin(ωt)
(1.8)
and
exp(−kx) = cos(kx) − i sin(kx)
Complex notation also simplifies differentiation and integration. For example, differentiation
or integration with respect to time becomes equivalent to multiplication or division by iω
respectively.
Whilst it can be convenient to work with complex notation, the final result that corresponds
to a physical quantity (sound pressure, velocity, etc.) must be real, rather than imaginary. In
general, it is the real part of the solution that represents the physical quantity.
The time-average of harmonic waves is frequently needed for practical purposes and is
denoted by t . The following time-averages often occur in derivations,
1 T
1 T
(1.9)
cos2 (ωt)dt = lim
sin2 (ωt)dt = 0.5
lim
T →∞ T 0
T →∞ T 0
and
1
T →∞ T
lim
0
T
1
T →∞ T
sin(ωt)dt = lim
0
T
1
T →∞ T
cos(ωt)dt = lim
0
T
sin(ωt) cos(ωt)dt = 0
(1.10)
where T is the averaging time.
The time-average of the product of two waves, p1 (t) and p2 (t), that are written in complex
exponential notation can be calculated using
1
(1.11)
p1 p2 t = Re p1 p2∗
2
where ∗ denotes the complex conjugate.
1.2.1.2 Plane waves
To gain an insight into the sound field in rooms we often assume that it is comprised of plane
waves; so called, because in any plane that is perpendicular to the propagation direction, the
sound pressure and the particle velocity are uniform with constant phase. These planes are
referred to as wavefronts. In practice, plane waves can be realized (approximately) in a long
hollow cylinder which has rigid walls. A sound source is placed at one end of the cylinder that
generates sound with a wavelength that is larger than the diameter of the cylinder. This results
in a plane wave propagating in the direction away from the source. The longitudinal wave
shown in Fig. 1.1 can also be seen as representing a plane wave in this cylinder. This onedimensional scenario may seem somewhat removed from real sound fields in typical rooms.
However, the plane wave model can often be used to provide a perfectly adequate description
of the complex sound fields that are encountered in practice.
Using a Cartesian coordinate system, the wave equation that governs the propagation of
sound through three-dimensional space is
∂2 p
∂2 p
∂2 p
1 ∂2 p
+ 2 + 2 − 2 2 =0
2
∂x
∂y
∂z
c0 ∂t
where p is the sound pressure.
4
(1.12)
Chapter 1
For a plane wave that is propagating in the positive x, y, and z-direction across this space, the
sound pressure is described by an equation of the form
p(x, y, z, t) = p̂ exp(−ikx x) exp(−iky y) exp(−ikz z) exp(iωt)
(1.13)
where p̂ is an arbitrary constant for the peak value, and kx , ky , and kz are constants relating to
the wavenumber.
As we are using harmonic time dependence, exp(iωt), the wave equation can now be written
in terms of the wavenumber as
∂2 p
∂2 p
∂2 p
+ 2 + 2 + k 2p = 0
2
∂x
∂y
∂z
(1.14)
The relationship between the wavenumber and the constants, kx , ky , and kz is found by
inserting Eq. 1.13 into the wave equation, which gives
k 2 = kx2 + ky2 + kz2
(1.15)
We will need to make use of this relationship to describe the sound field in rooms. However,
the wave equation only governs sound propagation across three-dimensional space. It does
not describe the sound field in a room because it does not take account of the waves impinging
upon the room surfaces. In other words, this equation does not take account of boundary
conditions. Hence the constants, kx , ky , and kz can only be determined once we have defined
these boundary conditions.
The particle motion gives rise to sound pressure; hence we can relate the sound particle
velocities, ux , uy , and uz (in the x, y, and z directions respectively) to the sound pressure by
using the following equations of motion,
∂uy
∂p
= −ρ0
∂y
∂t
∂ux
∂p
= −ρ0
∂x
∂t
∂p
∂uz
= −ρ0
∂z
∂t
(1.16)
kz
p
ωρ0
(1.17)
and therefore the particle velocities are
ux =
kx
p
ωρ0
uy =
ky
p
ωρ0
uz =
The ratio of the complex sound pressure to the complex sound particle velocity at a single
point is the specific acoustic impedance, Za . For a plane wave propagating in a single direction
(we will choose the x-direction, so that k = kx and u = ux ), this impedance is referred to as the
characteristic impedance of air, Z0 , and is defined as
ρ0
p
(1.18)
Z0 = = ρ0 c0 =
u
κ
where κ is the gas compressibility (adiabatic).
The particle velocity is related to the sound pressure by a real constant that is independent
of frequency. Therefore, the sound pressure and the particle velocity always have the same
phase on the plane that lies perpendicular to the direction of propagation.
In order to predict or measure sound transmission we will need to quantify the sound intensity,
I; the energy that flows through unit surface area in unit time. The sound intensity is the
time-averaged value of the product of sound pressure and particle velocity,
I = put =
p2 t
ρ0 c0
(1.19)
5
S o u n d
I n s u l a t i o n
where p2 t is the temporal average mean-square sound pressure given by
1 T 2
p2 t =
(1.20)
p dt
T 0
and T is the averaging time. Note that p2 t is described as the root-mean-square (rms)
sound pressure.
1.2.1.3
Spherical waves
For spherical waves the sound pressure and the particle velocity over a spherical surface
are uniform with constant phase; these surfaces are referred to as wavefronts (see Fig. 1.3).
For a sound source such as a loudspeaker used in sound insulation measurements, a useful
idealized model is to treat the loudspeaker as a point source that generates spherical waves.
A point source is one for which the physical dimensions are much smaller than the wavelength
of the sound, and the sound radiation is omnidirectional.
We now need to make use of a spherical coordinate system defined by a distance, r , from
the origin at r = 0. For spherically symmetrical waves, the wave equation that governs the
propagation of sound through three-dimensional space is
∂2 p
2 ∂p
1 ∂2 p
+
− 2 2 =0
2
∂r
r ∂r
c0 ∂t
(1.21)
For a spherical wave propagating across this space, the sound pressure can be described by
an equation of the form
p̂
p(r , t) = exp(−ikr ) exp(iωt)
(1.22)
r
λ
Figure 1.3
Spherical wavefronts produced by a point source.
6
Chapter 1
where p̂ is an arbitrary constant for the peak value and r is the distance from a spherical
wavefront to the origin.
Substitution of Eq. 1.22 into the following equation of motion,
∂p
∂ur
= −ρ0
∂r
∂t
gives the radial particle velocity, ur , as
p
p
i
c0
ur =
=
1−
1−i
ρ0 c 0
kr
ρ0 c0
2πfr
(1.23)
(1.24)
This gives the acoustic impedance, Za , as
Za =
ρ 0 c0
p
=
ur
1 − kri
=
ρ0 c0 k 2 r 2
ρ0 c0 kr
+i
1 + k 2r 2
1 + k 2r 2
(1.25)
In contrast to plane waves, the particle velocity is related to the sound pressure by a complex
variable that is dependent on both the wavenumber and distance. So although the phase of
the sound pressure and the phase of the particle velocity are constant over a spherical surface
at a specific frequency, they do not have the same phase over this surface.
The time-averaged sound intensity for a harmonic spherical wave is
1
I=
2ρ0 c0
2
p2 t
p̂
=
r
ρ0 c0
(1.26)
For spherical waves, the intensity is seen to be proportional to 1/r 2 ; this feature is often
referred to as spherical divergence. The sound power associated with a point source producing
spherical waves can now be calculated from the intensity using
W = 4πr 2 I =
2πp̂2
ρ0 c0
(1.27)
Rather than use p̂ in Eqs 1.26 and 1.27, a point source can be described using a peak volume
velocity, Q̂, given by
Q̂ =
4πp̂
ωρ0
(1.28)
When kr ≫ 1 (i.e. at high frequencies and/or large distances) the imaginary part of Za is small,
therefore the particle velocity has almost the same phase as the sound pressure and Za tends
towards Z0 . The time-averaged sound intensity for the harmonic spherical wave then tends
towards the value for a plane wave (Eq. 1.19). These links between plane waves and spherical waves indicate why we are able to use the simpler plane wave model in many of the
derivations involved in sound insulation. Any errors incurred through the assumption of plane
waves are often negligible or insignificant compared to those that are accumulated from other
assumptions.
1.2.1.4
Acoustic surface impedance and admittance
As rooms are formed by the surfaces at the boundaries of the space we need to know the
acoustic impedance of a room surface as seen by an impinging sound wave. The normal
7
S o u n d
I n s u l a t i o n
acoustic surface impedance, Za,n , is defined as the ratio of the complex sound pressure at a
surface, to the component of the complex sound particle velocity that is normal to this surface,
Za,n =
p
un
(1.29)
Although we are mainly interested in plates (representing walls or floors) that form the room
boundaries, the above definition applies to any surface, including sheets of porous materials
such as mineral wool or foam.
The specific acoustic impedance, Za,s , is defined using the characteristic impedance of air,
Za,s =
Za,n
ρ0 c0
(1.30)
In some calculations it is more appropriate or convenient to use the specific acoustic
admittance, βa,s , rather than the specific acoustic impedance, where
βa,s =
1
Za,s
(1.31)
When calculating sound fields in rooms it is often convenient to assume that the room surfaces
are rigid. This is reasonable for many hard surfaces in buildings. At a rigid surface, the particle
velocity that is normal to this surface is zero; hence Za,n and Za,s become infinitely large and
βa,s is taken to be zero.
1.2.1.5 Decibels and reference quantities
The human ear can detect a wide range of sound intensities. The decibel scale (dB) is commonly used to deal with the wide range in pressure, intensity, power, and energy that are
encountered in acoustics. Levels in decibels are defined using the preferred SI reference
quantities for acoustics in Table 1.1 (ISO 1683); these reference quantities are used for all
figures in the book.
Table 1.1. Sound – definitions of levels in decibels
Level
Sound pressure
Energy
Intensity
Sound power
Loss factors (Internal,
Coupling, Total)
8
Definition
p
Lp = 20 lg
p0
where p is the rms pressure
E
LE = 10 lg
E0
I
LI = 10 lg
I0
W
LW = 10 lg
W0
η
LILF /LCLF /LTLF = 10 lg
η0
Reference quantity
p0 = 20 × 10−6 Pa
NB only for sound in air
E0 = 10−12 J
I0 = 10−12 W/m2
W0 = 10−12 W
η0 = 10−12
Chapter 1
A-weighting
1.2.1.6
A-weighting is used to combine sound pressure levels from a range of frequencies into a
single value. This is the A-weighted sound pressure level, Lp,A . It is intended to represent
the frequency response of human hearing and is often used to try and make a simple link
between the objective and subjective assessment of a sound. A-weighting accounts for the
fact that with the same sound pressure level, we do not perceive all frequencies as being
equally loud. In terms of the building acoustics frequency range it weights the low-frequency
range as being less significant than the mid- and high-frequency range. This does not mean
that the low-frequency range is unimportant for sound insulation, usually quite the opposite is
true; the A-weighted level depends upon the spectrum of the sound pressure level. Although
it is common to measure and predict sound insulation in frequency bands, assessment of the
sound pressure level in the receiving room is often made in terms of the A-weighted level.
For N frequency bands, the sound pressure level Lp (n) in each frequency band, n, is combined
to give an A-weighted level using
Lp,A = 10 lg
N
10
(Lp (n)+A(n))/10
n=1
(1.32)
where the A-weighting values, A(n), are shown in Fig. 1.4 for one-third-octave-bands (IEC
61672-1).
For regulatory and practical purposes, the airborne sound insulation is often described using
a single-number quantity that corresponds to the difference between the A-weighted level in
the source room and the A-weighted level in the receiving room for a specific sound spectrum
(e.g. pink noise) in the source room (ISO 717 Part 1). Use of this A-weighted level difference
10
A-weighting, A(n) (dB)
0
⫺10
⫺20
⫺30
⫺40
Low-frequency
range
⫺50
Mid-frequency
range
High-frequency
range
⫺60
20
40
80
160
315
630
1250 2500 5000
One-third-octave-band centre frequency (Hz)
10 000 20 000
Figure 1.4
A-weighting values over the range of human hearing indicating the low-, mid-, and high-frequency ranges for the building
acoustics frequency range.
9
S o u n d
I n s u l a t i o n
Time domain
Frequency domain
t
Amplitude
Amplitude
Sine wave
f
t
Amplitude
Amplitude
Impulse
f
Figure 1.5
Illustration of a sine wave and an impulse in the time and frequency domains.
simplifies calculation of the A-weighted level in the receiving room. It can also be used to make
a link to subjective annoyance (Vian et al., 1983).
1.2.2
Impulse response
In sound insulation as well as in room acoustics, we need to make use of the impulse response
in both measurement and theory. The frequency spectrum of an impulse is flat. It therefore
contains energy at all frequencies, whereas a sine wave only has energy at a single frequency
(see Fig. 1.5).
The general principle for an impulse response applies to any acoustic system, whether it is
sound pressure in a room or a cavity, or the vibration of a plate or a beam. It is based upon the
response of an acoustic system to a Dirac delta function, δ(t), sometimes called a unit impulse.
The delta function is infinite at t = 0 and infinitely narrow, such that δ(t) = 0 when t = 0, and it
has the property
∞
(1.33)
δ(t)dt = 1
−∞
Excitation of a linear time-invariant (LTI) acoustic system with a delta function results in the
impulse response of the system, h(t). The delta function is important because any kind of signal
can be described by using a train of impulses that have been appropriately scaled and shifted
in time. Hence, the impulse response completely describes the response of an LTI system to
any input signal, x(t). The output signal, y(t), can then be found from the convolution integral
y(t) =
∞
−∞
h(u)x(t − u)du =
where u is a dummy time variable.
10
∞
−∞
h(t − u)x(u)du
(1.34)
Impulse response, h (t) (linear)
Chapter 1
Time, t (s)
Figure 1.6
Example of a measured impulse response in a room.
For brevity, convolution is often written as y(t) = x(t) ∗ h(t). The convolution integral uses the
dummy time variable to multiply the time-reversed input signal by the impulse response (or the
time-reversed impulse response by the input signal), and integrate over all possible values of
t to give the output signal.
An example of a measured impulse response for sound pressure in a room is shown in Fig. 1.6.
1.2.3
Diffuse field
One of the assumptions commonly made in the measurement and prediction of sound insulation
is that the sound field in rooms can be considered as being diffuse. A diffuse sound field can be
considered as one in which the sound energy density is uniform throughout the space (i.e. the
sound field can be considered to be homogeneous), and, if we choose any point in the space,
sound waves arriving at this point will have random phase, and there will be equal probability of
a sound wave arriving from any direction. The diffuse field is a concept; in practice there must
be dissipation of energy, so there cannot be equal energy flow in all directions, there must be
net energy flow from a sound source towards part(s) of the space where sound is absorbed.
In diffuse fields it is common to refer to diffuse reflections; this means that the relationship
between the angle of incidence and the angle of reflection is random. This is in contrast to
specular reflection, where the angle of incidence equals the angle of reflection. Walls and
floors commonly found in buildings (excluding spaces specially designed for music performance
such as studios or concert halls) tend to be flat and smooth, from which one might assume that
specular reflections were the norm, and that diffuse reflections were the exception. However,
walls commonly have objects placed near them that partially obscure the wall from the incident
sound wave, such as tables, chairs, bookcases, filing cabinets, and cupboards. These can
11
S o u n d
I n s u l a t i o n
cause the incident wave to be scattered in non-specular directions. Non-specular reflection
also occurs when the acoustic impedance varies across the surface; for example a wall where
the majority of the surface area is concrete but with areas of glazing, wooden doors, or recessed
cupboards, each of which have different impedances. Hence there will usually be a degree
of non-specular reflection, such that some of the incident energy is specularly reflected and
some is diffusely reflected.
The diffuse field is a very useful concept. It allows many simplifications to be made in the measurement and prediction of sound insulation, as well as in other room acoustics calculations.
These make use of the mean free path that will be defined in the following section. In the
laboratory we can create close approximations to a diffuse field in the central zone of a room.
However, the sound field does not always bare a close resemblance to a diffuse field over the
entire building acoustics frequency range. In the low-frequency range this is primarily due to
the fact that sound waves arriving at any point come from a limited number of directions. In the
mid- and high-frequency ranges, waves arriving at any point tend to come from many different
directions. In the central zone of typical rooms it is often reasonable to assume that there is
a diffuse field in the mid- and high-frequency ranges. However it is not always appropriate to
assume that there is a diffuse field when: (a) there are regular room shapes without diffusing
elements, (b) there are non-diffuse reflections from room surfaces, and (c) there is non-uniform
distribution of absorption over the room surfaces. For the above reasons, we need to note the
limitations in applying diffuse field theory to the real world.
1.2.3.1 Mean free path
The mean free path, dmfp , is the average distance travelled by a sound wave between two
successive diffuse reflections from the room surfaces. From the basic relationship, c0 = dmfp /t,
we can calculate the time, t, taken to travel this distance. Upon each reflection, a fraction of
the sound energy is absorbed; hence the mean free path allows us to calculate the build-up
or decay of sound energy in a room over time. It will therefore be needed later on when we
derive the reverberation time in diffuse fields as well as when calculating the power incident
upon walls or floors that face into a room with a diffuse sound field. The following derivation is
taken from Kosten (1960) and starts by deriving the mean free path in a two-dimensional space
before extending it to three dimensions. This two-dimensional space has an area, S, and a
perimeter length, U. An arbitrary two-dimensional space can be defined by a closed curve as
shown in Fig. 1.7; note that although the space is defined by curved lines we assume that
all reflections are diffuse. The dashed lines within this curve represent free paths in a single
direction, where each free path has a length, l.
Projective geometry is now used to transform points along the perimeter of the space onto a
projection plane. Each of the free paths lies perpendicular to a projection plane that defines
the apparent length of the surface, La . When the space is uniformly filled with free paths, the
surface area of the space can be written in terms of the free path lengths using
(1.35)
S=
ldLa = La l
La
where l is the mean free path in one direction.
The number of paths in a single direction is proportional to the apparent length, so using a fixed
number of paths per unit of the projection length, and accounting for all N possible directions
12
Chapter 1
dU
θ
Projection
plane
La
Figure 1.7
Two-dimensional space showing some of the free paths (dashed lines) in a single direction that lie perpendicular to the
projection plane. The apparent length, La , is calculated by using a line integral to sum the projection of the small perimeter
length, dU, onto the projection plane.
gives the mean free path, dmfp , as
lim
dmfp =
N→∞
lim
N
N→∞
n=1
N
La,n l n
n=1
La,n
(1.36)
From Eq. 1.35, S = La,n l n in each direction, n, so Eq. 1.36 can be rewritten as
dmfp =
S
La
(1.37)
where La is the average apparent length.
The next step is to determine this average apparent length, but first we just look at a single
direction and calculate the apparent length. This is done by using a line integral for the closed
curve. At each point along the closed curve, the vector in the direction of the curve makes an
angle, θ, with the projection plane (see Fig. 1.7). The projection of each small perimeter length,
dU, onto the projection plane is a positive value, | cos θ|dU. By integrating around the entire
closed curve, the integral is effectively counting each free path twice, so a multiplier of one-half
is needed. The apparent length is therefore given by
1
La =
(1.38)
| cos θ|dU
2 C
To find the average apparent length, it is necessary to average over all possible directions.
This results in an average cosine term,
2
1 π
| cos θ|dθ =
| cos θ| =
(1.39)
π 0
π
13
S o u n d
I n s u l a t i o n
which gives the average apparent length as
La =
1
2
C
| cos θ|dU =
U
π
(1.40)
The mean free path for a two-dimensional space can now be found from Eqs 1.37 and 1.40,
giving
πS
U
dmfp =
(1.41)
We now consider a three-dimensional space with a volume, V , and a total surface area, ST .
Moving to three dimensions means that the projection plane becomes a surface (rather than a
line) onto which small parts of the surface area, dST , are projected (rather than small perimeter
lengths). Hence we need to define an apparent surface area, Sa .
The volume of the space can be written in terms of the free path lengths using
V=
ldSa = Sa l
(1.42)
Sa
where l is the mean free path in one direction.
The number of paths in a single direction is proportional to the apparent surface area, so using
a fixed number of paths per unit area of the projection surface, and accounting for all N possible
directions gives the mean free path, dmfp , as
lim
dmfp =
N→∞
lim
N
n=1
N→∞
N
Sa,n l n
n=1
Sa,n
(1.43)
From Eq. 1.42, V = Sa,n l n in each direction, n, so Eq. 1.43 can be rewritten as
dmfp =
V
Sa
(1.44)
where S a is the average apparent surface area.
The average apparent surface area is found from the surface integral,
Sa =
1
2
S
| cos θ|dST
(1.45)
Averaging over all possible directions gives the average cosine term. For any enclosed volume
with convex surfaces, the average apparent surface area is given by
Sa =
ST
4
(1.46)
The assumption that the volume effectively forms a convex solid does not limit its applicability to
real rooms as long as the surface area associated with any concave surfaces within the volume
are included in the calculation of ST (Kosten, 1960). Note once again that it is assumed that
all of these curved surfaces result in diffuse reflections.
14
Chapter 1
For any shape of room in which all room surfaces diffusely reflect sound waves, the mean free
path for a three-dimensional space is given by Eqs 1.44 and 1.46, hence
dmfp =
4V
ST
(1.47)
where ST is the total area of the room surfaces and V is the room volume.
It is important to note that Eq. 1.47 gives the mean value; as with any random process, there
will be a spread of results. The mean free path applies to any shape of room with diffusely
reflecting surfaces. However, the statistical distribution of the mean free path in rooms with
diffusely reflecting surfaces depends upon the room shape and its dimensions as well as the
presence of scattering objects within the room (Kuttruff, 1979).
1.2.4 Image sources
A geometrical approach to room acoustics allows calculation of the room response using image
sources. It is briefly described here to introduce the concept of image sources for specular
reflections from surfaces. This will be needed in later sections to describe sound fields within
rooms, as well as sound incident upon a building façade from outside.
This approach assumes that the wavelength is small compared with the dimensions of the
surface that the wave hits. In the study of room acoustics in large rooms and/or at high frequencies this allows sound to be considered in terms of rays rather than waves. Using rays
means that diffraction and phase information that causes interference patterns is ignored. In a
similar way to the study of optics, a ray can be followed from a point source to the boundary
where it undergoes specular reflection, such that the angle of incidence equals the angle of
reflection.
Image sources are defined by treating every boundary (e.g. wall, floor, ground) as mirrors in
which the actual source can be reflected (see Fig. 1.8). The length of the propagation path from
source to receiver is then equal to the distance along the straight line from the image source to
the receiver. As we are considering spherical waves from a point source it is necessary to use
this distance to take account of spherical divergence when calculating the intensity (Eq. 1.26).
For certain receiver positions in rooms with shapes that are much more complex than a simple
box, some of the image sources generated by the reflection process will correspond to paths
that cannot physically exist in practice. Hence for rooms other than box-shaped rooms, it is
necessary to check the validity of each image source for each receiver position.
1.2.4.1 Temporal density of reflections
For a box-shaped room containing a single point source, the image source approach that
was described above can be used to create an infinitely large number of image rooms each
containing a single image source. A small portion of this infinite matrix of image rooms in twodimensional space is shown in Fig. 1.9. Assuming that the point source generates an impulse
at t = 0, each image source must also generate an identical impulse at t = 0. This ensures that
all propagation paths have the correct time lag/gain relative to each other. A circle of radius,
c0 t, with its origin in the centre of the source room will therefore enclose image sources (i.e.
propagation paths with reflections) with propagation times less than t. Moving on to consider
15
S o u n d
I n s u l a t i o n
S⬘
S
Source
Image source
Boundary
Figure 1.8
Source and image source.
Figure 1.9
Source ( ) and image sources ( ) for a box-shaped room (dark solid lines) and some of its image rooms (dotted lines).
16
Chapter 1
three-dimensional space, it follows that the volume of a sphere of radius, c0 t, divided by the
volume associated with each image source (i.e. the room volume, V ) will equal the number of
reflections, N, arriving at a point in the room within time, t. Hence,
N=
4π(c0 t)3
3V
(1.48)
and the temporal density of reflections, dN/dt (i.e. the number of reflections arriving per second
at time, t) is
4πc03 t 2
dN
=
dt
V
(1.49)
This equation applies to any shape of room with a diffuse field; the derivation simply uses a
box-shaped room to simplify use of the image source approach.
1.2.5 Local modes
Having looked at the diffuse field, we will now look at the other idealized model, the modal
sound field. We start by defining room modes. To do this we can follow the journey of a plane
wave as it travels around a box-shaped room. To simplify matters we assume that all the
room surfaces are perfectly reflecting and rigid. Therefore the incident and reflected waves
have the same magnitude and the sound pressure is reflected from the surface without any
change in phase. A rigid wall or floor is defined as one which is not caused to vibrate when
a sound wave impinges upon it; hence the particle velocity normal to the surface is zero. In
practice, walls and floors do vibrate because this is the mechanism that is responsible for sound
transmission; however, this assumption avoids having to consider the wide range of acoustic
surface impedances that are associated with real surfaces.
We now follow the path that is travelled by a sound wave as it travels across a box-shaped
room (see Fig. 1.10). At some point in time it will hit one of the room boundaries from which it
will be reflected before continuing on its journey to be reflected from other room boundaries.
z
Ly
Lz
y
Lx
x
Figure 1.10
Box-shaped room.
17
S o u n d
I n s u l a t i o n
Figure 1.11
Room modes. Plan view of a box-shaped room showing one possible journey taken by a plane wave. A room mode occurs
when the wave travels through the same starting point (•) travelling in exactly the same direction as when it first left, whilst
achieving phase closure.
These reflections are assumed to be specular as would occur with smooth walls and floors that
have uniform acoustic surface impedance over their surface. We can also follow the journeys
taken by other sound waves travelling in other directions. Some of these waves will return
to the starting point travelling in exactly the same direction as when they first left. In some
instances the length of their journey, in terms of phase, will correspond to an integer multiple
of 2π such that there will be continuity of phase; we will refer to this as phase closure. Each
journey that returns to the same starting point travelling in the same direction whilst achieving
phase closure defines a mode with a specific frequency (see Fig. 1.11).
The term ‘local mode’ is used because the modes are ‘local’ to a space that is defined by its
boundaries; in a similar way we will define local modes of vibration for structure-borne sound
on plates and beams in Chapter 2. For rooms this definition assumes that there is no interaction
between the sound waves in the room and the structure-borne sound waves on the walls and
floors that face into that room. The walls and floors are only considered as boundaries that
determine the fraction of wave energy that is reflected and the phase change that occurs upon
reflection. It is also assumed that there is no sound source exciting these modes; we have
simply followed the journey of a plane wave without considering how it was generated. Hence
it is important to note that local modes of spaces and structures (e.g. rooms, walls, and floors)
are a concept; they do not actually exist in real buildings where the spaces and structures
are coupled together. Although the definition of local modes is slightly removed from reality,
the concept is very useful in studying certain features of sound or vibration fields, as well as
the interaction between these fields using methods such as Statistical Energy Analysis. Local
modes are also referred to as natural modes or pure standing waves; they are a property of the
space, rather than a combined function of the space and the excitation. The latter is referred
to as a resonance. The term local mode is sometimes abbreviated to mode; only using the full
name where it is necessary to distinguish it from a global mode.
To calculate the frequencies of the room modes in this box-shaped room it is necessary to
calculate the wavenumbers. Hence we refer back to our discussion in Section 1.2.1.2 on
plane waves and the wave equation where the relationship between the wavenumber and the
constants, kx , ky , and kz , was given by Eq. 1.15. These constants are calculated by using the
equation for sound pressure in a plane wave which must satisfy both the wave equation (Eq.
1.14) and the boundary conditions. For a box-shaped room with dimensions Lx , Ly , and Lz ,
the following boundary conditions are required to ensure that the particle velocity normal to the
rigid room surfaces is zero,
∂p
= 0 at x = 0 and x = Lx
∂x
18
∂p
= 0 at y = 0 and y = Ly
∂y
∂p
= 0 at z = 0 and z = Lz .
∂z
(1.50)
Chapter 1
By taking the real part of Eq. 1.13 that describes the sound pressure for a plane wave and
ignoring time dependence we have the following solution
p(x, y, z) = p̂ cos(kx x) cos(ky y) cos(kz z)
(1.51)
This will only satisfy the boundary conditions when sin(kx Lx ) = sin(ky Ly ) = sin(kz Lz ) = 0, hence
kx =
pπ
Lx
ky =
qπ
Ly
kz =
rπ
Lz
(1.52)
where the variables p, q, and r can take zero or positive integer values.
Each combination of values for p, q, and r describes a room mode for which the mode
wavenumber, kp,q,r , (also called an eigenvalue) is found from Eqs 1.15 and 1.52 to be
2 2
q
r
p 2
+
+
kp,q,r = π
(1.53)
Lx
Ly
Lz
Therefore the mode frequency, fp,q,r (also called an eigenfrequency) is
2 2
c0
q
r
p 2
fp,q,r =
+
+
2
Lx
Ly
Lz
(1.54)
where p, q, and r take zero or positive integer values.
In a box-shaped room there are three different types of room mode: axial, tangential, and
oblique modes.
Axial modes describe the situation where wave propagation is parallel to the x, y, or z axis.
They have one non-zero value for p, q, or r , and zero values for the other two variables (e.g.
f1,0,0 , f0,3,0 , f0,0,2 ).
Tangential modes can be described by defining a ‘pair of surfaces’ as two surfaces that lie
opposite each other, where each pair of surfaces partially defines the box-shaped room. Hence,
tangential modes describe wave propagation at an angle that is oblique to two pairs of surfaces,
and is tangential to the other pair of surfaces. They have non-zero values for two of the variables
p, q, or r , and a zero value for the other variable (e.g. f1,2,0 , f3,0,1 , f0,2,2 ).
Oblique modes describe the situation where wave propagation occurs at an angle that is oblique
to all surfaces; hence they have non-zero values for p, q, and r (e.g. f2,3,1 ).
We have assumed that all the room surfaces are perfectly reflecting and rigid, in practice there
is interaction between the sound pressure in the room and the vibration of the walls and floors
facing into that room. However, the assumption of rigid walls and floors is reasonable in many
rooms because this interaction results in relatively minor shifts in the eigenfrequencies.
1.2.5.1 Modal density
It is often useful to calculate the first 10 or so modes to gain an insight into their distribution
between the frequency bands in the low-frequency range. However, in a room of approximately 50 m3 there are almost one-million modes in the building acoustics frequency range.
Fortunately there is no need to calculate all of these modes because we can adopt a statistical
viewpoint. A statistical approach also helps us to deal with the fact that very few rooms are
19
S o u n d
I n s u l a t i o n
kz
π/Ly
k1,2,3
π/Lz
0,0,0
ky
π/Lx
kx
Figure 1.12
Mode lattice for a three-dimensional space. The vector corresponding to eigenvalue, k1,2,3 , is shown as an example.
perfectly box-shaped with rigid boundaries. For this reason the calculation of any individual
mode frequency will rarely be accurate when the wavelength is smaller than any of the room
dimensions. We can usually expect to estimate the one-third-octave-band in which a mode
frequency will fall to an accuracy of plus or minus one-third-octave-band.
The statistical descriptor for modes is the statistical modal density, n(f ), the number of modes
per Hertz. To calculate the modal density it is necessary to arrange the eigenvalues in such
a way that facilitates counting the modes in a chosen frequency range. It is implicit in the
form of Eq. 1.15 that this can be achieved by creating a lattice in Cartesian coordinates where
the x, y, and z axes represent kx , ky , and kz (Kuttruff, 1979). This lattice of eigenvalues in
k-space is shown in Fig. 1.12, where each intersection in the lattice represents an eigenvalue
indicated by the symbol •. The length of the vector from the origin to an eigenvalue equals
kp,q,r . Eigenvalues that lie along each of the three axes represent axial modes; those that
lie on the coordinate planes kx ky , kx kz , and ky kz (excluding the eigenvalues on the axes)
represent tangential modes; all other eigenvalues (i.e. all eigenvalues excluding those on the
axes and the coordinate planes) represent oblique modes. From Eq. 1.52 it is evident that the
distance between adjacent eigenvalues in the kx , ky , and kz directions are π/Lx , π/Ly , and π/Lz
respectively. Hence the volume associated with each eigenvalue is a cube with a volume of
π3 /Lx Ly Lz , which equals π3 /V .
The number of modes below a specified wavenumber, k, is equal to the number of eigenvalues that are contained within one-eighth of a spherical volume with radius, k, as indicated
in Fig. 1.13. If there were only oblique modes this would simply be carried out by dividing
(4πk 3 /3)/8 by π3 /V ; however, the existence of axial and tangential modes means that this
20
Chapter 1
kz
k
Oblique mode
π/Lz
π/Lx
π/Ly
k
ky
Tangential mode
Axial mode
k
kx
Figure 1.13
Sketch indicating how the volumes associated with the eigenvalues for axial, tangential, and oblique modes fall inside or
outside the permissible volume in k-space. The shaded volumes indicate those fractions of the volumes associated with the
axial and tangential modes that fall outside the permissible volume in k-space. The octant volume with radius, k, encloses
eigenvalues below wavenumber, k.
would be incorrect. This is because part of the cube volume that is associated with these
mode types falls outside of the permissible volume in k-space that can only have zero or
positive values of kx , ky , and kz . From Fig. 1.13 we also see that for tangential modes on the
coordinate planes, one-half of the cube volume falls outside this permissible volume and for
axial modes, three-quarters of the cube volume falls outside. Therefore calculating the number
of modes is a three-step process. The first step is to divide (4πk 3 /3)/8 by π3 /V to give an
estimate for the number of oblique modes that also includes one-quarter of the axial modes
and one-half of the tangential modes. The second step is to account for the other halves of the
tangential modes that lie in the area on the three coordinate planes; this fraction of the total
number of modes is calculated by taking one-half of (πk 2 /4)/(π2 /(Lx Ly + Lx Lz + Ly Lz )). The
latter step included the axial modes on each of the three coordinate axes as halves. Hence
there only remains one-quarter of the axial modes that have not yet been accounted for. The
third step determines this remaining fraction of the total number of modes by taking one-quarter
of k/(π/(Lx + Ly + Lz )). The sum of these three components gives the number of modes, N(k),
below the wavenumber, k,
N(k) =
k 2 ST
kLT
k 3V
+
+
6π2
16π
16π
(1.55)
21
S o u n d
I n s u l a t i o n
where the total area of the room surfaces, ST is 2(Lx Ly + Lx Lz + Ly Lz ) and the total length of
all the room edges, LT , is 4(Lx + Ly + Lz ).
As we are working in k-space we calculate the modal density, n(ω), in modes per radian, and
then convert to the modal density, n(f ), in modes per Hertz, which is more convenient for
practical calculations. The general equation for the modal density, n(ω), in terms of ω is
n(ω) =
dN(k)
dN(k) dk
=
dω
dk dω
(1.56)
To calculate n(ω) we now need to find dk/dω which is equal to the reciprocal of the group
velocity, cg . The group velocity is the velocity at which wave energy propagates across the
space. For sound waves in air, the group velocity is the same as the phase velocity, c0 . Hence
the general equation to convert n(ω) to n(f ) is
n(f ) = 2πn(ω) =
2π dN(k)
cg dk
(1.57)
which gives the modal density for a box-shaped room as
n(f ) =
πfST
LT
4πf 2 V
+
+
3
2
8c0
c0
2c0
(1.58)
where the modal density for each frequency band is calculated using the band centre frequency.
For rooms that are not box-shaped, and for typical rooms in the high-frequency range, a
reasonable estimate of the modal density can be found by using only the first term in Eq. 1.58,
to give
n(f ) =
4πf 2 V
c03
(1.59)
Estimates for the statistical modal density of axial, tangential, and oblique modes can be
estimated from (Morse and Ingard, 1968)
naxial (f ) =
ntangential (f ) =
noblique (f ) =
6V 1/3
c0
6πfV 2/3
6V 1/3
−
2
c0
c0
3πfV 2/3
3V 1/3
4πf 2 V
−
+
3
2
2c0
c0
c0
(1.60)
(1.61)
(1.62)
for which it is assumed that LT ≈ 12V 1/3 and ST ≈ 6V 2/3 (Jacobsen, 1982). This simplifies the
calculation for rooms that are almost (but not exactly) box-shaped.
1.2.5.2 Mode count
The mode count, N, in a frequency band with a bandwidth, B, can be determined in two ways.
Either by using Eq. 1.54 to calculate the individual mode frequencies and then by counting the
number of modes that fall within the band or by using the statistical modal density to determine
a statistical mode count, Ns , in that band, where
Ns = n(f )B
22
(1.63)
Chapter 1
1000 000
Box-shaped room
Lx ⫽ 4.64 m, Ly ⫽ 3.68 m, Lz ⫽ 2.92 m
100000
Tangential modes
V ⫽ 50 m3
10000
Mode count, N
Axial modes
Oblique modes
1000
100
10
1
0.1
20
31.5
50
80
125
200
315
500
800 1250
One-third-octave-band centre frequency (Hz)
2000
3150
5000
Figure 1.14
Mode count for axial, tangential, and oblique modes in a 50 m3 box-shaped room.
In a box-shaped room, the mode with the lowest frequency will always be an axial mode. As
the band centre frequency increases, the number of oblique modes in each band increases
at a faster rate than the number of axial modes or the number of tangential modes. As an
example we can look at the trends in the mode count for a 50 m3 room. The room dimensions
are determined using the ratio 41/3 :21/3 :1 for x:y:z. This ratio is sometimes used in the design
of reverberation rooms to avoid dimensions that are integer multiples of each other; this avoids
different modes having the same frequency. As we usually work in one-third-octave-bands it
is of interest to know the number of modes that fall within each band. The mode counts for the
axial, tangential, and oblique modes are shown in Fig. 1.14. For typical rooms we can describe
the mode count using three different ranges: A, B, and C. In range A, the frequency bands
either contain no modes or a few axial and/or a few tangential modes. In range B, the blend
of the three different mode types varies between adjacent frequency bands depending on the
room dimensions. In range C, the mode count is always highest for oblique modes and always
lowest for axial modes. For this particular example, range A corresponds to one-third-octavebands below 80 Hz, range B lies between the 80 and 200 Hz bands, and range C corresponds
to bands above 200 Hz.
1.2.5.3 Mode spacing
The average frequency spacing between adjacent modes, δf , is calculated from the modal
density using
δf =
1
n(f )
(1.64)
As sound insulation calculations are almost always carried out in one-third-octave or octavebands it tends to be more informative to calculate the mode counts in these frequency bands
rather than use the mode spacing.
23
S o u n d
I n s u l a t i o n
1.2.5.4 Equivalent angles
Part of the definition of a diffuse field is that there is equal probability of a sound wave arriving
from any direction, i.e. from any angle. Hence it is instructive to look at the range of angles associated with the plane waves that form local modes. We have previously described a local mode
in a qualitative manner by following the journey of a plane wave around a room. To quantitatively describe the plane wave field we need to account for the different propagation directions
after reflection from each surface. The general equation for a plane wave (Eq. 1.13) describes
propagation in a single direction; hence each mode is comprised of more than one plane wave.
For each mode we can define equivalent angles, θx , θy , and θz ; these angles are defined from
lines that are normal to the x, y, and z-axis respectively. They are defined in k-space for any
eigenvalue in the lattice (see Fig. 1.15). For each mode, one plane wave points in the direction
of this vector in k-space. The direction of the other plane waves can be found by reflecting the
vector into the other octants of k-space. Axial, tangential, and oblique modes are therefore
described by two, four, and eight plane waves respectively. For each mode, the equivalent
angles are related to the mode wavenumber, kp,q,r , and the constants, kx , ky , and kz by
kx = kp,q,r sin θx
ky = kp,q,r sin θy
kz = kp,q,r sin θz
(1.65)
hence, from the constant definitions in Eq. 1.52, the equivalent angles for each mode are
θx = asin
pc0
2Lx fp,q,r
θy = asin
qc0
2Ly fp,q,r
θz = asin
rc0
2Lz fp,q,r
(1.66)
Later on we will need to consider the angles of incidence for the waves that impinge upon a
room surface in the calculation of sound transmission. Here we are only assessing the range
of equivalent angles for plane waves that propagate across the space to form room modes.
Figure 1.16 shows the equivalent angles for the same 50 m3 room that was used for the mode
count, where each point corresponds to a single room mode. Note that we are ignoring the
fact that specular reflection would not occur in real rooms at high frequencies. In the lowfrequency range, where there are relatively few modes, there is a limited range of angles. As
the frequency increases, the number of modes increases (the majority tending to be oblique
modes), and the range expands to cover the full range of angles between 0◦ and 90◦ . For
axial modes, one angle is 90◦ and the other two angles are 0◦ (e.g. f1,0,0 has θx = 90◦ , θy = 0◦ ,
and θz = 0◦ ). For tangential modes, one angle is 0◦ , and the other two angles are oblique. For
oblique modes, all three angles are oblique.
Equivalent angles do not in themselves identify a frequency above which the modal sound field
approximates to a diffuse field; we have already noted other important features that define a
diffuse field. However, they do illustrate how one aspect of a diffuse field concerning sound
arriving from all directions can potentially be satisfied in a modal sound field. In the study of
sound transmission it is useful to be able to switch between thinking in terms of modes, and in
terms of waves travelling at specific angles.
1.2.5.5 Irregularly shaped rooms and scattering objects
The description of local modes was based on an empty box-shaped room. Rooms in real
buildings are not all box-shaped and they usually contain scattering objects such as furniture.
24
Chapter 1
kz
kp,q,r
θy
θx
θz
ky
kx
Figure 1.15
Equivalent angles in k-space.
90
80
70
60
Angle (°)
θx
50
θy
40
θz
30
20
10
0
10
100
1000
10 000
Frequency (Hz)
Figure 1.16
Equivalent angles for the modes of a 50 m3 box-shaped room.
25
S o u n d
I n s u l a t i o n
In the laboratory, it is common to use non-parallel walls and diffusers to try and create
a diffuse field in the central zone of the room. This does not mean that the local mode
approach is instantly irrelevant; far from it. Scattering objects can be seen as coupling
together the local modes of the empty room, giving rise to hybrid versions of the original
mode shapes (Morse and Ingard, 1968). These hybrid versions no longer have the symmetrical sound pressure fields associated with individual local modes in an empty box-shaped
room. In the limit, as the room shape becomes increasingly irregular (or sufficient scattering
objects are placed inside a box-shaped room) and the room surfaces have a random distribution of acoustic surface impedance, we can effectively consider all the room modes to be
some form of oblique mode. As we approach this limit we can leave the local mode model
behind us and assume there is a close approximation to a diffuse field in the central zone of
the room.
1.2.6 Damping
In our discussion on room modes we assumed that there was a perfect reflection each time
the plane wave was reflected from a room boundary. In reality there will always be damping mechanisms that reduce the sound pressure level. When discussing room acoustics we
usually refer to absorption and reverberation times, rather than damping. From a room acoustics perspective, the sound source and the listener are located within one space, so from
the point-of-view of a listener in the source room, any sound that doesn’t return to them has
been absorbed. However, with sound insulation our concern is usually for the person that
hears the sound in the receiving room, and from their point-of-view the sound has been transmitted. As we are particularly interested in the exchange of sound energy between spaces
and structures, it is useful to start treating them in a similar manner by using the same terminology to describe absorption and transmission. Hence it is convenient to relate different
damping mechanisms to the loss factors used in Statistical Energy Analysis; these are the
internal loss factor, the coupling loss factor, and the total loss factor (Lyon and DeJong,
1995).
With internal losses the sound energy is converted into heat. Hence high internal loss factors
are beneficial for the noise control engineer who is trying to reduce sound levels. Internal losses
occur when the sound wave hits absorptive surfaces or objects (e.g. sound absorbent ceiling
tiles, carpet, porous materials) and as the wave travels through the air due to air absorption. The
former is usually more important than the latter because air absorption only becomes significant
at high frequencies and in large rooms. Information on sound absorption mechanisms and
sound absorbers can be found in a number of textbooks (e.g. Mechel, 1989/1995/1998; Mechel
and Vér, 1992; Kuttruff, 1979).
With coupling losses, the sound energy is transmitted to some other part of the building that
faces into the room. This could be an open door or window where the sound exits, never to
return. It could also be a wall or a floor in the room which is caused to vibrate by the impinging
sound waves.
The sum of the internal and coupling loss factors equals the total loss factor, and this is
related to the reverberation time of the room. We therefore start this section on damping by
deriving reflection and absorption coefficients for room surfaces that will lead to a discussion
of reverberation times and loss factors for rooms.
26
Chapter 1
y
Incident wave
p̂⫹
θ
Reflected wave
p̂⫺
x⫽0
x ⫽ ⫺⬁
x
Figure 1.17
Plane wave incident at an angle, θ , upon a surface, and the specularly reflected wave.
1.2.6.1 Reflection and absorption coefficients
A plane wave incident upon a room surface at x = 0 can be described using the term,
p̂+ exp(−ikx) where p̂+ is an arbitrary constant. The term for the reflected wave is R p̂+ exp(ikx)
where R is defined as the reflection coefficient,
R = |R| exp(iγ)
(1.67)
As seen from Eq. 1.67, the reflection coefficient is complex. It describes the magnitude and
phase change that occurs upon reflection. In diffuse fields the waves that are incident upon a
surface have random phase, so the information on the phase change is usually ignored.
We now consider a plane wave that is incident upon a surface at an angle, θ; defined such that
θ = 0◦ when the wave propagates normal to the surface. The aim here is to relate the reflection
coefficient to the specific acoustic impedance or admittance. It is assumed that the surface
is locally reacting so that the normal component of the particle velocity only depends on the
region at the surface where the sound pressure is incident.
The incident wave propagates in the xy plane towards a surface at x = 0 as shown in Fig. 1.17.
The incident wave is described by
p+ (x, y, t) = [ p̂+ exp(−ik(x cos θ + y sin θ))] exp(iωt)
(1.68)
The incident wave is specularly reflected from the surface and the reflection coefficient is used
to describe the amplitude of the reflected wave, p̂− = R p̂+ . Hence the reflected wave is
p− (x, y, t) = [R p̂+ exp(−ik(−x cos θ + y sin θ))] exp(iωt)
(1.69)
The particle velocity in the x-direction (i.e. normal to the surface) is found using Eq. 1.16, which
gives
ux = −
1 1 ∂p
iω ρ0 ∂x
(1.70)
27
S o u n d
I n s u l a t i o n
Therefore the particle velocities for the incident and reflected waves are
ux+ (x, y, t) =
cos θ
[ p̂+ exp(−ik(x cos θ + y sin θ))] exp(iωt)
ρ0 c0
ux− (x, y, t) = −
cos θ
[R p̂+ exp(−ik(−x cos θ + y sin θ))] exp(iωt)
ρ0 c0
(1.71)
(1.72)
At the surface (i.e. at x = 0), the resultant pressure is p+ + p− , and the resultant particle velocity
normal to the surface is ux+ + ux− . The ratio of the resultant pressure to this resultant particle
velocity equals the normal acoustic surface impedance (Eq. 1.29). Hence the specific acoustic
impedance of a surface is related to the reflection coefficient by
Za,s =
1 1+R
cos θ 1 − R
(1.73)
which is re-arranged to give the reflection coefficient in terms of either the specific acoustic
impedance or admittance
R=
cos θ − βa,s
Za,s cos θ − 1
=
Za,s cos θ + 1
cos θ + βa,s
(1.74)
In practice it is usually more convenient to work in terms of absorption rather than reflection.
The sound absorption coefficient, α, is defined as the ratio of the intensity absorbed by a
surface to the intensity incident upon that surface; hence it takes values between 0 and 1.
The intensity in a plane wave is proportional to the mean-square pressure (Eq. 1.19), so the
absorption coefficient is related to the reflection coefficient by
α = 1 − |R|2
(1.75)
The absorption coefficient can be calculated using Eqs 1.74 and 1.75 in terms of either the
specific acoustic impedance or admittance. For a plane wave that is incident upon a locally
reacting surface at an angle, θ, the angle-dependent absorption coefficient, αθ , is
αθ =
4βRe cos θ
4ZRe cos θ
=
2
2
2
(ZRe
+ ZIm
) cos2 θ + 2ZRe cos θ + 1
(βRe + cos θ)2 + βIm
(1.76)
where the real and imaginary parts of the specific acoustic impedance are
Za,s = ZRe + iZIm
(1.77)
and the real and imaginary parts of the specific acoustic admittance are
βa,s =
1
= βRe − iβIm
Za,s
(1.78)
At normal incidence, θ = 0◦ , hence the normal incidence absorption coefficient, α0 , is
α0 =
4βRe
2
(βRe + 1)2 + βIm
(1.79)
There can be significant variation in the absorption coefficient with angle. However, when
there is a diffuse sound field incident upon a surface we assume that there is equal probability
of sound waves impinging upon the surface from all directions. For diffuse fields we therefore
use the statistical sound absorption coefficient, αst , given by
π/2
αst =
(1.80)
αθ sin(2θ)dθ
0
28
Chapter 1
The statistical absorption coefficient is calculated from the specific acoustic admittance using
(Morse and Ingard, 1968)
αst = 8βRe
2
β2 − βIm
βIm
atan
1 + Re
2
2
βIm
βRe + βIm
+ βRe
− βRe ln
2
(βRe + 1)2 + βIm
2
2
βRe
+ βIm
(1.81)
1.2.6.2 Absorption area
Rooms not only have absorbent surfaces, but they also contain absorbent objects (e.g. furniture, people) and there will be air absorption. For practical purposes, the absorption area, A, in
m2 is useful in describing the absorption provided by surfaces, objects, and air. The absorption
area is defined as the ratio of the sound power absorbed by a surface or object, to the sound
intensity incident upon the surface or object. For a surface the absorption area is the product of
the absorption coefficient and the surface area. The absorption area essentially describes all
the absorption in the room using a single area; hence an absorption area of 10 m2 corresponds
to an area of 10 m2 that is totally absorbing.
For a room with I surfaces, J objects, and air absorption, the total absorption area, AT is
AT =
I
i=1
S i αi +
J
j=1
Aobj, j + Aair
(1.82)
Air absorption depends upon frequency, temperature, relative humidity, and static pressure.
The absorption area for air is calculated from the attenuation coefficient in air, m, in Neper/m
and the volume of air in the space, V , using
Aair = 4mV
(1.83)
The attenuation coefficient in dB/m can be calculated according to ISO 9613-1 and converted
to Neper/m by dividing by 10 lg(e).
Calculated values for Aair at 20◦ C, 70% RH and P0 = 1.013 × 105 Pa are shown in Fig. 1.18.
For rooms with an absorption area of at least 10 m2 due to surfaces and objects, Aair will only
usually form a significant fraction of A in the high-frequency range. For furnished, habitable
rooms (such as those in dwellings, commercial buildings, and schools), air absorption in the
building acoustics frequency range can often be ignored in volumes <150 m3 .
1.2.6.3 Reverberation time
When a sound source in a room is stopped abruptly, the sound energy decays away due to
the damping mechanisms that are present in the room. This feature is called reverberation
and is assessed by plotting a decay curve. This is a plot of the decaying sound pressure level
against time, starting from the time at which the sound source is stopped, usually denoted as
the time, t = 0.
For sound insulation, the reverberation time is needed to relate the sound power radiated into
a space to the average sound pressure level in that space and to quantify either the absorption
29
S o u n d
I n s u l a t i o n
7
Environmental
conditions
20°C
70% RH
1.013 ⫻ 105 Pa
Absorption area for air, Aair (m2)
6
5
4
Room volumes (m3)
50
3
100
150
2
200
1
0
50
80
125
200
315
500
800
1250 2000
One-third-octave-band centre frequency (Hz)
3150
5000
Figure 1.18
Absorption area for air in different room volumes.
in a space or the total loss factor of a space. In room acoustics, several other parameters are
used to describe different aspects of the sound field relating to reverberation and the subjective
evaluation of sound in spaces (ISO 3382).
The reverberation time, T , is the time in seconds that is taken for the sound pressure level
to decay by 60 dB, or in terms of energy, for the sound energy to decay to one-millionth of
its initial value. This definition is well-suited to the decay curve that occurs in a diffuse field;
a straight line decay as shown in Fig. 1.19. Some decay curves can be approximated by a
single straight line over the full 60 dB decay, but there are many that do not follow this simple
form. In addition, it is not always possible to measure a 60 dB decay due to the presence
of background noise. Therefore we need a definition that can be used when decay curves
have more than one slope over the 60 dB decay range. This definition also needs to quantify
the time taken for the level to decay by 60 dB by using linear regression over a specified
decay range (e.g. 30 dB) so that it is not imperative to use the full 60 dB decay. Hence, the
reverberation time is more usefully defined as the time in seconds that would be required for
the sound pressure level to decay by 60 dB when using linear regression over a specified
part of the decay curve. As we can now use any range for the linear regression, such as 10,
15, 20, or 30 dB, it is necessary to use the notation, TX , where X identifies the evaluation
range used in the linear regression, i.e. T10 , T15 , T20 , T30 . With measured decay curves,
the starting point for the linear regression is usually 5 dB below the initial level, to the end
point at X + 5 dB (ISO 354 and ISO 3382). In Section 3.8.3 we will see that 5 dB is used as
the starting point primarily because the signal processing distorts the initial part of the decay
curve.
We can now look at reverberation times with a diffuse field and a non-diffuse field in a boxshaped room.
30
Chapter 1
Sound pressure level, Lp (t) (dB)
Steady-state level before
excitation is stopped at t = 0 s
60 dB
T
t=0s
Time, t (s)
Figure 1.19
Ideal straight line decay curve showing the decrease in the sound pressure level with time after the excitation has stopped.
1.2.6.3.1 Diffuse field
In a diffuse sound field, the mean free path, dmfp , can be used to calculate the average time, t,
between two successive diffuse reflections from the room boundaries,
t =
dmfp
c0
(1.84)
we can then use t to calculate the number of diffuse reflections, N, in time, t, using
N=
t
c0 t
=
t
dmfp
(1.85)
The decay process can now be assessed by looking at the probability of waves impinging upon
surface areas in the room with different absorption coefficients. This uses the approach taken
by Kuttruff (1979). We take a room containing a sound source fed by a stationary signal such
as white noise and assume that the resulting sound field is diffuse. When the sound source is
stopped at time, t = 0, the waves continue to travel across the room volume, and each time they
impinge upon a room boundary, a fraction of the energy is reflected, with the remaining fraction
being absorbed. The binomial probability distribution is used to assess two possible outcomes
when a sound wave impinges upon a room boundary: either the wave is reflected from (and
absorbed by) surface area, S1 , or it is reflected from (and absorbed by) the remaining surface
area in the room, S2 = ST − S1 . These outcomes must be statistically independent; hence the
probability that a wave is reflected from (and absorbed by) surface area, S1 , is the same every
time that the wave impinges upon a room boundary. This is conceivable in a room where each
surface area with a different absorption coefficient is uniformly distributed over the total surface
area of the room, and there are diffuse reflections from all the room surfaces.
31
S o u n d
I n s u l a t i o n
The binomial probability distribution, P(N1 \N) gives the number of times, N1 , that a wave is
reflected from surface area, S1 , out of a total number of reflections, N. The probability of a
reflection from surface area, S1 , is S1 /ST , hence the probability of a reflection from the remaining
surface area in the room, S2 , is S2 /ST or 1 − (S1 /ST ). The binomial probability distribution is
S1 N1
S1 N−N1
N
(1.86)
P(N1 \N) =
1−
N1
ST
ST
where the binomial coefficient is calculated using
N!
N
=
N1
N1 !(N − N1 )!
(1.87)
Each time a wave impinges upon surface area, S1 , the mean-square pressure is reduced by the
factor (1 − α1 ) where α1 is the diffuse field sound absorption coefficient for surface area, S1 .
The same process occurs with surface area, S2 , with the coefficient, α2 . So, after N reflections,
of which N1 reflections are from surface area, S1 , and N − N1 reflections are from the remaining
surface area, S2 , the mean-square pressure as a function of N1 , p2 (N1 ), is
p2 (N1 ) = p2 (0)(1 − α1 )N1 (1 − α2 )N−N1
(1.88)
where p2 (0) is the mean-square pressure at time, t = 0.
Taking into account all possible values of N1 from zero to N, we can calculate the expected
value, E(N1 ) of random variable, N1 , which has the probability distribution, P(N1 \N). As we
are particularly interested in the decay of the sound pressure we note that E(N1 ) equals the
population mean (i.e. the average value) of the mean-square pressure,
E(N1 ) =
N
N1 =0
p2 (N1 )P(N1 \N) = p2 (0)
S1
ST
N
S1
(1 − α1 ) + 1 −
(1 − α2 )
ST
(1.89)
The term within the square bracket is equal to (1 − α) where α is the average diffuse field
sound absorption coefficient
α=
S 1 α1 + 1 −
S1
ST
α2
ST
(1.90)
Hence for rooms with I surface areas that each have a different absorption coefficient (instead
of I = 2 as has just been assumed), the expression can be generalized to
I
I
1
S i αi
=
α = i=1
S i αi
I
ST
i=1 Si
i=1
(1.91)
The average diffuse field sound absorption coefficient, α, is a weighted arithmetic average of
the absorption coefficients, where each coefficient has been weighted according to its surface
area. From Eq. 1.89 the mean-square pressure p2 (t) at time, t, is therefore described by
p2 (t) = p2 (0)(1 − α)N = p2 (0) exp(N ln (1 − α))
(1.92)
It is important to note from Eq. 1.92 that the mean-square pressure in a diffuse field has an
exponential decay. Therefore when plotting the sound pressure level in decibels against time,
the decay curve is a straight line. We will soon look at decay curves in non-diffuse fields, where
32
Chapter 1
the mean-square pressure does not have an exponential decay, and the decay curve is not a
single straight line across the 60 dB decay range.
At time t = T , the definition of the reverberation time is such that p2 (T ) is one-millionth of p2 (0),
hence combining Eqs 1.85 and 1.92 gives
p2 (T )
c0 T
−6
= 10 = exp
ln (1 − α)
p2 (0)
dmfp
(1.93)
This gives the reverberation time formula that is commonly referred to as Eyring’s equation
(Eyring, 1930),
T =
−dmfp 6 ln 10
−24V ln 10
=
c0 ln (1 − α)
c0 ST ln (1 − α)
(1.94)
and by taking air absorption into account, this becomes
T =
−24V ln 10
c0 (ST ln (1 − α) − 4mV )
(1.95)
When considering the effect of the room volume, V , and the total surface area, ST , on the diffuse
field reverberation time, it is useful to think in terms of the mean free path. The reverberation
time is proportional to the mean free path for a diffuse field (Eq. 1.47), hence the longer the
mean free path, the longer the time between successive reflections from the room surfaces.
So if we choose a point in this room to measure the reverberation time, it will have taken longer
for the waves to travel around the room before returning to our chosen point. Each time the
wave hits a surface, a fraction of the wave energy will be absorbed. Assuming a fixed value
for the absorption coefficient of room surfaces, the reverberation time will therefore increase
with increasing room volume.
For a diffuse field where the average diffuse field absorption coefficient, α, is much smaller
than unity, we can assume that α ≈ − ln (1 − α). Using this approximation in Eq. 1.94 leads to
Sabine’s equation (Sabine, 1932),
T =
dmfp 6 ln 10
24V ln 10
=
c0 α
c0 ST α
(1.96)
To take account of absorption from objects and the air, as well as from the room surfaces,
Eq. 1.96 is more conveniently written in terms of the total absorption area, AT (Eq. 1.82) as
T =
24V ln 10
c0 AT
(1.97)
Assuming that the steady-state sound pressure level in the diffuse field is 60 dB at time t = 0,
the time-varying sound pressure level in decibels that defines the idealized decay curve is
Lp (t) = 10 lg
p2 (t)
p2 (0)
= 60 −
60t
T
(1.98)
The Sabine equation is based on the assumption that α is sufficiently small that α ≈ − ln (1−α),
whereas the Eyring equation is applicable to any value of α. In general, Eyring’s equation gives
reasonable estimates in rooms where there is uniform surface absorption and diffuse surface
reflections, however, it is also appropriate in box-shaped rooms with uniform surface absorption
and specular surface reflections (Hodgson, 1993, 1996).
33
S o u n d
1.2.6.3.2
I n s u l a t i o n
Non-diffuse field: normal mode theory
In non-diffuse fields the decay curves cannot usually be approximated by a straight line across
the entire 60 dB decay range. To understand some of the reasons for this, we return to consider
local room modes in a box-shaped room. In each frequency band, the decay curve will be
determined by the individual room modes that are decaying within that band, and the interaction
between these modes.
We will focus on decay curves in the low-frequency range. When we consider the mode count in
one-third-octave-bands for typical rooms, bands in the low-frequency range have relatively few
modes compared to those in the mid- and high-frequency ranges (refer back to the 50 m3 room
in Fig. 1.14). In reality, we cannot strictly compartmentalize the decaying modes into individual
frequency bands. This is due to the damping associated with each mode; a decaying mode may
influence the decay in the two bands that are adjacent to the band in which the mode is strictly
assumed to fall. However, compartmentalization is used here to provide some insight into the
way that axial, tangential, and oblique modes determine the decay curve for a band. The normal
mode theory used to calculate the decay curves is taken from Kuttruff (1979) and Bodlund
(1980); the latter reference also provides corrections to earlier investigations by Larsen (1978).
Assume that we have a box-shaped room with locally reacting surfaces. This room contains
a sound source that is fed by a sinusoidal signal with the same frequency as the mode of
interest. In reality, most rooms have modes that are relatively closely spaced. This means that
a sinusoidal signal will also excite other room modes unless all the surfaces have very low
absorption coefficients and the modes are all well-separated in terms of frequency. However,
here we will assume that we are able to excite only a single mode. When the sound source
is stopped at time, t = 0, the waves continue to travel along the path that is defined for this
particular room mode. Upon each reflection from a room surface, a fraction of the energy is
reflected, and the remaining fraction is absorbed.
When looking at the reverberant decay of an individual mode, m, the mean-square pressure
decays away exponentially according to
p2 (t) = p2 (0) exp(−2δm c0 t)
(1.99)
where δm = βa,s (εp,m /Lx + εq,m /Ly + εr ,m /Lz ) in which βa,s is the specific acoustic admittance,
and εp,m , εq,m and εr ,m correspond to mode, fp,q,r (if p = 0, then εp,m = 1 else εp,m = 2; if q = 0,
then εq,m = 1 else εq,m = 2; if r = 0 then εr ,m = 1 else εr ,m = 2).
The reverberation time, Tm , for an individual mode can be calculated from Eq. 1.99 at time
t = Tm using
2
pm
εp,m
εq,m
εr ,m
(Tm )
−6
=
10
=
exp
−2c
T
β
+
+
(1.100)
0
m
a,s
2 (0)
pm
Lx
Ly
Lz
which gives,
Tm =
3 ln 10
c0 βa,s
εp,m
Lx
+
εq,m
Ly
+
εr ,m
Lz
(1.101)
The denominator in Eq. 1.101 is referred to as the damping constant of the mode. From
Eq. 1.101 it is possible to identify three trends for the different mode types when the specific acoustic admittance is independent of frequency: (1) the axial modes associated with
each room dimension have different reverberation times to each other when Lx = Ly = Lz ;
34
Chapter 1
(2) when the tangential modes are considered in three groups defined by p = 0, q = 0, and
r = 0, each group will have the same reverberation time; and (3) all oblique modes have the
same reverberation time.
For engineering calculations it is convenient to relate the reverberation time directly to the
absorption coefficient. Normal mode theory uses the specific acoustic admittance, which is the
reciprocal of the specific acoustic impedance; hence it is linked to the absorption coefficient.
From Eq. 1.76 we see that the absorption coefficient is dependent upon the angle of incidence
and the specific acoustic impedance. Depending on the mode, the waves will be incident
upon the room surfaces at different angles. For axial modes the waves always impinge upon
two opposite surfaces at an angle of incidence that is normal to these surfaces. For oblique
and tangential modes, the angle of incidence varies depending upon the mode and the room
boundary upon which the waves are impinging. For simplicity, the angle dependence is ignored
in the following examples and a single value for the specific acoustic admittance is used for all
of the room surfaces. This still allows us to see the general effect of the modes on the decay
curves; we simply acknowledge that the situation is more complex in reality.
Equation 1.101 can now be used to calculate the decay curve for each of the M individual
modes within a frequency band. This can be compared with the decay curve for the frequency
band itself. It is assumed that the sound source is fed with a white noise signal with an rms
volume velocity spectral density, Qsd . It is convenient to set the sound pressure level at t = 0
to a level of 60 dB for the frequency band, hence we need to establish the level at t = 0 for
each of the M modes in that band. For the mth mode, the spatial average mean-square sound
pressure at time t = 0 is (Bodlund, 1980)
2
pm
s =
2
ρ02 c04 Qsd
π
Tm
2
12V ln 10
(1.102)
for which it is assumed that the specific acoustic admittance for the room surfaces is a real
value, much less than unity, and uniform over all the surfaces.
Using Eq. 1.102 to set the level for each mode at t = 0, the sound pressure level, Lp,m (t), in
decibels, for the mth decaying mode in a frequency band is
2
60t
60t
pm
s
Tm
pm (t) 2
Lp,m (t) = 10 lg
= 60 −
= 60 −
+ 10 lg M
+ 10 lg M
2
pm (0)
Tm
Tm
m=1 pm s
m=1 Tm
(1.103)
The sound pressure level, Lp (t), in decibels for the frequency band can then be calculated from
the energetic sum of the decay curves for the individual modes in the band,
Lp (t) = 10 lg
M
m=1
10Lp,m (t)/10
(1.104)
We can now look at decays in rooms with the same volume (50 m3 ) but different Lx , Ly , and
Lz dimensions. The specific acoustic admittance for all room surfaces is assumed to be real,
independent of frequency, and independent of the angle of incidence. Although it does not
correspond to any particular material commonly used for walls and floors, a value of βa,s = 0.01
is used to give reverberation times less than 2 s. Note that smooth, heavy concrete walls and
floors would usually have much smaller values, which would lead to longer reverberation times.
In contrast to measured decay curves, the decay curves from this model can be evaluated from
35
S o u n d
I n s u l a t i o n
Sound pressure level, Lp(t) and Lp,m(t) (dB)
60
All 9 modes in the 125 Hz
one-third-octave band (T15 = 1.57 s)
50
Axial modes
(0,3,0 : 0,0,2 : 4,0,0)
40
Tangential modes
(3,0,1 : 1,3,0 : 3,2,0)
30
Oblique modes
(1,2,1 : 2,2,1 : 3,1,1)
Room No. 1
Lx ⫽ 5 m
20
Ly ⫽ 4 m
Lz ⫽ 2.5 m
10
V ⫽ 50 m3
ST ⫽ 85.0 m2
0
0
0.2
0.4
0.6
0.8
Sound pressure level, Lp(t) and Lp,m(t) (dB)
60
1
1.2
Time, t (s)
1.4
1.6
1.8
2
All 11 modes in the 125 Hz
one-third-octave band (T15 = 1.43 s)
50
Axial mode
(5,0,0)
40
Tangential modes
(4,1,0 : 2,2,0 : 4,0,1 : 0,2,1 : 3,2,0 : 5,1,0)
Oblique modes
(3,1,1 : 1,2,1 : 4,1,1 : 2,2,1)
30
Room No. 2
Lx ⫽ 6.88 m
20
Ly ⫽ 3.16 m
Lz ⫽ 2.3 m
10
V ⫽ 50 m3
ST ⫽ 89.7 m2
0
0
0.2
0.4
0.6
0.8
1
1.2
Time, t (s)
1.4
1.6
1.8
2
Figure 1.20
Decay curves for box-shaped rooms No. 1 and No. 2. Each room has a volume of 50 m3 but with different Lx , Ly , and Lz
dimensions. The curves are shown for individual modes in the 125 Hz one-third-octave-band along with the resulting curve
for that band.
t = 0 to calculate the reverberation time. This is because the model does not include the effect
of direct sound from the sound source, and, unlike a measurement, there is no distortion of the
initial part of the decay curve from the signal processing.
As sound fields in practice can rarely be considered as diffuse in the low-frequency range
we will initially look at the 125 Hz one-third-octave-band. The decay curves from two different
50 m3 rooms (No. 1 and No. 2) are shown in Fig. 1.20 for the frequency band and the individual
36
Chapter 1
modes in this band. For each individual mode the decay curves are straight lines. In contrast,
the resulting decay curve for the frequency band is a curve; this is easier to see if a straight
edge is placed against it. The decay curves for these two rooms are different. However, in this
particular example the initial part of the decay curve has a similar slope in both rooms, hence
the reverberation time, T15 , is similar too.
The axial modes tend to have longer reverberation times than the tangential modes, which, in
turn, tend to be higher than for the oblique modes. For this reason the energy of individual axial
modes at t = 0 is slightly higher than individual tangential or oblique modes. We recall that the
decay curve for the frequency band is calculated from the energetic sum of the decays for the
individual modes (Eq. 1.104). Therefore, it is only in the early part of the decay, say within
the initial 20 or 30 dB, that the majority of the different room modes play a role in determining
the decay curve of the frequency band. In the later part of the decay, the decay curve for the
frequency band is primarily determined by the modes with the longest reverberation times.
These are always axial modes. This is clearly seen with room No. 2 where there is only a single
axial mode, f5,0,0 in the frequency band. In the late part of the decay curve, the slope is primarily
determined by this one axial mode. Hence for typical rooms in the low-frequency range, the axial
modes play an important role in determining the decay curve of the frequency band. This will be
more apparent when one room dimension is significantly longer than the other two dimensions.
In this situation, the decay curve of the frequency band is predominantly determined by the
axial mode(s) with wave propagation along the longest dimension. This allows some insight
into which surfaces require low-frequency absorbers to reduce the reverberation time of the
frequency band by reducing the reverberation time of specific modes.
In practice, the measured decay curve for the frequency band will fluctuate about the predicted
straight line decay due to interaction between the modes causing beating. Also, in using a
single value for the specific acoustic admittance we have effectively assumed a single absorption coefficient for all angles of incidence which is not appropriate for many common walls and
floors. In the low-frequency range the absorption coefficient can be lower at normal incidence
than at oblique incidence. This would make the curvature more distinct due to even longer
reverberation times for the axial modes.
The model also allows us to compare the decay curve for a diffuse field with the decay curves
for individual frequency bands as the band centre frequency increases. To do this we will
choose the 100, 1000, and 5000 Hz one-third-octave-bands for a different 50 m3 room (room
No. 3). It is important to note that the assumption of purely specular reflection for the 1000
and 5000 Hz bands is unrealistic in practice as walls and floors are often slightly irregular
with scattering objects near the room surfaces. However, it gives us a useful insight into the
effects of different mode counts and the different blends of mode types in each frequency band.
Figure 1.21a shows the decay curves for the three frequency bands. The curvature of the decay
in the 100 Hz band is in marked contrast to the approximately straight decay of the 5000 Hz
band. The reasons for this difference can be seen by grouping together the decay curves for
each mode type (axial, tangential, or oblique) as shown in Fig. 1.21b. In the 100 Hz band
the initial 20 dB decay is determined by all three mode types. However, the later part of the
decay curve is predominantly determined by the axial modes with minimal influence from
the tangential and oblique modes. This is in contrast to the 1000 and 5000 Hz bands where
the number of axial modes is small compared to the number of tangential or oblique modes;
hence, there is only a minor influence from the axial modes on the decay curve for the frequency
band. As the frequency increases, there are many more oblique modes than axial or tangential
37
S o u n d
I n s u l a t i o n
modes. As all oblique modes have the same reverberation time, and the decay curve for each
individual mode is a straight line, it follows that the decay curve at high frequencies tends
towards a straight line determined by the oblique modes. This is seen in Fig. 1.21b where the
decay curve for the 5000 Hz band is very similar to the grouped decay curve for the oblique
modes.
For a room with uniform locally reacting surfaces, normal mode theory gives a useful insight
into the reasons for curvature of the decay curves. However, it does not fully describe the
degree of curvature that occurs in practice. This is partly due to the fact that walls and floors
are not purely locally reacting; they can act as surfaces of extended reaction.
Locally reacting surfaces and surfaces of extended reaction: It is very convenient to be
able to consider the reverberation time as independent of any interaction between the room
modes and the structural modes of the walls and floors. So far it has been assumed that
although the sound waves impinging upon a surface are absorbed, this absorption is a ‘local
matter’ between the sound wave and the point on the surface from which it is reflected. From
the point-of-view of an impinging sound wave there are two types of room surface that are
responsible for absorption: locally reacting surfaces and surfaces of extended reaction (Morse
and Ingard, 1968). These two types can be defined by referring to the normal acoustic surface
impedance; this is the ratio of the complex sound pressure at the surface to the component of
the complex sound particle velocity that is normal to the surface. If a wave impinges upon a point
on the surface, a locally reacting surface is one where the particle velocity normal to the surface
is only affected by the sound pressure at that point, and is unaffected by the pressure at adjacent
points on the surface. In contrast, a surface of extended reaction is one where the particle velocity normal to the surface is affected by the pressure at adjacent points on the surface. Surfaces
of extended reaction therefore include plates undergoing bending wave motion, and porous
surfaces where the sound propagates inside the porous material in a direction parallel to the
surface.
60
100 Hz (T15 = 1.63 s)
Sound pressure level, Lp(t) (dB)
50
1000 Hz (T15 = 1.26 s)
5000 Hz (T15 = 1.22 s)
40
30
Room No. 3
Lx ⫽ 4.64 m
20
Ly ⫽ 3.68 m
Lz ⫽ 2.92 m
10
V ⫽ 50 m3
ST ⫽ 82.7 m2
0
0
0.2
0.4
0.6
0.8
1
1.2
Time, t (s)
1.4
Figure 1.21(a)
Decay curves for box-shaped room No. 3 (100, 1000, and 5000 Hz one-third-octave-bands).
38
1.6
1.8
2
Chapter 1
60
100 Hz
Sound pressure level, Lp(t) (dB)
50
Axial modes (2)
Tangential modes (3)
40
Oblique modes (1)
30
20
10
0
0
0.2
0.4
0.6
0.8
1
1.2
Time, t (s)
1.4
1.6
1.8
2
60
1000 Hz
Sound pressure level, Lp(t) (dB)
50
Axial modes (15)
Tangential modes (496)
40
Oblique modes (3386)
30
20
10
0
0
0.2
0.4
0.6
0.8
1
1.2
Time, t (s)
1.4
1.6
1.8
2
60
5000 Hz
Sound pressure level, Lp(t) (dB)
50
Axial modes (76)
Tangential modes (12821)
40
Oblique modes (452686)
30
20
10
0
0
0.2
0.4
0.6
0.8
1
1.2
Time, t (s)
1.4
1.6
1.8
2
Figure 1.21(b)
Decay curves for the 100, 1000, and 5000 Hz one-third-octave-bands along with the grouped decay curve for each mode
type. The number of modes corresponding to each mode type in the one-third-octave-band are shown in brackets.
39
S o u n d
I n s u l a t i o n
For practical purposes, the assumption of locally reacting room surfaces is very useful in simplifying the calculation of reverberation time. In many rooms, small areas of rigid frame, porous,
absorbent material are fixed to the room surfaces to reduce the reverberation time. Many of
these absorbers can be considered as locally reacting. The assumption of locally reacting surfaces is often reasonable due to other factors that introduce uncertainty into the calculation;
these are non-uniform distribution of absorption, application of laboratory measurements of
absorption coefficients to rooms with non-diffuse fields, specific room geometry, and scattering from objects and surfaces in the room. Bare walls and floors undergoing bending wave
motion are surfaces of extended reaction, responsible for transmitting sound to other parts of
the building. In this situation the reverberation time is not only determined by the room modes,
it is determined by the interaction between the room modes and the structural modes of the
walls and floors (Pan and Bies, 1988). This blurs the boundary between the study of room
acoustics and structure-borne sound.
For a room with locally reacting surfaces that have a frequency-independent value for the
specific acoustic admittance, the normal mode model indicates that the group of oblique modes
will have the same reverberation time, and each group of tangential modes with p = 0 or q = 0 or
r = 0 have the same reverberation time. However, experimental evidence from a reverberation
chamber with 280 mm thick concrete walls and floors indicates that there can be significant
variation between the reverberation times of individual modes within these groups (Munro,
1982). This is primarily because the walls and floors are not locally reacting (Pan and Bies,
1988). It has been shown both theoretically and experimentally that the reverberation time in a
room can be altered by changing the total loss factor and/or the modal density of its walls and
floors (Pan and Bies, 1990). Experiments in the same reverberation chamber demonstrate that
by increasing the total loss factor of a bending wave mode on one concrete wall (by wedging
wooden blocks between this wall and another wall to increase the structural coupling losses),
it is possible to change the reverberation time of an individual room mode (Pan and Bies, 1988).
However, this needs to be kept in perspective when predicting reverberation times in rooms.
There are other reasons why it is difficult to accurately predict reverberation times; mainly
the existence of non-diffuse sound fields and the application of laboratory measurements of
absorption or scattering coefficients to a specific situation in the field. The fact that walls and
floors are not locally reacting is simply one more reason.
In practice, walls and floors are usually partly or completely covered with a locally reacting
absorber, such as carpet on a heavy concrete floor. Therefore it is not always necessary
for calculations to consider the effect of modal interaction; reasonable estimates can often be
obtained by assigning an absorption coefficient to the areas of wall and floor that are not covered
by the locally reacting absorber. This absorption coefficient may be based on measurements
or empiricism. In many cases the wall or floor will act as both a surface of extended reaction
and a locally reacting surface, although one of these may be more important than the other. For
example, some masonry/concrete walls have highly porous surfaces. A reasonable estimate
of the reverberation time can often be found by using a measured absorption coefficient and
simply treating the wall as a locally reacting surface; it may be unnecessary to consider the fact
that it also acts as a surface of extended reaction due to bending wave vibration. This allows
calculation of the room reverberation time using absorption coefficients in equations such as
Eq. 1.94 or 1.96.
Despite the fact that real walls and floors are not purely locally reacting, normal mode theory shows that the curved decay for a frequency band is due to the different reverberation
40
Sound pressure level, Lp(t) (dB)
Chapter 1
Early part of the decay
60 dB
Later part of the decay
z
y
x
t⫽0s
Time, t (s)
Figure 1.22
Example of a decay curve with a double slope; this can occur in rooms with a highly absorptive ceiling, but where the walls
and the floor have relatively low absorption.
times for the modes within that band. Interaction between room modes and structural modes
therefore results in a range of reverberation times for individual modes; hence there will still
be curved decays in some frequency bands. In fact, measured data suggest that the range
of reverberation times for individual modes is much larger than calculated from normal mode
theory, resulting in decay curves with a greater degree of curvature (Bodlund, 1980).
1.2.6.3.3 Non-diffuse field: non-uniform distribution of absorption
Non-diffuse fields also occur due to non-uniform distribution of absorption over the room surfaces; one common example occurs when there is a highly absorptive ceiling but the walls
and the floor have relatively low absorption. In these situations the decay curve can also show
curvature or a distinct double slope as illustrated in Fig. 1.22. Considering the different modes,
it is possible to make a basic qualitative assessment of the reasons for this double slope. When
the early part of the decay is predominantly determined by the oblique modes (as in the previous example for the 5000 Hz band) we can expect large numbers of these modes to be rapidly
attenuated as they impinge upon the highly absorbent ceiling. This gives rise to the fast decay
rate in the early part of the decay curve. However, some of the axial and tangential modes will
only be reflected from the side walls which have low absorption. Hence we can expect these
modes to have relatively long reverberation times and contribute to the late part of the decay,
which compared to the early part, will have a much slower rate of decay. The main features of
the decay curve can be predicted by dividing the modes into two groups (Nilsson, 2004). The
first group contains modes where the waves propagate almost parallel to the ceiling (grazing
waves). In the second group the modes propagate at angles that are oblique to the ceiling
(non-grazing waves). Using this grouping, the non-grazing waves determine the early part of
the decay curve and the grazing waves determine the late part of the curve. Other prediction
41
S o u n d
I n s u l a t i o n
formulae for rooms with non-uniform distribution of absorption can be found in work by Fitzroy
(1959), Arau-Puchades (1988), and Neubauer (2001).
1.2.6.4 Internal loss factor
In later chapters we will look at predicting sound transmission between two rooms using Statistical Energy Analysis (SEA). It will then become useful to denote the different rooms using
a subscript. The internal loss factor is usually denoted as ηint but here we will start using the
notation, ηii , for the internal loss factor of a room subsystem, i, in an SEA model.
Internal losses describe the conversion of sound energy into heat by absorption; if this is the
only process that is described by the total absorption area, the internal loss factor is given by
ηint = ηii =
c 0 AT
8πfV
(1.105)
1.2.6.5 Coupling loss factor
The coupling loss factor, ηij , describes resonant transmission between a room (subsystem i)
and a plate (subsystem j) that faces into the room. This is described in Section 4.3.1.1.
1.2.6.6 Total loss factor
The total loss factor, ηi , of a subsystem, i, is the sum of its internal loss factor and all the
coupling loss factors from that subsystem,
ηi = ηii +
J
ηij
j=1
(i = j)
(1.106)
and is related to the reverberation time by
ηi =
2.2
6 ln 10
=
2πfT
fT
(1.107)
For most rooms, the sum of the coupling loss factors is much smaller than the internal loss
factor, and the latter provides a reasonable estimate of the total loss factor. For a room where
the total absorption area, AT , is calculated from the measured reverberation time (and hence
includes both internal and coupling losses), the total loss factor can be written as
ηi =
c0 AT
8πfVi
(1.108)
1.2.6.7 Modal overlap factor
The modal overlap factor, M, describes the degree of overlap in the modal response. It is
defined as the ratio of the 3 dB modal bandwidth, f3 dB , to the average frequency spacing
between mode frequencies, f , and is calculated from
M=
f3 dB
= f ηn
f
(1.109)
where f3 dB (which is also referred to as the half-power bandwidth) is equal to the frequency
spacing between the two points on the modal response where the level is 3 dB lower than the
peak level, and η is the loss factor.
42
Chapter 1
M << 1
Mⴝ1
⌬f
⌬f
Level (dB)
3 dB
⌬f3dB
Frequency (Hz)
⌬f3dB
Frequency (Hz)
Figure 1.23
Modal response for two adjacent modes with modal overlap factors, M ≪ 1 and M = 1.
An example of the response due to two adjacent modes with a frequency spacing, f is shown
in Fig. 1.23. This idealized response could represent either the sound pressure level in a room
or the velocity level on a wall. When M ≪ 1 there is no overlap of the 3 dB bandwidths and
there can be deep troughs between the two modes. When M = 1, the modal responses overlap
at the point where the levels are 3 dB below the peak level. As M ≫ 1 the response becomes
increasingly uniform due to the absence of deep troughs.
Figure 1.24 shows the modal overlap factor for different room volumes and reverberation times.
The modal overlap factor is often less than unity in the low-frequency range.
A cut-off frequency, fM , that identifies the lowest frequency associated with a minimum value of
the modal overlap factor can be found by substituting Eqs 1.59 and 1.107 in Eq. 1.109 to give
MTc03
fM =
(1.110)
8.8πV
With a modal overlap factor of three, this cut-off frequency is often referred to as the Schroeder
cut-off frequency, fS , and quoted as (Schroeder, 1962)
T
fS = 2000
(1.111)
V
For a room of fixed volume, long reverberation times mean that the damping loss factor is low;
hence the modal overlap is also low which results in higher cut-off frequencies. Usually we
want to calculate the cut-off frequency from measured reverberation times. When these are
approximately constant over the building acoustics frequency range, the average reverberation
43
S o u n d
I n s u l a t i o n
1000
Modal overlap factor (⫺)
100
Volume (m3)
10
50
100
T⫽2s
150
1
50
100
T⫽1s
150
0.1
50
80
125
200
315
500
800
1250 2000
One-third-octave-band centre frequency (Hz)
3150
5000
Figure 1.24
Modal overlap factors for different room volumes and reverberation times.
time can be used. Otherwise, an initial estimate for the cut-off frequency can be found from the
arithmetic average of the reverberation time over a large part of the frequency range. This will
then identify a more relevant part of the frequency range over which the reverberation times
can be averaged to refine the estimate.
Figure 1.25 shows the Schroeder cut-off frequency for a range of room volumes and reverberation times. For room volumes less than 60 m3 with reverberation times between 0.5 and 1 s,
the lowest cut-off frequency will be in the 200 Hz one-third-octave-band, and often in higher
frequency bands.
1.2.7
Spatial variation in sound pressure levels
In the measurement and prediction of sound insulation it is almost always the temporal and spatial average sound pressure level in each room that is of interest rather than the level at a particular point in space at a particular point in time. For this reason, measurement procedures require
time-averaged sound pressure levels to be measured at a number of different points in a room
and averaged. However, an average value is only useful in the analysis of sound insulation
measurements and predictions if we know what it represents. It is therefore necessary to look
at the spatial variation of time-averaged sound pressure levels both in theory and in practice.
We start with the theory for the sound field near room boundaries, and then move on to discuss
the sound pressure level distribution in a room due to individual modes and in the idealized
diffuse sound field. We then consider practical situations where there is a direct sound field
near the loudspeaker, and spaces in which the sound pressure level decreases with distance.
This allows us to interpret some example measurements of sound fields in frequency bands,
and to see the benefit in using statistical descriptions for the spatial variation in the sound
pressure level.
44
Chapter 1
1000
Schroeder cut-off frequency (Hz)
900
Reverberation time (s)
800
0.5
700
1.0
600
1.5
500
2.0
400
300
200
100
0
20
40
60
80
100
120
Volume (m3)
140
160
180
200
Figure 1.25
Schroeder cut-off frequency for different room volumes and reverberation times.
1.2.7.1 Sound fields near room boundaries
In a diffuse field the phase relationship between all waves passing through a single point in
space is random. However, near a room boundary there will be a non-random phase relationship between the incident wave and the reflected wave. When a sound wave is incident upon
a room boundary the reflected wave combines with the incident wave to give an interference
pattern in the vicinity of this boundary (Waterhouse, 1955).
Initially it is assumed that all the room boundaries are perfectly reflecting and rigid. Under this
assumption the sound field close to any surface, edge, or corner in a room can be compared
to the sound field far from the surface. This approach is often used to quantify the total sound
energy stored in a room where it can be assumed that there is a diffuse field in the central zone
of the room. In practice there are a wide range of acoustic surface impedances for walls and
floors in buildings and it is necessary to be aware of their effect on the sound field.
1.2.7.1.1 Perfectly reflecting rigid boundaries
To gain an insight into the sound field near a wall or floor in a room, we start with the situation
where a harmonic plane wave is incident upon a surface, such as a wall or floor, at an angle that
is perpendicular to the surface, i.e. at normal incidence. The surface is positioned at x = 0 in
the yz plane (see Fig. 1.26). It is assumed that the surface is large compared to the wavelength
and that there are no other surfaces that affect the sound field. The incident wave, p̂+ exp(−ikx)
travels from −∞ towards x = 0 where it is reflected from the surface to give the reflected wave,
p̂− exp(ikx). This conveniently means that the exponential terms equal unity at the surface
where x = 0. The resulting sound pressure due to the incident and reflected waves is
p(x, t) = [ p̂+ exp(−ikx) + p̂− exp(ikx)] exp(iωt)
(1.112)
where p̂+ is an arbitrary constant for the incident wave. The constant for the reflected wave,
p̂− , is related to p̂+ by the reflection coefficient of the surface, R (Eq. 1.67), where p̂− = R p̂+ .
45
S o u n d
I n s u l a t i o n
Incident wave
p̂⫹
p̂⫺
Reflected wave
x ⫽⫺⬁
x⫽0
x
Figure 1.26
Plane waves incident upon, and reflected from a room boundary.
By squaring the real part of Eq. 1.112 and taking the time-average, the mean-square sound
pressure at a distance, x, from the surface in the negative x-direction is
2
(0.5 + 0.5|R|2 + |R| cos(2kx + γ))
p2 t = p̂+
(1.113)
For perfect reflection from a rigid surface there is no phase shift (i.e. γ = 0), so R = |R| = 1 and
Eq. 1.113 becomes
2
(1 + cos(2kx))
p2 t = p̂+
(1.114)
2
At x = 0, the sound pressure for the incident and reflected waves is in phase, so p2 t = 2p̂+
. At
x = λ/4 the incident and reflected waves are out of phase with each other and the mean-square
pressure is zero.
The particle velocities for the incident and reflected waves are found from the sound pressure
terms using Eq. 1.18. This gives ρp̂0+c0 exp(−ikx) for the incident wave and − ρp̂0−c0 exp(ikx) for the
reflected wave that travels in the opposite direction. The resulting particle velocity is,
u(x, t) =
1
[ p̂+ exp(−ikx) − p̂− exp(ikx)] exp(iωt)
ρ0 c0
(1.115)
At x = 0, the particle velocity is zero when the surface is perfectly reflecting and rigid (i.e.
p̂− = p̂+ ).
In practice, sound waves in a room are incident from many different directions upon a reflecting surface, so the next step is to consider a single wave that is incident at an angle, θ, to the
x-axis. For the reflected wave we will assume specular reflection from the surface. For an
oblique angle of incidence, the Cartesian coordinate system is rotated by θ so that x in
Eq. 1.112 is replaced by x ′ , where x ′ = x cos θ + y sin θ. This gives
p(x, y, t) = [ p̂+ exp(−ik(x cos θ + y sin θ)) + p̂− exp(−ik(−x cos θ + y sin θ))] exp(iωt) (1.116)
Hence for oblique incidence, the time-averaged mean-square pressure at a distance, x, from
a perfectly reflecting, rigid surface is
2
(1 + cos(2kx cos θ))
p2 t = p̂+
46
(1.117)
Chapter 1
(a)
(b)
y
y
sin θ d φ
dθ
dV
φ
θ
x
x
z
z
Figure 1.27
(a) Sound waves incident from all possible angles upon a perfectly reflecting surface form a hemisphere around a small area
on the surface. (b) Spherical coordinate system with the element of the solid angle used to average the mean-square pressure
over the hemisphere.
We can now consider waves that are incident upon the surface from all possible angles. This
forms a hemisphere that encloses a small area on the surface as shown in Fig. 1.27. The
incident waves are assumed to be incoherent, i.e. to have random phase, therefore the meansquare pressure in Eq. 1.117 can be averaged over all possible angles of incidence using
p2 t =
1
2π
0
2π
π/2
0
sin (2kx)
2
2
(1 + cos (2kx cos θ))] sin θ dθdφ = p̂+
1+
[ p̂+
2kx
(1.118)
where the spherical coordinate system is shown in Fig. 1.27 and the element of the solid
angle, d, is sin θ dθdφ.
From Eq. 1.118 the asymptotic value for the mean-square pressure at a distance far from the
2
2
surface,p∞
t , is equal to p̂+
. At this point it is convenient to change over from using negative
x values for the distance and use positive values for the distance, d, along the x-axis. Hence
the ratio of the mean-square pressure at a distance, d, from this surface, to the mean-square
pressure at a point far away from the surface is (Waterhouse, 1955)
p2 t
sin(2kd)
=1+
2
p∞ t
2kd
(1.119)
This is plotted in Fig. 1.28 as the sound pressure level difference in decibels against 2kd. The
smallest distance at which there is no difference between the level near the surface and the
level far away from the surface occurs at 2kd = π, where d = λ/4. The largest level differences
47
S o u n d
I n s u l a t i o n
3.0
Sound pressure level difference (dB)
2.5
2.0
1.5
1.0
0.5
0.0
⫺0.5
⫺1.0
d⫽λ/4
⫺1.5
0
1
2
3
4
5
6
7
8
9 10 11 12 13 14 15 16 17 18 19 20
2kd
Figure 1.28
Sound pressure level difference between a point at a distance, d, from a perfectly reflecting surface and a point far away from
the surface.
(magnitude) occur at distances less than λ/4 from the surface. At high frequencies and/or
large distances from the surface the sound pressure level difference tends towards 0 dB. If we
start at a distance d = λ/4 from the surface and move towards the surface the level difference
tends towards 3 dB. Equation 1.119 applies to single frequencies rather than frequency bands,
although it also gives reasonable estimates for one-third-octave-bands and octave-bands of
white noise at distances up to λ/4 from the surface (Waterhouse, 1955). To illustrate its practical
application, the level difference against frequency at different distances from the reflecting
surface is shown in Fig. 1.29.
When measuring the reverberant sound pressure level in a room it is necessary to avoid
measuring near the room boundaries in these interference patterns. The requirements for
the minimum distance between a microphone and the room boundaries depend on the level
of accuracy required, and the practical aspect of finding sufficient measurement positions
in small rooms. Measurements are usually carried out simultaneously in all the frequency
bands over the building acoustics frequency range. For this reason it is common to quote the
minimum distance as a fixed value based upon the lowest frequency of interest; rather than
quoting a fraction of a wavelength. In the Standards for field and laboratory sound insulation
measurements the minimum measurement distances from the room boundaries are quoted as
0.5, 0.7, and 1.2 m (ISO 140 Parts 3, 4, 5, 6, & 7). Equation 1.119 can be used to estimate the
level difference at these distances from walls or floors. At 50 Hz the level difference is 1.4 dB
for a distance of 1.2 m. At 100 Hz a level difference of 1.8 dB occurs for a distance of 0.5 m,
and 0.8 dB for 0.7 m.
The interference patterns at the edges and the corners also need consideration. The ratio for
the mean-square sound pressure at a distance, d, from an edge or a corner due to both the
incident and reflected waves, relative to the mean-square sound pressure at a point far away
from the edge or corner is given by Waterhouse (1955).
48
Chapter 1
Sound pressure level difference (dB)
3
2.5
0.1 m
2
Distance d from the
surface is shown in
steps of 0.1 m, from
0.1 to 1.2 m
1.5
1
0.5
1.2 m
0
20
60
100
140
180
220 260 300
Frequency (Hz)
340
380
420
460
500
Figure 1.29
Sound pressure level difference between a point at a distance, d, from a perfectly reflecting surface and a point far away. Note
that the curves have been truncated at the frequency where d = λ/4. These single frequency values are also applicable to
white noise in one-third-octave or octave-bands.
For an edge that lies along the x-axis, the mean-square sound pressure ratio at a point (y, z) is
p2 t
= 1 + j0 (2ky) + j0 (2kz) + j0 (2kdx )
2
p∞
t
(1.120)
where j0 (a) = sin (a)/a and dx2 = y 2 + z 2 .
For a corner positioned at the origin of the Cartesian coordinates the mean-square sound
pressure ratio at a point (x, y, z) is
p2 t
= 1 + j0 (2kx) + j0 (2ky) + j0 (2kz) + j0 (2kdx ) + j0 (2kdy ) + j0 (2kdz ) + j0 (2kd)
2
p∞
t
(1.121)
where dy2 = x 2 + z 2 , dz2 = x 2 + y 2 and d 2 = x 2 + y 2 + z 2 .
Figure 1.30 allows comparison of the sound pressure level difference for a surface, edge, and
a corner. To create this particular example, d is used to represent different distances from the
surface, edge, or corner: namely, the distance perpendicular to the surface along the x-axis,
the distance from the edge along the line y = z, and the distance from the corner along the
line x = y = z. At the boundary position where d = 0, the level differences are 3, 6, and 9 dB
for the surface, edge, and corner respectively. Image sources can be used to visualize this
finding. Figure 1.31 shows the actual source for a plane wave front near these boundaries
along with the image sources. For the surface, edge, and corner there are 1, 3, and 7 reflected
waves respectively; this gives the total number of sources as 2, 4, and 8 respectively. The
image sources have the same amplitude and phase as the actual source, therefore the sound
pressure from the actual source and the image sources is in phase at the boundary position
(d = 0). Hence the level differences correspond to 10 times the logarithm (base 10) of the total
number of sources.
49
Sound pressure level difference (dB)
S o u n d
I n s u l a t i o n
12
z
9
x
y
6
Surface
d
3
distance d along the x-axis
0
d
Edge
⫺3
distance d along line y ⫽ z
⫺6
d
⫺9
Corner
distance d along line x ⫽ y ⫽ z
⫺12
0
2
4
6
8
10
2k d
12
14
16
18
20
Figure 1.30
Sound pressure level difference between a point at a distance, d, from a surface, edge, and corner relative to a point far away.
Surface
Edge
Corner
Figure 1.31
Source ( ) and image sources ( ) for a plane wave front incident upon a surface, edge, and corner.
50
Chapter 1
The level differences due to interference patterns near edges and corners tend to be larger in
magnitude than with a surface and extend to greater distances. However, we will soon look at
calculating the total energy stored in a room where it is necessary to account for the energy
stored in the interference patterns of surfaces, edges, and corners. When the sound field in
the central zone of the room is reasonably diffuse it is found that the energy stored in edge
and corner zones is relatively small compared to the energy stored near room surfaces. This
is because the surfaces account for large areas in most rooms (Waterhouse, 1955).
1.2.7.1.2 Other boundary conditions
Up till now we have assumed that the room boundaries are perfectly reflecting and rigid, i.e.
the normal acoustic surface impedance is infinite. Therefore at the surface, the particle velocity
normal to these boundaries is zero. Many walls and floors in buildings have low-absorption
coefficients in the low-frequency range where interference patterns in rooms are important;
so the assumption that they are perfectly reflecting is reasonable. However, habitable rooms
almost always have fairly absorptive surfaces to provide suitable acoustics for the occupants
of the building so we need to consider the effect of absorption. In addition we know that real
room surfaces are not rigid because they are set into vibration by impinging sound waves. In
reality the sound waves that impinge upon the walls and floors cause them to vibrate; hence the
particle velocity at, and normal to the room surfaces must be the same as the surface vibration
of the wall or floor, and not zero. When we focus on sound transmission, the vibration of these
room surfaces becomes particularly important. Room surfaces range from a single sheet of
12.5 mm plasterboard on a timber frame, to a few hundred millimetres of solid concrete; all of
these surfaces have finite values for the acoustic surface impedance.
If we restrict our attention to sound waves impinging upon a surface at normal incidence, then
the effect of absorptive non-rigid surfaces can be assessed by using Eq. 1.113 to calculate
the mean-square pressure. This requires knowledge of the reflection coefficient which can be
calculated from the specific acoustic impedance using Eq. 1.74, and can be related to the
absorption coefficient using Eq. 1.75.
In order to assess the sound field near a surface it is useful to reference the mean-square
2
pressure from Eq. 1.113 to the mean-square pressure of the free-field incident wave, p+
t ,
where
2
p+
t =
2
p̂+
2
(1.122)
2
The sound field can then be shown using the sound pressure level difference, 10 lg (p2 t /p+
t ),
where 0 dB corresponds to the level of the free-field incident wave.
Figure 1.32 shows examples of the sound pressure level difference for normal incidence as
a function of 2kx in front of a surface at x = 0. Note that the distance, x, is in the negative
x-direction. These examples use a range of values for the specific acoustic impedance to
represent different surfaces. For a rigid surface the specific acoustic impedance is infinite.
Real values for the specific acoustic impedance are chosen to show a range of absorption
coefficients up to unity. The complex value, 1 + 8i, is used to represent a single sheet of
12.5 mm plasterboard (10.8 kg/m2 ) at 50 Hz; this plate has a low surface density and is used
to provide contrast to the assumption of a rigid surface. The complex value, 1 − 2i, can occur
at a single frequency when a porous material is placed in front of a thick heavy wall to provide
absorption in the room.
51
S o u n d
I n s u l a t i o n
Specific acoustic Absorption
impedance
coefficient
⬁ (Rigid surface)
18
3
1
1 ⫹ 8i
1 ⫺ 2i
⫺10
⫺9
⫺8
⫺7
⫺6
0
0.20
0.75
1
0.06
0.50
⫺5
⫺4
⫺3
⫺2
⫺1
Sound pressure level difference (dB)
9
6
3
0
⫺3
⫺6
⫺9
⫺12
⫺15
⫺18
⫺21
⫺24
⫺27
⫺30
⫺33
⫺36
⫺39
⫺42
⫺45
⫺48
⫺51
⫺54
⫺57
⫺60
0
2kx
Figure 1.32
Sound field in front of a surface formed by the incident and reflected waves for normal incidence. Surfaces with different
specific acoustic impedances are positioned at x = 0.
When the specific acoustic impedance is unity then the absorption coefficient is also unity and
the incident wave is completely absorbed at the surface. Therefore the sound field in front of
the surface only comprises the incident wave, and the sound pressure level difference is 0 dB
for all values of 2kx. In contrast, at the rigid surface there is a perfect reflection so that there
is pressure doubling, a level difference of 6 dB. As we move away from the surface, there are
sharp interference minima (troughs) where the incident and reflected waves are exactly out of
phase with each other. There are also interference maxima with peak values of 6 dB where
the waves are in phase with each other. The minima occur when 2kx = −(2n − 1)π where
n = 1, 2, 3, etc.
When the specific acoustic impedance has real or complex values, resulting in absorption
coefficients between 0 and 1, the depths of the minima and the height of the maxima are
significantly reduced in comparison to the rigid surface. Also, in comparison to surfaces with
real or infinite impedance, complex impedances can significantly change the value of 2kx at
which the minima and maxima occur. In the next section we will look at the modal sound field
by assuming perfectly reflecting and rigid boundaries; hence the features we have seen here
for real boundaries will be of relevance again.
1.2.7.2
Sound field associated with a single mode
Before we look at sound fields where there are many modes, it is instructive to look at the
sound field associated with an individual mode. We will use the box-shaped room with perfectly
reflecting and rigid boundaries and send a sinusoidal signal to a loudspeaker positioned in one
of the corners.
52
Chapter 1
For a sound source positioned at xs , ys , zs , the mean-square sound pressure level at a
receiver point x, y, z that is associated with mode, fp,q,r , is calculated using normal mode
theory (Morse and Ingard, 1968)
2
ωρ c 2 Q ψ
(x,
y,
z)ψ
(x
,
y
,
z
)
p,q,r s s s
0 0 rms p,q,r
2
(1.123)
pp,q,r (x, y, z, xs , ys , zs )t =
V
2
2
2
2
2
p,q,r 4ωp,q,r ζp,q,r + (ω − ωp,q,r )
where ω is the frequency of the sinusoidal signal, Qrms is the rms volume velocity of the source,
ζp,q,r is the damping constant, and p,q,r = 1/(εp εq εr ) for which εp , εq , and εr have already been
defined next to Eq. 1.99.
The damping constant, ζp,q,r , can be linked back to δm that was previously used to describe
the reverberant decay of an individual mode in Eq. 1.99. In terms of the damping constant, the
decay of the mean-square sound pressure for mode, fp,q,r , is
p2 (t) = p2 (0) exp(−2ζp,q,r t)
(1.124)
We can now describe the damping constant in terms of the reverberation time or loss factor. From Eq. 1.101 the damping constant is related to the reverberation time, Tp,q,r , for an
individual mode by
3 ln 10
(1.125)
ζp,q,r =
Tp,q,r
which is related to the loss factor for an individual mode, ηp,q,r , using
ηp,q,r =
6 ln 10
ζp,q,r
=
2πfTp,q,r
πf
(1.126)
From Eqs 1.51 and 1.52 the local mode shape (also called an eigenfunction), ψp,q,r , that
describes the sound pressure distribution in space for the receiver position is
qπy
r πz
pπx
(1.127)
cos
cos
ψp,q,r (x, y, z) = cos
Lx
Ly
Lz
and for the source position is
ψp,q,r (xs , ys , zs ) = cos
pπxs
Lx
cos
qπys
Ly
cos
r πzs
Lz
(1.128)
We will soon look at how different source positions affect the excitation of individual modes.
For the moment it is only necessary to ensure that we can excite any mode; hence the source
needs to be positioned at any one of the corners, for example at 0,0,0. When the source is
at a corner, |ψp,q,r (xs , ys , zs )| = 1 for any mode. This allows us to focus on the way that ψp,q,r
(x, y, z) affects the spatial distribution of the mean-square sound pressure.
The maximum value that |ψp,q,r (x, y, z)| can take is 1; hence for an individual mode, the
maximum mean-square sound pressure occurs at receiver positions where |ψp,q,r (x, y, z)| = 1.
These maximum values occur at positions referred to as anti-nodes and their position in the
room depends upon the individual mode, fp,q,r . However, when the receiver is positioned at
any of the eight corners of the box-shaped room, then |ψp,q,r (x, y, z)| = 1 for all modes. Hence,
the corner of a room is an ideal point to detect which modes have been excited.
For any individual mode in a box-shaped room, the sound field on any of the three orthogonal
planes forming the room is symmetrical about the lines perpendicular to the axes.
53
S o u n d
I n s u l a t i o n
Sound
pressure
level
(dB)
z
x
y
z
y
Figure 1.33
Illustration showing how the sound pressure levels from each plane in the box-shaped room is displayed on each threedimensional surface plot. In this example it is the yz plane midway along the x dimension, Lx . Each plane is shown with a
graduated shaded area so that corners with the darkest and the lightest shading can be used as reference points to identify
sound pressure levels at different positions on the plane. If more than one plane is shown, then all these planes have the
same sound pressure level distribution.
When at least one of the cosine terms in ψp,q,r (x, y, z) is 0, the sound pressure is also 0, so
there will be planes of zero pressure perpendicular to the x, y, or z-axes. In a box-shaped room
these are referred to as nodal planes; these exist where x = nLx /2p, y = nLy /2q and z = nLz /2r
for n = 1, 3, 5, etc. For any mode, fp,q,r , there will be p nodal planes perpendicular to the x-axis,
q nodal planes perpendicular to the y-axis, and r nodal planes perpendicular to the z-axis.
For rooms with volumes less than 30 m3 , there will only be one mode or a few modes in individual one-third-octave-bands between 50 and 100 Hz. These modes are usually the axial modes
f1,0,0 , f0,1,0 , f0,0,1 , f2,0,0 , f0,2,0 , the tangential modes f1,1,0 , f0,1,1 , f1,0,1 and the oblique mode f1,1,1 .
The graphs in this section can be interpreted with reference to Fig. 1.33. Examples of the sound
pressure level distribution are shown for axial modes f1,0,0 and f0,0,1 (Fig. 1.34), the tangential
mode, f1,1,0 (Fig. 1.35), and the oblique mode f1,1,1 (Fig. 1.36). To allow a practical interpretation
of the sound field, the decibel scale has been used. However the use of decibels is not ideal
because the mean-square pressure is zero on the nodal planes. In reality there will not be zero
mean-square sound pressure for two reasons. Firstly, there will always be some background
noise in the measurement, and secondly, not all real surfaces are perfectly reflecting and rigid.
For the latter reason the surfaces will absorb some of the incident sound and there will not be
perfect cancellation along nodal planes, i.e. there will not be pure standing waves. This was
previously seen in Section 1.2.7.1.2 when we looked at the sound field near room boundaries
where the specific acoustic impedance of the surfaces had complex or finite real values. The
maximum level for each mode has therefore been normalized to 0 dB, and we will assume that
the background noise level is 60 dB below this maximum level; so levels of −60 dB on these
graphs represent the nodal planes with zero mean-square pressure.
The main features of the modal sound field are the large spatial variations in the sound pressure
level. The highest levels occur at the room boundaries, with the nodal planes sited away from
these boundaries.
This model of the sound field gives us a basic insight into the sound field in a room. However by
assuming that the room boundaries are perfectly rigid we have not considered the interaction
between the sound in the room and the vibration of the walls and floors. In addition we have
not yet considered the situation where frequency bands ‘contain’ zero, one, or more modes.
54
Chapter 1
(a) f1,0,0
0
dB
⫺60
z
0
y
x
Ly/2
0
Lz/2
0
dB
dB
⫺60
⫺60
dB
Lx/2
⫺60
(b) f0,0,1
z
0
y
dB
x
0
dB
Lx/2
⫺60
0
⫺60
dB
Ly /2
⫺60
Lz/2
0
dB
⫺60
Figure 1.34
Sound pressure level distribution for axial modes: (a) f1,0,0 and (b) f0,0,1 .
z
y
0
0
dB
dB
⫺60
⫺60
x
4Lx/10
4Lx/10
0
dB
dB
⫺60
⫺60
⫺60
Lx/2
Lz/2
0
dB
4Ly/10 4Ly/10
0
0
dB
dB
⫺60
Ly/2
0
⫺60
Figure 1.35
Sound pressure level distribution for the tangential mode, f1,1,0 .
55
S o u n d
z
y
I n s u l a t i o n
0
0
0
dB
dB
dB
⫺60
⫺60
⫺60
0
4Lz/10
dB 4Lz/10
dB
⫺60
⫺60
x
0
4Lx /10
4Lx/10
dB
⫺60
Lx/2
4Ly/10 4Ly/10
Lz/2
0
0
0
dB
dB
dB
⫺60
⫺60
⫺60
Ly/2
0
Figure 1.36
Sound pressure level distribution for the oblique mode f1,1,1 .
1.2.7.3 Excitation of room modes
We need to consider the excitation of room modes in two different situations: one in a source
room where the airborne sound insulation is being measured by using a loudspeaker to
generate the sound field, and the other in a receiving room where the room surfaces radiate
sound into the receiving room, which excites the receiving room modes. The latter situation
primarily concerns the coupling between energy stored in walls, floors, or other spaces, and the
energy stored in the receiving room modes and is discussed in Chapter 4. At this point, we just
consider the effect of using a single loudspeaker at different positions in the source room, and
how those different positions affect the excitation of room modes in the low-frequency range.
For individual frequencies the mean-square sound pressure due to each room mode can be
calculated using Eq. 1.123 and summed to give the overall mean-square sound pressure. This
is similar to carrying out a swept-sine measurement with a constant amplitude signal from the
loudspeaker. In practice the airborne sound insulation is usually measured using broad-band
noise, but single frequencies gives a clearer understanding of the effect of different source
positions.
Three different source positions are assessed: one near a corner, another at the mid-point
along one wall, and another that is exactly in the centre of the room. In practice it is rarely
possible to put the loudspeaker exactly at the corner, although we can place it near the corner;
so we will assume that the acoustic centre of the loudspeaker is at 0.5,0.5,0.25 m which is
nearest to the corner at 0,0,0.
The receiver position is at Lx ,Ly ,Lz , which is in the corner opposite the loudspeaker. A microphone would not be placed in a corner for standard sound insulation measurements; it would
be positioned away from the room boundaries. However, we only want to assess which modes
have been excited so we do not need to know the absolute sound pressure levels. For this
reason the receiver position is at a point in the room where all modes have an anti-node (i.e.
in a corner).
The damping constants for the individual modes are calculated from Eqs 1.125 and 1.101 using
a frequency-independent value for the specific acoustic admittance (βa,s = 0.01).
Predicted curves for the sound pressure level are shown in Fig. 1.37 for the same 50 m3 boxshaped room that was used to look at the mode count in one-third-octave-bands. The peaks in
56
Chapter 1
the sound pressure level are due to the room modes that have been excited. On each curve, the
axial, tangential, and oblique modes have been plotted at their respective eigenfrequencies.
When there are relatively few modes and the modes are all well-separated we see that if a
mode frequency coincides with a peak in the sound pressure level curve, this indicates that this
particular mode has been excited; if it coincides with a trough in the curve, then it has not been
excited. If a mode frequency occurs on the curve between a trough and a peak, then either the
mode has not been excited, or it has been excited but it has such a low sound pressure level
that there is no discernible peak. This numerical experiment indicates the difficulty in using
physical experiments to find the mode count where a microphone and a sound source are
placed in a single position. The problem is that counting the peaks in the response is unlikely
to identify all of the modes; this equally applies to vibration measurements used to identify
structural modes of vibration.
When the source is near a corner there are many more peaks in the sound pressure level
curve than when the source is at the centre of the room; hence many more modes are excited
by a corner position than the central point. When the source is near a corner, all the modes
below 100 Hz have been excited and are clustered at or near the peaks in the sound pressure
level curve. In contrast, when the source is at the centre of the room, the first five modes have
not been excited at all, and there are clusters of modes that have not been excited near the
troughs of the sound pressure level curve. The modes that have not been excited have one
or more nodal planes that cut through the source position at the centre of the room (e.g. see
f1,1,1 in Fig. 1.37). Similarly, when the source is mid-way along one wall, the modes that have
not been excited also have one or more nodal planes that cut through the source position (e.g.
see f1,0,0 and f1,1,1 in Fig. 1.37).
For field airborne sound insulation measurements in non-diffuse sound fields it is necessary
to excite the majority of the modes in the source room. For this reason, loudspeaker positions
near the corners are used in box-shaped rooms as well as in other shapes of room. In addition
it is necessary to take average measurements from more than one source position. However,
it must also be ensured that the direct sound from the loudspeaker does not cause significant
excitation of the walls or floors compared to excitation by the reverberant sound field.
1.2.7.4 Diffuse and reverberant fields
The Schroeder cut-off frequency is sometimes used to estimate the lowest frequency above
which the sound field can be considered to be diffuse. This is the frequency at which the modal
overlap factor equals three; hence it identifies the lowest frequency above which the sound
energy is relatively uniform in the central zone of a room.
In practice, a diffuse field cannot be realized throughout the entire room volume due to the
fact that rooms are defined by the walls and floors that form the room boundaries. These
boundaries give rise to interference patterns close to the surfaces, edges, and corners; hence
the energy density is not uniform throughout the space. In addition these boundaries absorb
sound (to varying degrees) so that there must be a net power flow from the sound source to the
boundaries, whereas in a diffuse field the net power flow is zero. Under laboratory conditions
it is possible to achieve a suitable degree of diffusivity through careful room design, by using
diffusing elements in the room, and by defining measurement positions away from the room
boundaries. Hence close approximations to diffuse fields in the building acoustics frequency
range only tend to exist in the central zones of large reverberant chambers. Such chambers
are carefully designed and validated for the laboratory measurement of absorption or sound
57
S o u n d
I n s u l a t i o n
Box-shaped room
Lx = 4.64 m, Ly = 3.68 m, Lz = 2.92 m
V = 50 m3
1,0,0
0,1,0
Source at
0.5, 0.5, 0.25
(near a corner)
Sound pressure level (dB)
0,1,0
1,1,0
1,2,0
1,1,1
1,0,2
1,0,1
Source at
Lx/2, 0.5, 0.25
(mid-way along
one wall)
2,0,0
10 dB
0,1,1
1,0,0
0,2,0
2,0,1
Source at
Lx/2, Ly/2, Lz/2
(at the centre of
the room)
Axial modes
Tangential modes
1,0,1
1,1,1
0
20
40
60
80
Oblique modes
2,1,0
100
120
Frequency (Hz)
140
160
180
200
Figure 1.37
Excitation of room modes with three different source positions. Curves for the sound pressure level in the corner position
(Lx , Ly , Lz ) are shown along with the axial, tangential and oblique mode frequencies to assess which modes are, and which
modes are not excited by the source position. Note that the curves have been offset from each other; this allows the relative
levels along each individual curve to be assessed, but not the relative levels between different curves.
power levels. In laboratory measurements of airborne and impact sound insulation we also
come across close approximations to diffuse sound fields in the source and receiving rooms.
In general it is better to avoid referring to a diffuse sound field in a room because we must
add several caveats to any such statement. At frequencies above the lowest room mode it is
58
Chapter 1
simpler if we just refer to a reverberant field. This acknowledges the fact that the room response
varies over the building acoustics frequency range due to the existence of one, a few, several,
or many modes in the different frequency bands. It also serves as a reminder that interference
patterns exist at the room boundaries, and that it is only in the central zone of reverberant
rooms that sound fields resemble (to varying degrees) a diffuse field.
1.2.7.5 Energy density
The energy density, w, equals the energy per unit volume. For a sound field comprised of plane
waves, Eq. 1.19 describes the sound intensity, i.e. the energy that flows through a unit surface
area in unit time. Therefore, as a plane wave will travel a distance equal to c0 in unit time,
the energy density in any reverberant sound field comprised of plane waves (which includes
diffuse or modal sound fields) is
wr =
p2 t,s
I
=
c0
ρ0 c02
(1.129)
Hence the energy density is directly proportional to the temporal and spatial average meansquare sound pressure, p2 t,s .
1.2.7.5.1
Diffuse field
As any kind of signal can be described by using many impulses, the steady-state energy density
in a diffuse field can be derived by considering the impulse response of a room (Barron, 1973;
Kuttruff, 1979). We assume that a room contains a point source that generates an impulse at
time, t = 0. This point source generates spherical waves for which the intensity is inversely proportional to the square of the distance travelled (Eq. 1.26). At an arbitrary receiver position in the
room, the sound intensity is the sum of the intensity from the direct path and the many indirect
paths involving at least one reflection from a room boundary. For direct propagation from the
source to the receiver, we define the intensity at the receiver to be I0 , after it has travelled the
source–receiver distance, r0 . For each propagation path that involves at least one reflection, we
can consider an image source that generates an identical impulse to the actual source at t = 0.
At time, t, impulses from the actual source or an image source will have travelled a distance,
c0 t. For the image sources, a fraction of the sound intensity is absorbed upon each reflection
from the room boundaries; hence from Eq. 1.92 the intensity is attenuated by the factor
exp
c0 t
ln (1 − α)
dmfp
(1.130)
At an arbitrary receiver point, the intensity from an image source at time, t, is therefore given by
I0 r02
c0 t
ln
(1
−
exp
α)
(c0 t)2
dmfp
(1.131)
At time, t, we now need to determine how many impulses from image sources will arrive at
the receiver in a small time interval, δt. This is equivalent to finding the number of reflections
arriving during this time interval, and can be calculated from the temporal density of reflections
(Eq. 1.49), using
4πc03 t 2
δt
V
(1.132)
59
S o u n d
I n s u l a t i o n
Therefore the total intensity arriving at the receiver after a specific time, t1 , can be found by
integrating the intensity from all reflections according to
I=
∞
t1
I0 r02
4πc03 t 2
c0 t
exp
ln
(1
−
dt
α)
2
(c0 t)
dmfp
V
(1.133)
and the energy density can be found using
I
wr =
=
c0
∞
t1
c0 t
W
exp
ln (1 − α) dt
V
dmfp
(1.134)
where the sound power, W , for the spherical wave source is
W = 4πr02 I0
(1.135)
The choice of time, t1 , for the lower limit of the integral needs to consider the fact that the direct
sound does not arrive at the receiver until t = r0 /c0 , and that the exponential decay in a diffuse
field can only start after the first reflections arrive at the receiver (i.e. when t > r0 /c0 ). Therefore
if the lower limit of the integral is taken to be t1 = 0, the energy density will include the time
interval before the direct sound has arrived and the exponential decay has begun. Using t1 = 0
to estimate the energy density in a diffuse field, wr , gives the classical equation
wr =
4W
c0 A
(1.136)
where A = −ST ln (1 − α).
The time interval during which the direct sound travels to the receiver can be excluded by using
t1 = r0 /c0 to give an estimate of the energy density that is dependent upon r0 (Barron, 1973).
In large spaces such as concert halls, this dependence on source–receiver distance is often
important. However, for sound insulation in typical rooms it is not necessary (or practical) to
relate the energy density to specific source–receiver distances. On the basis that the exponential decay in a diffuse field can only start after the first reflections arrive at the receiver position;
it is necessary to find a lower limit for the integration that represents the average time taken
to travel from the source to the receiver when there is one reflection. This can be determined
by using the mean free path (Kuttruff, 1979; Vorländer, 1995). Although the mean free path is
the average distance travelled after leaving one boundary and striking the next boundary, it is
reasonable to assume that the average distance from either the source or the receiver to any
room boundary is approximately equal to half the mean free path. Therefore the mean free
path represents the average path length between the source and receiver when there is one
reflection. Taking the lower limit of the integral to be t1 = dmfp /c0 gives the energy density of
the diffuse sound field, wr , as
wr =
4W
(1 − α)
c0 A
(1.137)
where A = −ST ln (1 − α)
If there is significant air absorption then for each image source, the intensity is attenuated by
the factor
c0 t
exp
ln (1 − α) exp(−mc0 t)
(1.138)
dmfp
60
Chapter 1
and the resulting energy density is (Vorländer, 1995)
A
4W
exp −
wr =
c0 A
ST
(1.139)
where A = −ST ln (1 − α) + 4mV .
Equating Eq. 1.129 to any of the above equations for the diffuse field energy density (Eqs
1.136, 1.137, or 1.139) gives the basic relationship between reverberant sound pressure and
absorption area for a fixed sound power input into a room; namely, that the sound pressure in
a room is reduced by increasing the absorption area, and vice versa. In diffuse fields with an
exponential decay, the decay curve is a straight line over the full 60 dB range; hence there is a
simple and unambiguous relationship between the reverberation time and the absorption area.
Having seen that there is more than one equation for the energy density in a diffuse field,
we need to discuss the equations that are used in practice. Although Eq. 1.139 is the more
accurate equation, it is not always necessary to consider air absorption. In practice, Eq. 1.137
is usually adequate. In most rooms with volumes less than 200 m3 and reverberation times
less than 2 s at 20◦ C and 50% RH the error in neglecting air absorption in the calculation of the
diffuse field energy density is only greater than 1 dB in one-third-octave-bands above 3150 Hz.
In the measurement Standards for sound insulation, the classical equation (Eq. 1.136) is used
to derive equations that link the sound power to the reverberant sound pressure level in a room
(Section 3.5.1). This is reasonable in transmission suites (where reverberation times are at least
1 s) because negligible errors are incurred when using Eq. 1.136. However, for field sound
insulation measurements (where reverberation times in furnished rooms are approximately
0.5 s), consideration could be given to use of Eq. 1.137 to determine the apparent sound
reduction index (Vorländer, 1995).
1.2.7.5.2
Reverberant sound fields with non-exponential decays
In comparison with diffuse fields, it is more awkward to make a link between the sound power
radiated into a room and the reverberant sound pressure level for a non-diffuse field. For modal
sound fields, some insight can be gained by using normal mode theory and the specific acoustic
admittance of the room boundaries; however, this does not give a completely general solution
that corresponds to the practical situation (Jacobsen, 1982; Bodlund, 1980). For non-diffuse
fields, it is the reverberation time that provides a practical link between the sound power and
the reverberant sound pressure level. The difficulty lies in evaluating decay curves from nondiffuse fields because they are not straight lines across the entire 60 dB decay range. For this
reason it is necessary to identify which part of the decay curve should be used to calculate the
reverberation time.
From normal mode theory in Section 1.2.6.3.2 we have seen that it is only within the initial 20 or
30 dB of the decay curve that the majority of room modes play a role in determining the decay
curve of the frequency band. The late part of the decay is determined by a relatively small
number of modes with longer decay times. In the steady-state situation, energy is stored in all
modes; hence, it is appropriate to determine the reverberation time using an evaluation range
in which the majority of the room modes play a role in forming the decay curve. Furthermore, as
any signal can be represented by a train of impulses, we can describe the steady-state sound
pressure by the energetic sum of a train of impulse responses. For each impulse response, only
the initial 20 dB drop of its decay curve will determine the steady-state sound pressure level to
within 0.1 dB. Therefore T10 , T15 , or T20 should be used to determine the absorption area, rather
61
S o u n d
I n s u l a t i o n
than T30 or T60 . The energy density in reverberant sound fields with non-exponential decays
can then be estimated by using the equations for the diffuse field energy density (Eqs 1.136,
1.137, and 1.139).
1.2.7.6 Direct sound field
So far we have considered the sound field in the central zone of a room without considering the
sound field close to the sound source, i.e. the direct field. Most sound insulation measurements
use an omnidirectional loudspeaker and we need to consider the direct field near the sound
source in order to assess how far the microphone should be from the loudspeaker when we
want to measure the reverberant sound pressure level in the central zone of the room without
any strong influence from the direct field.
For an omnidirectional source that emits a sound power, W , and is positioned away from the
room boundaries, the energy density of the direct field, wd , at a distance, d, from the source is
wd =
W
4πc0 d 2
(1.140)
Figure 1.38 shows the energy density due to the direct and the reverberant fields in a 50 m3
room at distances up to 1 m from the sound source. The energy density due to the direct field
decreases by 6 dB every time the distance is doubled. The distance from the sound source
at which the energy density in the direct field (Eq. 1.140) equals the energy density in the
diffuse field (Eq. 1.137) is described as the reverberation distance, rrd . When α ≈ − ln (1 − α)
the reverberation distance can be estimated using
ST α
rrd ≈
(1.141)
16π
In rooms with volumes less than 150 m3 and reverberation times greater than 0.5 s, the reverberation distance is usually less than 1 m. To measure the reverberant sound pressure level the
preferred option is to position the microphone at distances slightly greater than the reverberation distance. However in most rooms a practical choice for the minimum distance between the
microphone and most loudspeakers is 1 m; this is commonly used for airborne sound insulation
measurements (ISO 140 Parts 3 & 4).
1.2.7.7
Decrease in sound pressure level with distance
There are two main types of room in which there can be a significant decrease in the sound
pressure level with distance from the source: (1) large rooms (often with volumes greater
than 200 m3 ) with absorbent surfaces and/or large scattering objects and (2) corridors or passageways, usually with highly absorbent ceilings. Computer models (usually based around a
geometrical ray approach) can be used to calculate the sound pressure level distribution, but
for corridors it is still possible to gain some insight using simpler models.
Long corridors, such as those in flats, offices, and schools, are usually broken up into smaller
lengths of corridor by fire doors. This typically results in sections of corridor that are less
than 30 m in length, and where the dimensions of the cross-section are between 1.5 and 5 m.
Noise control measures in the corridor often require absorptive ceilings; hence these elongated
spaces can show a significant decrease in the sound pressure level with distance. As the sound
propagates down the corridor, there are two types of internal damping that reduce the sound
62
Chapter 1
Direct field
Energy density (dB)
Reverberant field
V ⫽ 50 m3
ST ⫽ 83 m2
A ⫽ 10 m2
T ⫽ 0.8 s
6 dB
rrd
0
0.2
0.4
0.6
Distance, d (m)
0.8
1
Figure 1.38
Energy density due to the direct and reverberant fields in a room at distances up to 1 m from the sound source.
pressure level: air absorption (which will be considered as negligible) and absorption by the
corridor surfaces.
For a thorough overview of models for long enclosures the reader is referred to the book by Kang
(2002). A simple model can be based on a corridor of infinite length that is divided into a number
of very thin box-shaped sections of depth, dL (Redmore and Flockton, 1977). Effectively we are
considering a large number of two-dimensional sound fields that are coupled together along
the length of the corridor (see Fig. 1.39). The following derivation and the resulting equation
(Eq. 1.145) are different from that in Redmore and Flockton (1977). However, it gives the
same equation that was later determined empirically in scale model experiments of corridors
by Redmore (1982).
It is assumed that the energy density in each section of volume, dV , is uniform and that the
corridor surfaces have an average absorption coefficient, α. We arbitrarily choose a section at
x = 0 in this infinite corridor, and follow sound propagating in the positive x-direction. In each
two-dimensional sound field, sound energy impinges upon the corridor surfaces, c0 /dmfp times
every second; the mean free path for a two-dimensional field is given later in Eq. 1.185. The
power absorbed by the corridor surfaces is
Wabs = wdV
Ec0 Uα
c0 A
=
dmfp UdL
πLy Lz
(1.142)
where A = UdLα, and the perimeter of the corridor section, U = 2Ly + 2Lz .
The absorbed power is related to the loss factor by
Wabs = ωηE
(1.143)
63
S o u n d
I n s u l a t i o n
z
Ly
Lz
y
dL
x
Figure 1.39
Corridor divided into sections.
hence equating Eqs 1.142 and 1.143 gives the loss factor as
η=
c0 Uα
2π2 fLy Lz
(1.144)
After travelling a distance, d, down the infinitely long corridor, the sound energy is reduced by
the factor exp(−kηd), which gives the decrease in the sound pressure level in decibels as
Linf =
10 1 Uαd
ln 10 π Ly Lz
(1.145)
Equation 1.145 applies to an infinitely long corridor (i.e. without ends). In practice, sound will
be partially reflected and partially absorbed from the end(s) of the corridor. We now consider a
corridor that extends to infinity in the negative x-direction, but has a termination at x = D (e.g.
at the fire doors). It will be assumed that there are no interference effects between the incident
and reflected sound at the receiver. In addition, we will not consider how the sound is injected
into the corridor; this avoids consideration of the direct field from the sound source. To do this
it is assumed that the thin corridor section at x = 0 starts with uniform energy density, and that
sound propagates in the positive x-direction. The decrease in sound pressure level after the
sound has travelled a distance, d, down the corridor is (Redmore and Flockton, 1977)
L = −10 lg (10−Linf /10 + R10−Linf ((2D−d)/d)/10 )
where the surface at the end of the corridor has a reflection coefficient, R.
64
(1.146)
Chapter 1
24
Decrease in sound pressure level, ∆L (dB)
22
A: α ⫽ 0.3
20
18
16
A: α ⫽ 0.2
14
B: α ⫽ 0.3
12
10
B: α ⫽ 0.2
8
6
A: α ⫽ 0.1
4
2
B: α ⫽ 0.1
0
0
2
4
6
8
10
12 14 16 18
Distance, d (m)
20
22
24
26
28
30
Figure 1.40
Decrease in sound pressure level along a corridor. Two different cross-sections (Ly × Lz ) are shown: A (1.5 × 2.5 m) and
B (3 × 3 m). The corridor surfaces have average absorption coefficients of 0.1, 0.2, or 0.3. The surface that forms the end of
the corridor has a reflection coefficient, R = 0.95.
Figure 1.40 shows the calculated decrease in level along two different corridors with different
average absorption coefficients. The partial reflection at the end of the corridor causes the
initial decrease in sound pressure level per unit distance to be larger than it is towards the end
of the corridor.
1.2.7.8 Sound fields in frequency bands
So far we have focused on the sound field at single frequencies in box-shaped rooms that
have perfectly reflecting rigid boundaries. We usually measure sound insulation in one-thirdoctave or octave-bands in rooms with a wide variety of boundaries. These bands contain many
modes with different decay times and, when considering interaction with the room boundaries,
it becomes more complex to predict this sound field. To look at the superposition of modes that
occurs in real rooms we now look at some example measurements of sound pressure levels
in one-third-octave-bands.
1.2.7.8.1 Below the lowest mode frequency
Below the lowest mode frequency we expect the sound field to be homogeneous and uniform
throughout the space. This assumption is reasonable when there is no significant overlap from
the response of the lowest mode into frequency bands below the lowest mode frequency.
To gain an impression of the sound field we can look at sound pressure level measurements
in a 29 m3 source room, and an 18 m3 receiving room, both with timber-frame walls and floors
65
S o u n d
I n s u l a t i o n
(Hopkins and Turner, 2005). The lowest mode frequency is calculated to be 39 Hz for the
source room and 59 Hz for the receiving room. Therefore we will look at the 20 Hz one-thirdoctave-band because this is well-below the lowest mode in both rooms. A broad-band noise
source was used with measurement positions in a three-dimensional grid (including positions
at the room boundaries). To gain a visual impression of the sound field, sound pressure levels
from three different measurement planes are shown in Figs 1.41a and 1.41b for the source
and receiving rooms respectively. In the source room there is a peak in the sound pressure
level immediately next to the loudspeaker. Further away from the loudspeaker in the source
room, and throughout the receiving room, the spatial variation within a single measurement
plane can be as low as 2 dB or as high as 12 dB. Although the sound field is relatively uniform
in some measurement planes it will not always be homogeneous throughout the room volume.
The non-uniform sound field in the receiving room may be attributed to structural modes and
resonances of the walls and floors that occur below the lowest room mode. The mass–spring–
mass resonance frequency of the timber-frame separating wall is estimated to fall in the 31.5 Hz
one-third-octave-band.
For field measurements it is worth noting that both source and receiving rooms can have
quite large spatial variations in the sound pressure level below the lowest calculated mode
frequency.
1.2.7.8.2
Reverberant field: below the Schroeder cut-off frequency
In rooms with volumes less than 30 m3 the sound field in one-third-octave-bands below 100 Hz
is sometimes dominated by the response of a single mode. However, in the low-frequency
range the situation is usually more complex due to the influence of one, two, or three modes
that fall within a frequency band.
In most dwellings the height (Lz ) of a room is less than the width (Lx ) and the depth (Ly ).
Therefore the lowest mode frequency will be f1,0,0 or f0,1,0 , and when Lx = Ly , one of these
modes will usually be the only mode that falls exactly within the lower and upper limits of the
associated one-third-octave-band. Depending on the amount of damping and the bandwidth,
the response from one or more modes in adjacent bands can overlap into this band. As an
example we can look at the measured sound pressure level distribution in a 34 m3 receiving
room with masonry/concrete walls and floors for the 50 Hz one-third-octave-band; this has a
measured reverberation time of 1.2 s. For this room the Schroeder cut-off frequency is in the
500 Hz band. Broad-band excitation was applied in the source room, so it is representative
of the situation that is encountered in field sound insulation tests. The sound field is shown
in Fig. 1.41c and can generally be described as symmetrical in each plane. The first three
modes are f1,0,0 , f0,1,0 , and f1,1,0 for which the calculated mode frequencies (assuming rigid
walls) are 43, 47, and 64 Hz respectively. Although f0,1,0 is calculated to be the only mode that
falls exactly within the limits of the 50 Hz band, there is evidence of overlapping response from
one or both of the f1,0,0 and f1,1,0 modes. This is evident from the nodal planes along both the x
and y-axes where the sound pressure levels are lowest in the middle of both these axes, rather
than just the y-axis as would occur if the response was only due to f0,1,0 . We previously noted
that real walls and floors are not rigid and will dissipate energy; hence the sound pressure in
nodal planes will not be zero as implied by normal mode theory for sinusoidal excitation of a
single mode in a room with rigid boundaries. However, these nodal planes still cause a high
degree of spatial variation in the sound field. For the 50 Hz band in this example, there is a
difference of 28 dB between the lowest level in the central zone of the room and the highest
66
(a) Source room (29 m3), 20 Hz one-third-octave-band. The loudspeaker position is indicated by
z
z
z
3 dB
3 dB
x
x
3 dB
x
y
y
y
(b) Receiving room (18 m3), 20 Hz one-third-octave-band
z
z
x
y
3 dB
z
x
y
3 dB
x
y
3 dB
Figure 1.41
Measured sound pressure level distribution in rooms. For each figure, the same scale is used in each of the three plots to allow an assessment of the differences between the sound pressure
levels in the different planes; however, different scales are used for different rooms and different frequency bands.
(c) Receiving room (34 m3), 50 Hz one-third-octave-band
z
z
3 dB
x
z
3 dB
x
x
3 dB
y
y
y
(d) Receiving room (18 m3), 80 Hz one-third-octave-band
z
z
3 dB
x
y
Figure 1.41
(Continued)
z
x
y
3 dB
x
y
3 dB
(e) Receiving room (18 m3), 160 Hz one-third-octave-band
z
z
3 dB
x
z
x
y
x
3 dB
3 dB
y
y
(f) Receiving room (34 m3), 500 Hz one-third-octave-band
z
z
3 dB
x
y
Figure 1.41
(Continued)
z
3 dB
x
y
x
y
3 dB
S o u n d
I n s u l a t i o n
level that is ≈0.5 m from the room boundaries. For sound insulation measurements with broadband noise sources, measured data suggests that this difference will usually be between 17
and 28 dB for typical rooms in the low-frequency range (Hopkins and Turner, 2005; Simmons,
1996); background noise will always limit the lowest level that is measurable in the nodal
planes.
The highest sound pressure levels exist in corners and near wall/floor surfaces, with low levels
near the centre of the room. This highlights an important issue for spatial average sound
pressure level measurements in small rooms. Field sound insulation measurement procedures
usually require that the microphone is positioned at a minimum distance of 0.5 m from the
boundaries with guidance that this minimum distance should be increased to 1.2 m below
100 Hz (ISO 140 Part 4). In small rooms this means that the microphone is positioned in
the central zone of the room where the sound pressure level is lowest. Therefore the spatial
average levels are not representative of either the room average sound pressure level or the
level perceived by room occupants who often sit and sleep near the room boundaries.
For a room height, Lz , which is between 2.1 and 2.4 m, the f0,0,1 mode will fall within the
63 Hz or 80 Hz one-third-octave-band. This mode gives rise to low sound pressure levels on
the z = Lz /2 measurement plane (the plane that lies in the middle of the z-axis) compared to the
levels on the z = 0 and the z = Lz measurement planes. Figure 1.41d shows the data for the
80 Hz one-third-octave-band in an 18 m3 receiving room with timber frame walls and floors with
a measured reverberation time of 0.5 s. For this room the Schroeder cut-off frequency is in the
315 Hz band. We might expect the time-averaged sound pressure level in each measurement
plane for a box-shaped room to be symmetrical in frequency bands containing the first few
modes. However, the spatial variation is asymmetric; this usually occurs when one or more of
the room dimensions are equal to at least one wavelength. In most rooms there are recessed
windows or lobby areas associated with the door, i.e. smaller volumes connected to the main
rectangular space. As the room shape and the surface impedance of the room surfaces become
increasingly irregular with scattering objects in the room, the local mode shapes effectively
become hybrid mode shapes which do not have symmetrical sound pressure fields. Above the
first few modes there is usually a marked degree of asymmetry and complexity in the sound field.
An example is shown in Fig. 1.41e for the 160 Hz one-third-octave-band in the 18 m3 receiving
room of timber-frame construction. Asymmetry not only occurs in the receiving room, but also in
the source room, although this is partly due to higher levels in the direct field of the loudspeaker.
1.2.7.8.3
Reverberant field: at and above the Schroeder cut-off frequency
At and above the Schroeder cut-off frequency the sound field becomes increasingly uniform in
the centre of the room. However, at positions that are very close to the room boundaries there
are still higher sound pressure levels due to the interference patterns.
An example is shown in Fig. 1.41f for the 500 Hz one-third-octave-band in a 34 m3 receiving
room with masonry/concrete walls and floors. This frequency band contains the Schroeder cutoff frequency. Compared to the average level in the centre of the room, the average measured
levels at the wall surfaces are ≈3 dB higher, the edges are ≈6 dB higher, and the corners are
≈8 dB higher; these correspond to the predicted values 3, 6, and 9 dB that were discussed
earlier in Section 1.2.7.1.1. Higher sound pressure levels did not occur at a few grid points near
the room surface with a recessed window in the external wall because the microphone was
then ≈200 mm further away from the room boundary. This feature can be seen in Fig. 1.41f
where the window position is indicated on the diagram.
70
Chapter 1
Above the Schroeder cut-off frequency the spatial variation usually decreases significantly so
it is more useful to look at the standard deviation rather than plots of the spatial distribution of
the sound pressure level.
1.2.7.9 Statistical description of the spatial variation
When measuring and predicting the sound pressure level in rooms, it is almost always the
spatial average value that is required. Hence we need to be able to quantify the spatial variation
of the sound pressure level in terms of the normalized variance of the mean-square sound
pressure, and the standard deviation of the sound pressure level in decibels. The normalized
standard deviation, ε, is the ratio of the standard deviation to the mean, which is squared to
give the normalized variance, ε2 .
To determine the standard deviation and confidence intervals, it is necessary to know the probability distribution (probability density function) for the mean-square pressure, or identify one that
gives a reasonable representation of the actual distribution. The standard deviation depends
upon the type of excitation. For sound insulation we almost always use broad-band noise and
measure in frequency bands; although sound insulation against pure tones is occasionally of
interest with environmental noise sources. For frequency band measurements in rooms containing a single omnidirectional sound source emitting broad-band noise, estimates of the standard deviation can be found in the same way as for sound power measurements in a reverberant
chamber (Schroeder, 1969; Lubman, 1974). For airborne sound insulation measurements we
carry out spatial sampling of the sound pressure in the room that contains the loudspeaker,
the source room. We will assume that the sound pressure is sampled at stationary microphone
positions located at random points in the room; these positions are away from the room boundaries and at positions where the direct field from the source is insignificant. In this situation, the
spatial variation of the mean-square pressure is represented by a gamma probability distribution for either modal or diffuse sound fields (Bodlund, 1976; Lubman, 1968; Schroeder, 1969;
Waterhouse, 1968). This gamma distribution is asymmetric, right-skewed and is bounded at
the lower end of the distribution by the minimum possible value for the mean-square pressure.
In practice it is necessary to consider temporal as well as spatial averaging. Here it is assumed
that the uncertainty due to time-averaging of the random noise signal at each position is
negligible; for further discussion of temporal averaging in measurements see Section 3.3.3.
The valid frequency ranges for the variance and standard deviation formulae in this section
are defined in terms of the Schroeder cut-off frequency. At frequencies between 0.2fS and
0.5fS the normalized variance of the mean-square sound pressure can be estimated from
(Lubman, 1974)
N −1
(1.147)
ε2 (p2 ) = 1 +
π
where N is the mode count in the frequency band (Section 1.2.5.2).
The corresponding standard deviation of the sound pressure level in decibels, σdB , can be
estimated from Eq. 1.147 by taking account of the gamma distribution using (Craik, 1990)
σdB ≈
4.34
−0.22 + 1 +
(1.148)
N
π
Using the statistical modal density to calculate the mode count for use in Eqs 1.147 and 1.148
gives a non-integer mode count compared to the integer value determined from the individual
71
S o u n d
I n s u l a t i o n
mode frequencies by counting the number of modes that fall within each frequency band. The
difference between the two methods is rarely large but the statistical approach is more robust. It
accounts for the fact that most rooms are not perfect box-shaped rooms with rigid boundaries;
in practice, a modal response will often occur in a frequency band that is adjacent to the band
in which it was predicted to lie. It also avoids arbitrary decisions when an individual mode is
calculated to lie very close to the boundary between two adjacent frequency bands. By using
a statistical approach at frequencies with such low modal overlap we must expect the actual
standard deviation in decibels to fluctuate about the smooth curve predicted by Eq. 1.148.
If the lower limit of 0.2fS does not include the lowest frequency band of interest, then a reasonable estimate can still be calculated with Eq. 1.148 when the limit is lowered to that of the
frequency band containing the lowest mode frequency (Hopkins and Turner, 2005). As noted
by Lubman, this equation takes no account of other important factors: modal damping (reverberation time); combinations of small numbers of different room modes (axial, tangential, or
oblique); and the degree to which the modes are excited. Despite these omissions, it generally
gives estimates within ±1 dB of measured values.
In a diffuse field at frequencies above fS , the normalized variance can be calculated from the
bandwidth, B, and the room reverberation time, T , according to (Schroeder, 1969)
ε2 (p2 ) =
1
1 + 0.145BT
(1.149)
for which the corresponding standard deviation in decibels is (Schroeder, 1969)
5.57
σdB = √
1 + 0.238BT
(1.150)
In most rooms, Eqs 1.149 and 1.150 will give reasonable estimates at and above 0.5fS (rather
than above fS ); this allows continuity from Eqs 1.147 and 1.148 across the entire building
acoustics frequency range. Whilst these equations are normally sufficient, it is sometimes
necessary to take account of the direct sound field from the omnidirectional sound source. In
this situation, the normalized variance above 0.5fS can be calculated from (Michelsen, 1982)
⎡
⎤
√ 3
A
ST A ⎥
1
16π
⎢
+⎣
(1.151)
ε2 (p2 ) =
⎦
1 + 0.145BT
1602 dmin
V
where dmin is the minimum distance between the microphone and the sound source.
For airborne sound insulation measurements in both the laboratory and the field, dmin = 1 m
(ISO 140 Parts 3 & 4), and the difference between Eqs 1.149 and 1.151 can become significant
in the high-frequency range for low reverberation times and/or large rooms.
For Eq. 1.151, the corresponding standard deviation in decibels is (Michelsen, 1982)
σdB ≈ 4.34 ε2 (p2 )
(1.152)
Calculated standard deviations over the building acoustics frequency range (using Eqs 1.148,
1.151, and 1.152) are shown in Fig. 1.42 for a 50 m3 room with different reverberation times
and dmin = 1 m. Above 100 Hz the standard deviation increases as the reverberation time
decreases. Note that these standard deviations are for one-third-octave-bands and that octavebands will have lower values; hence octave-bands are sometimes used to reduce the required
number of microphone positions in a room. As an aside it is worth noting that the standard
72
Chapter 1
6
Box-shaped room
Lx ⫽ 5 m, Ly ⫽ 4 m, Lz ⫽ 2.5 m
V ⫽ 50 m3
Standard deviation, σdB (dB)
5
Reverberation time (s)
0.5
4
1.0
3
1.5
2
1
0
50
80
125
200
315
500
800
1250 2000
One-third-octave-band centre frequency (Hz)
3150
5000
Figure 1.42
Predicted standard deviation for the spatial variation of the sound pressure level for a 50 m3 source room with different
reverberation times.
deviation for a pure tone in a diffuse sound field is 5.6 dB (Schroeder, 1969); this is much
higher than the standard deviations that typically occur with broad-band noise measured in
one-third-octave or octave-bands over the building acoustics frequency range.
Describing the spatial variation in the receiving room is of equal importance to the source room.
However, the situation is more complex for three main reasons.
Firstly, the sound transmitted into the receiving room is not always broad-band in nature. It
may contain peaks in the sound pressure level at single frequencies, for example at the critical
frequencies of walls/floors/windows, or the mass–spring resonances of wall linings. In reality
this may only occur with a few types of homogeneous isotropic building elements because the
majority of building elements are constructed from small components (e.g. bricks forming a
brick wall) such that there will be spatial variation in the dynamic properties of the element due
to both workmanship and the material properties. This makes it less likely that there will be a
well-defined pure tone for a critical frequency or mass–spring resonance.
Secondly, we are no longer dealing with a single point source; one or more room surfaces are
acting as the sound sources. In the laboratory we assume that all the sound is radiated into
the receiving room by the test element, whereas in the field, it is radiated by both separating
and flanking elements. Michelsen (1982) and Olesen (1992) have investigated the standard
deviation of sound pressure levels in the source and receiving rooms for sound insulation
measurements in both the laboratory and the field. Radiating surfaces in the receiving room
can be represented as an equivalent number of uncorrelated point sources, hence the larger
the surface, the larger the number of point sources. The implication for the standard deviation
in a receiving room is that it should be lower than the source room because of the increased
number of uncorrelated point sources. Measured data does not confirm that lower values
always occur in practice (Michelsen, 1982; Olesen, 1992; Hopkins and Turner, 2005).
73
S o u n d
I n s u l a t i o n
6
Measured
Standard deviation, σdB (dB)
5
Predicted
4
3
Number of measurement points: 181
Minimum distance from room boundaries: 0.5 m
Minimum distance from sound source: 1 m
Schroeder cut-off frequency: 315 Hz band
Lowest eigenfrequency: 39 Hz
Average reverberation time (above 31.5 Hz band): 0.5 s
2
1
0
20
31.5
50
80
125
200
One-third-octave-band centre frequency (Hz)
315
500
Figure 1.43
Comparison of measured and predicted standard deviations for the spatial variation of the sound pressure level in a 29 m3
source room. Measured data are reproduced with permission from Hopkins and Turner (2005).
Thirdly, the gamma probability distribution may not be a reasonable representation of the
actual distribution for mean-square sound pressure in a receiving room. If we consider the
interaction between all the radiating surfaces and the space it is clear that the mean-square
sound pressure at any point in the room is determined by a large number of variables. By
assuming that these are independent random variables, the sound pressure level will be the
sum of a large number of random quantities. The central limit theorem can therefore be used
to infer that the spatial variation of the mean-square sound pressure will have a log-normal
probability distribution, and the sound pressure level in decibels will have a normal (Gaussian)
probability distribution (Lyon and DeJong, 1995).
Despite these three complexities, empirical evidence suggests that reasonable estimates for
receiving rooms can be found by using the same equations as for source rooms.
Figures 1.43 and 1.44 show measured and predicted standard deviations for a 29 m3 source
room and a 34 m3 receiving room (Hopkins and Turner, 2005). The microphone positions are
at least 0.5 m from the room boundaries and at least 1 m from the sound source. Generally,
there is good agreement between the measurements and the calculated values. The largest
discrepancies tend to occur below 0.5fS ; in practice, measured standard deviations will rarely
be greater than 6 dB for typical rooms within this frequency range. The equations discussed
above are only valid for sound fields above the lowest mode frequency. However, the measurements in the source room allow us to see the trend for the standard deviation below the
lowest mode frequency in the 20 to 31.5 Hz frequency bands. In this range we assume that
the sound field is uniform. Below the lowest mode frequency the standard deviation rapidly
decreases due to the fading influence of the modes on the sound field. In practice, it is unlikely
that the sound field in both source and receiving rooms can be considered as homogeneous
and uniform in the first few frequency bands below the band that contains the lowest mode
frequency.
74
Chapter 1
6
Measured
Standard deviation, sdB (dB)
5
Predicted
4
Number of measurement points: 105
Minimum distance from room boundaries: 0.5 m
Schroeder cut-off frequency: 500 Hz band
Lowest eigenfrequency: 43 Hz
Average reverberation time: 1.8 s
3
2
1
0
50
80
125
200
315
500
800 1250 2000
One-third-octave-band centre frequency (Hz)
3150
5000
Figure 1.44
Comparison of measured and predicted standard deviations for the spatial variation of the sound pressure level in a 34 m3
receiving room. Measured data are reproduced with permission from Hopkins and Turner (2005).
1.2.8 Energy
Although we are ultimately interested in the temporal and spatial average sound pressure
level in rooms, sound transmission involves energy flow between spaces and structures. This
makes it convenient to work with a single variable, energy; so we need to know the relationship
between the temporal and spatial average values of sound pressure and energy in a room.
The energy in a room can be derived with two different approaches, one using sound pressure,
and the other using sound particle velocity. The latter approach is used to describe the energy
stored in structures such as plates and beams; hence it is included here to highlight the link
between the way we deal with sound in spaces and the vibration of structures.
In both diffuse and non-diffuse fields, we can assume that the sound field is comprised of plane
waves and calculate the sound energy using the plane wave intensity described by Eq. 1.19.
The plane wave intensity quantifies the energy travelling through an imaginary surface of unit
area in 1 s, where this surface lies perpendicular to the direction of wave propagation. The
group velocity, cg , is defined as the velocity at which wave energy propagates, which for
longitudinal waves in air is the same as the phase velocity, c0 . Therefore the energy density
in a reverberant field, wr , that describes the energy in a unit volume is
wr =
I
c0
(1.153)
p2 t,s V
ρ0 c02
(1.154)
Hence the energy stored in volume, V , is
E = wr V =
where p2 t,s is the temporal and spatial average mean-square sound pressure.
75
S o u n d
I n s u l a t i o n
An alternative way of deriving the room energy is from the product of the mass of air within
the room and the temporal and spatial average mean-square sound particle velocity, u 2 t,s ,
where the latter can be found from the characteristic impedance of air (Eq. 1.18). This gives
E = mu 2 t,s = ρ0 V
p2 t,s
p2 t,s V
=
ρ02 c02
ρ0 c02
(1.155)
1.2.8.1 Energy density near room boundaries: Waterhouse correction
When calculating the sound energy stored in a reverberant room we need to consider the fact
that energy density is not uniformly distributed throughout the space. Near the room boundaries
the phase relationships between sound waves impinging upon a point are no longer random.
This causes interference patterns and an increase in energy density close to the boundaries
(Section 1.2.7.1).
To determine the total sound energy stored in a reverberant room from the energy calculated with Eq. 1.154 (using the spatial average sound pressure measured in the central zone
of the room) the energy is multiplied by the following frequency-dependent correction term
(Waterhouse, 1955)
1+
ST λ
8V
(1.156)
where ST is the total surface area of the room.
This term is widely referred to as the Waterhouse correction which is usually more convenient
to use in decibels,
ST λ
CW = 10 lg 1 +
(1.157)
8V
In the derivation of the correction term it is assumed that the room surfaces are perfectly
reflecting, which is often a reasonable assumption in the low-frequency range where the term
is most important. It is also assumed that sound waves are incident from all directions upon
a reflecting surface which has dimensions that are large compared to the wavelength. This
will not be a valid assumption where the sound field in the central zone of the room cannot
be classified as reasonably diffuse, and the room dimensions are small. Another important
assumption is that the energy stored in edge and corner zones is relatively small compared to
the energy stored near room surfaces. This is a reasonable assumption in many shapes and
sizes of room when the walls and floors have large surface areas. However, it is not necessarily
appropriate for modal sound fields in the low-frequency range where there are relatively few
oblique modes compared to axial and tangential modes. In this situation, numerical calculations
indicate that the Waterhouse correction sometimes appears to give accurate results because
it overestimates the energy stored near room surfaces (Agerkvist and Jacobsen, 1993). This
compensates for the fact that the energy stored in edge and corner zones is not included in
the correction term.
Example values for the Waterhouse correction are shown in Fig. 1.45 for box-shaped rooms
with volumes in the range 50 to 200 m3 . For these rooms the correction term is greater than 1 dB
in the low-frequency range which is often below the Schroeder cut-off frequency. This means
that significant values for the correction term tend to occur at frequencies where the sound field
in the central zone of the room is not a close approximation to a diffuse field. However, as a
rule-of-thumb for the building acoustics frequency range, the Waterhouse correction term tends
76
Chapter 1
4
Waterhouse correction, CW (dB)
Room volume (m3)
50
3
100
150
200
2
x : y : z ratio for each
box-shaped room is 41/3 : 21/3 : 1
1
0
50
80
125
200
315
500
800
1250 2000
One-third-octave-band centre frequency (Hz)
3150
5000
Figure 1.45
Waterhouse correction, CW , for different room volumes.
to give a reasonable estimate for most empty box-shaped rooms with a minimum volume of
50 m3 . This assumes that the sound pressure level in the central zone is adequately sampled.
When calculating the sound power radiated into a room from sound pressure measurements
made in the central zone of a reverberant room, the Waterhouse correction in decibels should
be added to the sound pressure level in decibels. However, we do not need to account for
interference patterns at the room boundaries when we calculate the diffuse field intensity that
is incident upon a surface from the diffuse field sound pressure level. Hence there are some
situations where we need to use the Waterhouse correction and some where we don’t. Specific
applications of the Waterhouse correction that apply to the measurement of sound insulation
are noted in Chapter 3.
1.3 Cavities
Cavities exist in many different parts of a building, for example: ceiling voids, roof voids,
between the joists in timber floors, in thermal glazing units, within cavity walls, and behind wall
linings. They can play an important role in sound transmission because vibration is not only
transmitted via structural connections between the plates that form a cavity, but also by the
sound field in the cavity itself.
As with rooms we will retain use of the convenient box-shaped space but for cavities we will
use Lz as the smallest dimension, the cavity depth (see Fig. 1.46).
1.3.1
Sound in gases
Almost all cavities in buildings are filled with air, so Eq. 1.1 for the speed of sound in rooms
is also applicable to cavities. However, cavities such as those in insulating glass units are
77
S o u n d
I n s u l a t i o n
z
Ly
Lz
y
Lx
x
Figure 1.46
Box-shaped cavity.
sometimes filled with other gases. Therefore a more general approach to calculate the phase
velocity, c, for any ideal gas is given by
1
V
γPV
γR(T + 273.15)
c=
=
=
=
(1.158)
κρ
κnM
nM
M
where κ is the gas compressibility (adiabatic), ρ is the gas density, V is the volume occupied
by n moles of a gas, M is the molar mass of the gas (kg/mol), γ is the ratio of specific heats
at constant pressure and constant volume which is 1.67 for monatomic gases such as helium,
1.41 for diatomic gases such as air, and 1.33 for polyatomic gases, P is the static pressure
which is 1.013 × 105 Pa for air at atmospheric pressure, R is the universal gas constant which
is 8.314 J/mol.K, and T is the temperature in ◦ C.
Properties of gases that are components of air or gases that are sometimes used in insulating
glass units are listed in the Appendix, Table A1.
1.3.2
Sound in porous materials
Cavities in walls and floors are sometimes partly filled or fully filled with porous materials to
absorb sound energy and provide other benefits such as thermal insulation. Porous materials
are also used around the perimeter of cavities to absorb sound and/or to control the spread of
fire. Some examples are shown in Fig. 1.47.
Porous materials essentially consist of a skeletal frame (which could be formed from fibres,
granules, or a polymer, etc.) that is surrounded by air. A wide range of porous materials are
used in buildings, with a range of frames (e.g. mineral wool, polystyrene balls, open-cell foam,
masonry blocks). Sound transmission through porous materials takes place due to airborne
propagation through the pores and structure-borne propagation via the frame. However, there
are varying degrees of coupling between these types of propagation, and they cannot simply
be assumed to occur independently of each other. For this reason, sound propagation through
78
Chapter 1
Lightweight wall (fully filled)
Secondary glazing
(porous reveal lining)
Timber joist
floor cavity
(partly filled)
Figure 1.47
Examples of porous materials used in cavities.
porous materials is considerably more complex than in air; the subject is only touched upon
here to introduce basic concepts and parameters that are needed in other chapters. For a
thorough review of different models used to describe sound in porous materials, the reader is
referred to Allard (1993).
1.3.2.1 Characterizing porous materials
Two simple parameters that can be used to describe the properties of porous materials are
the porosity and the airflow resistance. For a more complete description of the material, other
parameters such as the structure factor, shape factor, and tortuosity can be used to describe
aspects relating to the propagation path through the pores. However, these parameters are
rarely available, more awkward to measure, and are used in more complex models than will
be looked at here.
1.3.2.1.1 Porosity
For porous materials, the porosity, φ, is defined as
φ=
Vair
Vbulk
(1.159)
where Vair is the volume of air within the material and Vbulk is the bulk volume (i.e. total volume)
of the material.
For porous materials used in buildings, the porosity is usually in the range, 0.90 < φ < 0.99.
For mineral wool it is typically 0.95 < φ < 0.99. Mineral wool (i.e. glass or rock fibre) is usually
79
S o u n d
I n s u l a t i o n
made of solid fibres, hence if the material that binds these fibres together has negligible mass,
the porosity can be estimated using
φ =1−
ρbulk
ρfibre
(1.160)
where ρbulk is the bulk density of the material and ρfibre is the density of the fibre.
1.3.2.1.2
Airflow resistance
Sound absorption by, and sound transmission through porous materials is partly described by
their ability to resist airflow. This is quantified by the following parameters: airflow resistance,
specific airflow resistance, and airflow resistivity.
The airflow resistance, R (Pa.s/m3 ) is defined as
R=
p
qv
(1.161)
where p is the air pressure difference (referred to as differential pressure) across a layer
of porous material with respect to the atmosphere (Pa), and qv is the volumetric airflow rate
passing through the layer (m3 /s). The volumetric airflow rate is
qv = uS
(1.162)
where u is the linear airflow velocity (m/s) and S is the cross-sectional area of the porous
material perpendicular to the direction of airflow (m2 ).
The specific airflow resistance, Rs (Pa.s/m) applies to a specific thickness of a porous
material; hence it is an appropriate specification parameter for both homogeneous and nonhomogeneous materials as well as materials with a porous surface coating or perforated surface
layer.
Rs = RS
(1.163)
The airflow resistivity, r (Pa.s/m2 ) is the specific airflow resistance per unit thickness, and is
only appropriate as a specification parameter for homogeneous materials.
r=
Sp
RS
Rs
=
=
dqv
d
d
(1.164)
where d is the thickness of the layer of porous material in the direction of airflow (m).
NB: Specific airflow resistance and airflow resistivity are sometimes quoted in Rayls and
Rayls/m respectively. The Rayl is used as a unit for the ratio of sound pressure to particle
velocity and is equivalent to Pa.s/m.
For fibrous materials the airflow resistance depends upon the direction of airflow through the
material. These materials are usually supplied and used in rectangular sheets, either cut from
slabs or from a roll, hence the airflow can be measured in two directions as shown in Fig. 1.48:
(1) in the plane of the sheet, the lateral airflow and (2) perpendicular to the plane of the sheet,
the longitudinal airflow. In the literature it is usually measurements of the longitudinal airflow that
are quoted (e.g. Bies and Hansen, 1980). In rooms or cavities where sheets of material are used
to cover a surface it is the longitudinal direction that is needed to calculate the sound absorption
coefficient for the surface. However, narrow cavities are sometimes separated by sheets of
80
Chapter 1
Longitudinal direction
Lateral direction
Figure 1.48
Airflow resistivity of a sheet of porous material – definition of lateral and longitudinal directions.
fibrous materials that form a junction between the different cavities. Depending on the orientation of these sheets it is either the lateral or longitudinal direction that is needed to calculate
the absorption coefficient for the cavity boundary or sound transmission between cavities.
1.3.2.1.3 Fibrous materials
Fibrous materials are commonly used in cavities of walls and floors. The airflow resistance
of fibrous materials is due to friction between the fibres and the air particles moving between
the fibres, hence it can depend upon: size of fibres, shape/type of fibres (e.g. crimped, hollow), density of fibres, number of fibres per unit volume, and fibre orientation/distribution (e.g.
random, stratified/layered, stratified with higher fibre density near the surface of the sheet).
Mineral wool is anisotropic as the fibres tend to lie in planes that are parallel to the plane of
the sheet; the orientation of the fibres within each plane being random. Therefore the airflow
resistivity in the lateral direction is significantly lower than in the longitudinal direction.
For mineral wool (i.e. glass or rock wool) empirical relationships can be found between airflow
resistance and bulk density according to (Bies, 1988; Nichols, 1947)
1+k
r=
k1 ρbulk2
2
dfibre
(1.165)
where k1 is a constant for a material that is manufactured in a particular way, k2 is a constant
that depends upon fibre orientation, and dfibre is the fibre diameter (microns).
For one type of mineral wool with a known average fibre diameter, the constants k1 and k2 can
be found from measured airflow resistivity data for a range of bulk densities. By plotting lg(r )
against lg(ρbulk ), the data points should cluster along straight lines, and linear regression can
be used to determine k1 and k2 . An example is shown in Fig. 1.49 for the lateral and longitudinal airflow resistivity of rock wool (random fibre orientation, average dfibre = 4.75 µm, average
ρfibre = 2600 kg/m3 , porosity range was 0.94 (highest bulk density) ≤ φ ≤ 0.99 (lowest bulk density), two different UK manufacturers). For the lateral airflow resistivity, k1 = 353, k2 = 0.63 over
the bulk density range, 31 ≤ ρbulk ≤ 155 kg/m3 . For the longitudinal airflow resistivity, k1 = 780,
k2 = 0.59 over the bulk density range, 38 ≤ ρbulk ≤ 162 kg/m3 .
81
S o u n d
I n s u l a t i o n
200 000
R2 ⫽ 0.95
Longitudinal
Lateral
r (Pa.s/m2)
100 000
10 000
R2 ⫽ 0.93
1000
10
100
200
rbulk (kg/m3)
Figure 1.49
Measured airflow resistivity (lateral and longitudinal directions) for rock wool. Individual measurements are shown along with
regression lines. Measured data from Hopkins are reproduced with permission from ODPM and BRE.
Measured airflow resistivities and empirical relationships for other porous materials can be
found from Bies and Hansen (1980), Mechel and Vér (1992), and Mechel (1995). To cover
the full density range for a material it may be necessary to have more than one empirical
relationship, this can occur with fibrous materials that can be produced in a wide range of
fibre diameters. For materials such as glass wool, the combination of different manufacturing
processes and different fibre diameters can lead to empirical relationships that are specific to
one manufacturer and/or density range (Bies, 1988).
To determine empirical relationships for materials other than mineral wool, the form of Eq. 1.165
may not be appropriate. For example, with polyester fibre materials it has been shown that
better correlation can be found between the airflow resistivity and the number of fibres per unit
volume (Narang, 1995).
1.3.2.2 Propagation theory for an equivalent gas
General theory for sound propagation in a fluid-saturated porous elastic material requires consideration of two longitudinal waves and one shear wave (Biot, 1956). Modelling these three
waves requires knowledge of the fluid density, frame density, porosity, airflow resistivity, tortuosity, complex shear modulus, and Poisson’s ratio. In buildings we are usually interested in
air-saturated porous materials, rather than liquid-saturated. This simplifies matters because
with gases it can often be assumed that the skeletal frame is not elastic, and is sufficiently rigid
that it does not move. This allows use of simpler sound propagation models.
82
Chapter 1
Air Porous material Air
r0
r0
k
k
c0
c0
≡
Air
Equivalent gas
Air
r0
reff
r0
k
keff
k
c0
cpm
c0
Figure 1.50
Equivalent gas model used for a porous material in air.
For porous materials with a rigid skeletal frame and porosities close to unity, sound propagation
can be modelled with a single longitudinal wave by using the concept of an equivalent gas to
represent the porous material and the gas (usually air) contained within it (Morse and Ingard,
1968). Within a porous material the compressibility of the gas is altered, and its effective mass
is increased because the flow of the gas is impeded by the porous structure. Hence, the
equivalent gas is described by using an effective gas compressibility, κeff , and an effective gas
density, ρeff .
The gas compressibility, κ, equals the reciprocal of the bulk compression modulus of a gas, K ,
such that
κ=
1 ∂ρ
1
=
K
ρ ∂P
(1.166)
where ρ is the gas density.
From this point onwards we will assume that the gas in the porous material is always air. The
equivalent gas model is shown in Fig. 1.50. For an infinite medium without internal losses, K
takes real values; K = P0 for an isothermal process and K = 1.4P0 for an adiabatic process,
where P0 is the static pressure for air (usually taken as 1.013 × 105 Pa at atmospheric pressure).
However, in a porous material it is necessary to use complex values to include the effect
of internal damping. For sound propagation in typical rooms, the distances are only usually
large enough to require consideration of air absorption in the high-frequency range; this is an
internal loss due to the conversion of sound into heat energy. These internal losses occur due
to both thermal conduction and viscosity, and result from the molecular constitution of the gas;
in an infinite medium the thermal conduction and viscosity contribute almost equally to the
internal damping (Morse and Ingard, 1968). In a porous material, sound propagates close to
the boundaries of the skeletal frame and the losses due to thermal conduction and viscosity
are much larger. Therefore we need to account for these internal losses by using complex
values for both the effective gas compressibility and the effective gas density.
The effective gas compressibility varies over the building acoustics frequency range, and
depends upon heat transfer between the air and the frame. At ‘low’ frequencies, the rate of
83
S o u n d
I n s u l a t i o n
compression and rarefaction for the longitudinal sound wave in a porous material is sufficiently
slow that heat is transferred back and forth between the air and the frame. This means that the
temperature remains relatively constant and the process can be assumed to be isothermal. At
‘high’ frequencies there is insufficient time for this heat transfer to take place, so it becomes
an adiabatic process. There is no general definition of ‘low’ and ‘high’ frequencies. As a ruleof-thumb for fibrous materials over the building acoustics frequency range, it can be assumed
that ‘low’ corresponds to the low-frequency range, and ‘high’ corresponds to the high-frequency
range, with a transition between isothermal and adiabatic in the mid-frequency range.
The effective gas density also varies with frequency. This can be described in terms of the mass
impedance of the skeletal frame, iωmframe (Beranek, 1947). At ‘low’ frequencies where the mass
impedance is small, the compressions and rarefactions of the air particles cause the frame
to move too; hence the effective gas density needs to take account of the mass of the
frame. At ‘high’ frequencies where the mass impedance is large, the frame effectively remains
motionless.
The concept of an equivalent gas allows sound propagation in porous materials to be described
using two parameters, both of which are complex: the complex wavenumber, kpm , and the characteristic impedance, Z0,pm . Assuming harmonic time dependence for a wave using the term
exp(iωt), the wave equation for sound propagation in the porous material has the same form as
the wave equation for an infinite medium (Eq. 1.14); the difference being that the wavenumber,
k, is replaced by kpm .
The complex wavenumber for sound in a porous material, kpm , is
kpm = Re{kpm } + iIm{kpm } =
ω
cpm
(1.167)
where the phase velocity for sound in the porous material, cpm , is also complex, and equals
cpm =
1
φρeff κeff
(1.168)
The complex wavenumber is used here to clarify the link between propagation of sound in air
and propagation in a porous material via the wave equation. Note that some texts prefer to use
the propagation constant, Ŵ, which is related to the complex wavenumber by
(1.169)
Ŵ = ikpm
The characteristic impedance for air in a porous material, Z0,pm , is determined in the same way
as the characteristic impedance for air in an infinite medium (Eq. 1.18), which gives
Z0,pm =
p
= ρeff cpm =
u
ρeff
φκeff
(1.170)
The complex wavenumber (Eq. 1.167) and the characteristic impedance (Eq. 1.170) are both
calculated from the effective density and the effective gas compressibility. The latter two parameters can be calculated if the structure of the porous material can be represented using idealized
geometry. For example, representing all pores by cylindrical tubes at a specified angle to the
surface of a sheet of porous material, or representing all the fibres in a sheet of fibrous material
by long cylindrical tubes that lie in planes parallel to the surface of the sheet. Microstructural
84
Chapter 1
models that assume idealized geometry can be quite complicated. However, they can give
an effective density and gas compressibility that adequately represents real porous materials
as well as giving an insight into which parameters are important for sound propagation (e.g.
see Allard, 1993). For many porous materials the geometry is not simple and requires a statistical description. However, an alternative, simpler approach can be taken that avoids direct
calculation of the effective density and the effective gas compressibility whilst retaining use of
the equivalent gas model. This makes use of empirical relationships to determine the complex
wavenumber and the characteristic impedance.
The most widely used empirical equations are those of Delany and Bazley (1969, 1970). These
form a benchmark against which many other theories are tested, and other empirical equations
are compared. They were derived from a large number of measurements on different fibrous
materials. The resulting empirical equations for Z0,pm and kpm only require knowledge of the
airflow resistivity which can be measured or determined from other empirical relationships.
Although these empirical equations were based upon fibrous materials they can be used to
estimate values for porous foams with r < 10 000 Pa.s/m2 (Allard, 1993).
The assumption of a rigid skeletal frame allows empirical laws to be used to calculate sound
propagation in isotropic, homogeneous, porous materials. Fibrous materials such as mineral
wool can be considered as relatively homogeneous, although they are formed from layers so
they are anisotropic. However, by considering propagation through the material in only a single
direction, they can be treated as isotropic, homogeneous materials.
Empirical equations are not absolute laws; there are many different materials and there is often
more than one way to group or plot the data to carry out regression analysis. Other empirical
equations to determine the characteristic impedance and the propagation constant for fibrous
materials can be found in the literature (e.g. Mechel and Vér, 1992). A theoretical model for rigid
frame fibrous materials from Allard and Champoux (1992) gives similar values to the Delany
and Bazley equations in the range of validity but improves the low-frequency trends.
The empirical equations of Delany and Bazley (1969, 1970) are
Z0,pm = ρ0 c0 (1 + 0.0571X −0.754 − i0.087X −0.732 )
(1.171)
and
kpm = Re{kpm } + iIm{kpm } =
2πf
(1 + 0.0978X −0.700 − i0.189X −0.595 )
c0
(1.172)
where the variable, X , is
X =
ρ0 f
r
(1.173)
The range of validity for Eqs 1.171 and 1.172 is (Delany and Bazley, 1969)
0.01 < X < 1.0
(1.174)
For the equivalent gas model, the wavelength of sound within the porous material, λpm , is
calculated using
λpm =
2π
Re{kpm }
(1.175)
85
S o u n d
I n s u l a t i o n
7
Wavelength in equivalent gas calculated
using Delany and Bazley equations for
different airflow resistivities
6
2000 Pa.s/m2 (50–1600 Hz)
4000 Pa.s/m2 (50–3150 Hz)
4
Ai
8000 Pa.s/m2 (50–5000 Hz)
ra
t2
3
16 000 Pa.s/m2 (80–5000 Hz)
0˚
C
Wavelength (m)
5
32 000 Pa.s/m2 (160–5000 Hz)
2
64 000 Pa.s/m2 (315–5000 Hz)
1
0
50
80
125
200
315
500
800 1250 2000
One-third-octave-band centre frequency (Hz)
3150
5000
Figure 1.51
Comparison of the wavelength in porous materials using the equivalent gas model with the wavelength in air. The range of
validity for the Delany and Bazley equations is shown in brackets in terms of frequency.
In Fig. 1.51 the wavelength in air can be compared with the wavelength for porous materials in
air that is calculated using the Delany and Bazley equations. The calculations use a range of
airflow resistivities (2000 to 64 000 Pa.s/m2 ) that represents porous materials commonly used
in buildings. The range of validity for the Delany and Bazley equations (Eq. 1.174) usually
allows use of the equivalent gas model for a large part, but not all, of the building acoustics frequency range. The wavelength in the equivalent gas is significantly shorter than air in the lowand mid-frequency ranges, but tends towards the wavelength in air within the high-frequency
range.
The sound pressure for a plane wave propagating through a porous material in the positive
x-direction is described by
p(x, t) = p̂ exp(−ikpm x) exp(iωt) = p̂ exp(−iRe{kpm }x) exp(Im{kpm }x) exp(iωt)
(1.176)
The definition of a complex wavenumber implies attenuation with distance, hence Im{kpm } is
negative; this can be seen in the empirical equation for fibrous materials (Eq. 1.172). Therefore
the amplitude of the plane wave decreases with distance according to the decaying exponential
term, exp(Im{kpm }x). This gives the decrease in sound pressure level in decibels, LP , after
propagating a distance, x, through the porous material,
LP =
20
|Im{kpm }|x
ln 10
(1.177)
The installation of porous materials in air spaces means that it is often necessary to account
for the reflection that occurs when sound enters the material from air, and when it exits the
material into air. This is described in Section 4.3.9 in the calculation of the normal incidence
sound reduction index for porous materials.
86
Chapter 1
In some cases there is no air space between the porous material and the plate that forms
part of a wall or floor, such as a cavity wall where a porous material fills the cavity. For a
plate undergoing bending wave vibration that is immediately next to a porous material, sound
transmission from the plate into and through the porous material may need to use Biot theory
for the porous material to take account of the shear wave and two longitudinal waves that can
propagate within it. In this case, the simplified assumption of an equivalent gas may no longer
be appropriate.
1.3.3 Local modes
From Eq. 1.54 the mode frequencies of closed cavities are calculated using
2 2
c
q
r
p 2
fp,q,r =
+
+
2
Lx
Ly
Lz
(1.178)
1.3.3.1 Modal density
To calculate the cavity modal density across the building acoustics frequency range we not
only need to consider three-dimensional sound fields like in rooms, but also one-dimensional
(p = 0 and q = r = 0), and two-dimensional (p = 0, q = 0, and r = 0) sound fields. Hence we
can represent cavities as a one-dimensional space of length, Lx , a two-dimensional space of
surface area, S = Lx Ly , and a three-dimensional space of volume, V = Lx Ly Lz .
One-, two-, and three-dimensional sound fields can occur in lightweight walls and floors where
cavities are formed by a framework of studs or joists. In these cases, the one or two-dimensional
modal density can be determined by using Eq. 1.178 to calculate the mode frequencies; the
number of modes that fall within each band are then divided by the bandwidth. As with rooms
it is simpler to use the following statistical approaches.
For a long (Lx ), narrow (Ly ), and thin (Lz ) cavity at low frequencies there is a one-dimensional
sound field consisting purely of axial modes. The modal density is calculated in the same way
as for structural waves on beams (Section 2.5.1.4), hence
n1D (f ) =
2Lx
c
(1.179)
At frequencies at and above f0,1,0 , but below f0,0,1 , the cavity acts as a two-dimensional space
that supports axial and tangential modes. To count the number of modes the eigenvalues are
arranged in a two-dimensional lattice as shown in Fig. 1.52 (Price and Crocker, 1970). Eigenvalues that lie along the x and y-axes represent axial modes; those that lie on the coordinate
plane kx ky (excluding the eigenvalues on the axes) represent tangential modes. The area
associated with each eigenvalue is a rectangle with an area of π2 /Lx Ly (which equals π2 /S).
The number of modes below a specified wavenumber, k, is equal to the number of eigenvalues
that are contained within one-quarter of the area of a circle with radius, k. However, one-half
of the area associated with each axial mode falls outside the permissible area in k-space that
can only have zero or positive values of kx and ky . Therefore calculating the number of modes
is a two-step process. The first step is to divide πk 2 /4 by π2 /S to give an estimate for the number of tangential modes that also includes one-half of the axial modes. The second step is to
account for the other halves of the axial modes that lie on the x and y-axes by taking one-half
87
S o u n d
I n s u l a t i o n
ky
k
Tangential mode
k2,3
π/Ly
0,0
k
π/Lx
kx
Axial mode
Figure 1.52
Mode lattice for a two-dimensional space. The vector corresponding to eigenvalue, k2,3 , is shown as an example. The shaded
area indicates the fraction of the area associated with axial modes that falls outside the permissible area in k-space. The area
enclosed by a circle with radius, k, encloses eigenvalues below wavenumber, k.
of k/(π/(Lx + Ly )). The sum of these two components gives the number of modes, N(k), below
the wavenumber, k, where
N(k) =
k(Lx + Ly )
k 2S
+
4π
2π
(1.180)
Hence, from Eq. 1.57 the modal density is
n2D (f ) =
Lx + Ly
2πfS
+
c2
c
(1.181)
For cavities that are not box-shaped, and for cavities where there is ambiguity about whether
it is reasonable to assume rigid boundaries for one or two of the four boundaries (i.e. those
that lie along the planes where x = 0, x = Lx , y = 0, and y = Ly ), the modal density can be
calculated by using only the first term in Eq. 1.181,
n2D (f ) =
88
2πfS
c2
(1.182)
Chapter 1
The crossover point from a two-dimensional to a three-dimensional sound field occurs at the
frequency where there is a half wavelength across the smallest dimension, Lz , which is usually the cavity depth. This corresponds to the axial mode f0,0,1 , the first cross-cavity mode,
where
f0,0,1 =
c
2Lz
(1.183)
At and above f0,0,1 there are axial, tangential, and oblique modes, hence the cavity acts as a
three-dimensional space for which the modal density is
n3D (f ) =
πfST
4πf 2 V
LT
+
+
c3
2c 2
8c
(1.184)
where ST is 2(Lx Ly + Lx Lz + Ly Lz ) and LT is 4(Lx + Ly + Lz ).
For cavities that are not box-shaped, and for box-shaped cavities in the high-frequency
range, a reasonable estimate of the modal density is found by using only the first term in
Eq. 1.184.
The statistical mode count in a frequency band is calculated from the modal density using
Eq. 1.63. Mode counts are now used to gain an insight into the distribution of modes for two
common cavities, a timber joist floor cavity and a wall cavity (see Fig. 1.53). The timber joist
floor cavity is long, narrow, and thin; in the low-frequency range this results in only axial modes
along the longest dimension, Lx . Above the first cross-cavity mode in the 800 Hz band there is
then a rapid increase in the number of modes with increasing frequency. In contrast, the wall
cavity has a two-dimensional sound field over the majority of the building acoustics frequency
range with the first cross-cavity mode in the 2500 Hz band.
As with rooms, the distribution of the different mode types is useful in determining which internal
cavity surfaces should be lined with absorbent material to reduce the sound level in the cavity.
We can take the timber joist floor cavity as an example. To absorb sound in the low-frequency
range where there are only axial modes along Lx , absorbent material could be positioned over
the surfaces perpendicular to the x-axis at the ends of the cavity where Lx = 0 m and Lx = 4 m.
In practice, floor cavities are often partially or fully filled with absorbent material along their
entire length to absorb sound energy stored in axial, tangential, and oblique modes.
1.3.3.2
Equivalent angles
Equivalent angles for local modes in rooms were introduced in Section 1.2.5.4. Figure 1.54
shows equivalent angles for the timber joist floor cavity and wall cavity described in Fig. 1.53.
These can be compared with the equivalent angles for a 50 m3 room (refer back to Fig. 1.16).
Below the first cross-cavity mode, θz = 0◦ , because the sound field is two-dimensional and
there is a limited range of angles. Above the first cross-cavity mode the range of angles tends
to cover the full range from 0◦ to 90◦ ; however, the elongated shape of the timber joist floor
cavity means that the distribution of angles between θx , θy , and θz is uneven when compared
with the 50 m3 room.
Compared with rooms, the small volumes and elongated shapes associated with typical cavities
means that in the building acoustics range there is often a limited range of angles from which
the sound waves will arrive at any point in the space.
89
S o u n d
I n s u l a t i o n
(a) Timber joist floor cavity
10 000
Timber joist floor cavity
1000
Lx ⫽ 4 m, Ly ⫽ 0.4 m, Lz ⫽ 0.225 m
f0,0,1 ⫽ 762 Hz
Axial modes
Tangential modes
Mode count, N
Oblique modes
100
10
1
0.1
20
31.5
50
80
125
200
315
500
800 1250
One-third-octave-band centre frequency (Hz)
2000
3150
5000
2000
3150
5000
(b) Wall cavity
10 000
1000
Wall cavity
Lx ⫽ 4 m, Ly ⫽ 2.5 m, Lz ⫽ 0.075 m
f0,0,1 ⫽ 2287 Hz
Axial modes
Tangential modes
Mode count, N
Oblique modes
100
10
1
0.1
20
31.5
50
80
125
200
315
500
800 1250
One-third-octave-band centre frequency (Hz)
Figure 1.53
Mode count for a timber joist floor cavity and a wall cavity.
1.3.4 Diffuse field
A diffuse field in a cavity is defined in the same way as for rooms. However, when there is a
two-dimensional sound field we need to account for the fact that waves can only arrive from
directions within one plane rather than from all possible directions in three-dimensional space.
Compared to rooms, cavities have much smaller volumes and the sound field can only usually
be considered as diffuse over a narrow part of the building acoustics frequency range.
90
Chapter 1
(a) Timber joist floor cavity
90
80
70
60
Angle (°)
ux
50
uy
40
uz
30
20
10
0
10
100
1000
10 000
1000
10 000
Frequency (Hz)
(b) Wall cavity
90
80
70
60
Angle (°)
ux
50
uy
40
uz
30
20
10
0
10
100
Frequency (Hz)
Figure 1.54
Equivalent angles for the modes of a timber joist floor cavity and a wall cavity.
91
S o u n d
I n s u l a t i o n
1.3.4.1 Mean free path
As with rooms, the mean free path is only defined for the situation where all reflections from the
boundaries are diffuse. When a cavity acts as a three-dimensional space, the mean free path
is the same as for rooms and is defined in Eq. 1.47. The mean free path for a two-dimensional
space has already been derived in Section 1.2.3.1 and is given by (Kosten, 1960)
dmfp =
πS
U
(1.185)
where U is the perimeter of the cavity (U = 2Lx + 2Ly for a rectangular cavity with a depth, Lz ).
1.3.5 Damping
In rooms, absorptive material is often distributed in one of two ways: either it is distributed over
all the surfaces, or one or two of the room surfaces provide the majority of the absorption area
(e.g. highly absorbent tiles that cover the ceiling). In cavities there is more scope to vary the
distribution of absorbent material; it can be placed within the cavity volume as well as over the
surfaces.
The implications of one, two, and three-dimensional sound fields in cavities becomes apparent
when we consider the position of the absorption within the cavity. Below the first cross-cavity
mode there are only axial and tangential modes in the cavity, hence sound waves are only
incident upon the perimeter of the cavity. To absorb sound in this frequency range, absorptive
material needs to be placed around the perimeter of the cavity. In fact, this is sometimes the
only practical place to position the absorption. An example of this is high performance windows
in music studios, where two or more glazing units are separated by wide cavities. To increase
the absorption of sound at and above the first cross-cavity mode the two main surfaces that
face into the cavity also need to be absorptive.
Cavities within plasterboard and masonry walls are often filled or partially filled with absorbent
porous material. This introduces additional internal losses as sound waves propagate through
the porous material.
Cavities tend to have relatively small volumes which often contain additional absorbent material
so it is not usually necessary to consider air absorption for the building acoustics frequency
range.
1.3.5.1 Reverberation time
Sound fields in cavities rarely approximate a diffuse field in either two or three dimensions,
hence the decay curves tend to show various degrees of curvature. The reasons for this are
similar to those previously discussed for non-diffuse fields in rooms; normal mode theory indicates that the degree of curvature varies depending upon the combination of axial, tangential,
and oblique modes in a frequency band.
Reverberation times in cavities tend to be shorter than those in rooms; examples are shown
in Fig. 1.55 which were measured using T10 , T15 , or T20 .
1.3.5.2
Internal losses
Below the frequency of the first cross-cavity mode, the internal loss factor is determined by
the absorption of the surface at the cavity perimeter. For locally reacting surfaces, Eq. 1.76
92
Chapter 1
(a) Timber joist floor cavity
0.8
0.7
Reverberation time (s)
0.6
0.5
Lx
Lz
0.4
Ly
0.3
0.2
0.1
Measured
0.0
50
80
125
200
315
500
800 1250 2000
One-third-octave-band centre frequency (Hz)
3150
5000
(b) Masonry wall cavity
0.8
Lx
0.7
Reverberation time (s)
0.6
Ly
0.5
Lz
0.4
0.3
0.2
Measured
0.1
0.0
50
80
125
200
315
500
800 1250 2000
One-third-octave-band centre frequency (Hz)
3150
5000
Figure 1.55
Examples of measured reverberation times in cavities. (a) Timber joist floor cavity. Lx = 4.2 m, Ly = 0.4 m, and Lz = 0.225 m.
Areas Ly Lz are fair-faced masonry. Areas Lx Lz are timber joists. Area Lx Ly (upper) is chipboard. Area Lx Ly (lower) is
plasterboard. (b) Masonry wall cavity. Lx = 3.6 m, Ly = 5.0 m, and Lz = 0.075 m. Areas Lx Ly , Lx Lz and Ly Lz (left side) are
fair-faced masonry. Area Ly Lz (right side) is 455 mm, 28 kg/m3 mineral wool (cavity stop). Measured data from Hopkins are
reproduced with permission from ODPM and BRE.
gives the angle-dependent sound absorption coefficient, however, to simplify the calculation it
is assumed that αθ = α0 cos(θ), which gives (Price and Crocker, 1970)
ηii =
S P αP c 0
2π2 fV
(1.186)
where SP is the surface area of the cavity perimeter and αP is the average statistical sound
absorption coefficient for the cavity perimeter. For box-shaped cavities, SP = 2(Lx Lz + Ly Lz )
93
S o u n d
I n s u l a t i o n
and αP = 4k=1 Sk αk /SP where Sk and αk correspond to the area and statistical absorption
coefficient for each side of the cavity perimeter. If the statistical absorption coefficients are
not available and the perimeter surface is locally reacting, then αP can be estimated from the
normal incidence absorption coefficient, α0 , using (Price and Crocker, 1970)
αP =
π
α0
4
(1.187)
At and above the frequency of the first cross-cavity mode, the internal loss factor is
ηii =
ST αc0
8πfV
(1.188)
where ST is the total area of all the cavity surfaces and α is the average statistical sound absorption coefficient for all the cavity surfaces. For box-shaped cavities, ST = 2(Lx Ly + Lx Lz + Ly Lz )
and α = 6k=1 Sk αk /ST where Sk and αk correspond to each surface of the cavity.
Near the frequency of the first cross-cavity mode an issue arises in using Eqs 1.186 and
1.53 to calculate the internal loss factor. This is because there is often a significant difference
between the values for two-dimensional and three-dimensional sound fields. Over the building
acoustics frequency range, this causes a sharp transition in the predicted internal loss factor
and the predicted reverberation time. In practice, damping measurements inside real cavities
indicate a more gradual transition. In the prediction of sound transmission, this is not usually a
problem as a sharp transition will not normally occur in the predicted sound insulation because
of the existence of many other sound transmission paths.
1.3.5.2.1
Sound absorption coefficient: Locally reacting porous materials
Calculation of the internal loss factor requires the normal incidence or statistical sound absorption coefficient for the cavity boundaries. For porous materials the absorption coefficient can
be calculated by treating the material as an equivalent gas and using wave theory to calculate
the specific acoustic impedance or admittance (e.g. see Allard, 1993). This can make use
of equations such as those of Delany and Bazley (Section 1.3.2.2) or Allard and Champoux
(1992) to determine Z0,pm and kpm for the equivalent gas.
It is assumed that the porous material is locally reacting with a thickness, h, and is positioned
a distance, d, from a rigid non-porous surface that has an infinite impedance (see Fig. 1.56).
For this calculation, most masonry/concrete walls and floors can be assumed to be rigid.
The calculations in this section are equally applicable to rooms where locally reacting porous
materials are placed near masonry/concrete walls or floors.
The normal incidence and statistical absorption coefficients can be calculated using Eqs 1.79
and 1.81 respectively where the specific acoustic admittance is calculated using
βa,s = βRe − iβIm
⎛
⎞−1
0 c0
− Zρ0,pm
tan(kpm h) tan 2πfd
c0
iZ
1
0,pm
⎠
=⎝
=
Za,s
ρ0 c0 tan 2πfd + ρ0 c0 tan(k h)
pm
c0
Z0,pm
(1.189)
When d = 0, the porous layer is next to the rigid surface (often referred to as rigid backing).
Equation 1.189 then reduces to
−1
−iZ0,pm
1
1
βa,s = βRe − iβIm =
(1.190)
=
Za,s
ρ0 c0 tan(kpm h)
94
Chapter 1
Air Equivalent Air Rigid
gas
surface
Incident wave
ρ0
ρeff
ρ0
κ
κeff
κ
c0
cpm
c0
h
d
Figure 1.56
Absorber: porous material – air gap – rigid surface. Equivalent gas model used to represent the porous material.
When d = nλ/2 for n = 1, 2, 3, etc., the specific acoustic admittance calculated from Eq. 1.189
is the same as Eq. 1.82, and the porous material can be considered as rigidly backed.
Examples for the statistical absorption coefficient are shown in Fig. 1.57 for a range of airflow
resistivities from 2000 to 64 000 Pa.s/m2 . Two thicknesses of porous material are considered,
h = 0.025 m and h = 0.1 m, each of which have air gaps of d = 0 m and d = 0.1 m. For rigid
backing, increasing the thickness of the material from 25 to 100 mm significantly increases
the absorption coefficient in the low- and mid-frequency ranges. However, by using a 100 mm
air gap with the 25 mm material it is possible to achieve similarly high values to the 100 mm
material with rigid backing in the low- and mid-frequency ranges; this is at the expense of lower
absorption coefficients in the high-frequency range. With an air gap, the curve for the absorption
coefficient has a ripple with troughs that tend to become less pronounced with increasing airflow
resistivity. In practice, this ripple is less pronounced due to the use of frequency bands, variation
in material properties, and variation in d due to workmanship.
The airflow resistance of porous materials tends to increase with increasing bulk density, but
there is no simple rule that porous materials with low or high airflow resistivity will always give
the highest absorption coefficients over the building acoustics frequency range. To determine
suitable values of r , h, and d, it is necessary to identify which part of the frequency range
requires the highest absorption coefficients. There are a large number of permutations for
these three variables, and measured absorption coefficients for a specific combination are not
always available. In order to assess their effect it is usually sufficient to calculate the absorption coefficient as described in this section. For fibrous materials, a wide range of densities
are available (typically 10 to 200 kg/m3 ) which gives a wide range of airflow resistivities from
which to choose a specific material. However, commonly available materials come in a limited
range of thicknesses, which, in combination with the cavity dimensions will limit the choice of
h and d.
95
S o u n d
I n s u l a t i o n
(a) h ⫽ 0.025 m, d ⫽ 0 m
(b) h ⫽ 0.025 m, d ⫽ 0.1 m
1.0
Statistical absorption coefficient, αst (⫺)
Statistical absorption coefficient, αst (⫺)
1.0
0.9
0.8
0.7
64 000
0.6
32 000
0.5
16 000
0.4
8000
4000
2000
0.3
0.2
0.1
0.0
100
1000
Frequency (Hz)
0.5
0.4
0.3
8000
0.2
4000
2000
0.1
1000
Frequency (Hz)
5000
1.0
Statistical absorption coefficient, αst (⫺)
Statistical absorption coefficient, αst (⫺)
0.6
(d) h ⫽ 0.1 m, d ⫽ 0.1 m
1.0
0.9 16 000
0.8 32 000
0.7 64 000
0.6
0.5
8000
0.4
4000
2000
0.3
0.2
0.1
0.0
100
0.7
0.0
100
5000
(c) h ⫽ 0.1 m, d ⫽ 0 m
32 000
16 000
0.8 64 000
0.9
1000
Frequency (Hz)
5000
8000
16 000
4000
0.8
0.9
0.7
32 000
64 000
2000
0.6
0.5
0.4
0.3
0.2
0.1
0.0
100
1000
Frequency (Hz)
5000
Figure 1.57
Statistical absorption coefficients of porous materials for a range of airflow resistivities in Pa.s/m 2 .
1.3.5.3 Coupling losses
Calculation of the coupling loss factors involving the cavity are discussed in Section 4.3.5.3.
1.3.5.4 Total loss factor
The total loss factor equals the sum of the internal and coupling loss factors. For most cavities in
walls and floors that have absorptive surfaces, the coupling loss factors are much smaller than
the internal loss factor, and the latter provides a reasonable estimate of the total loss factor.
As with rooms, Eq. 1.107 can be used to calculate the total loss factor from the reverberation
time and vice versa.
1.3.5.5 Modal overlap factor
The modal overlap factor for cavities is calculated using Eq. 1.109.
96
Chapter 1
1.3.6
Energy
Calculation of the sound energy stored in a cavity is calculated using Eq. 1.154 in the same
way as for rooms.
1.4 External sound fields near building façades
To assess the airborne sound insulation of the building façade from external sound sources it is
necessary to measure the sound pressure levels both inside and outside the building. Having
looked at the internal sound field, we will now look at the external sound field near a façade.
The sound pressure level near the façade depends upon: the position of the microphone in
relation to the façade and the ground, diffraction effects from the edges of the façade, diffraction
effects from protruding or recessed elements on the building façade (e.g. balconies), sound
propagation from the source (including the effects of ground impedance, façade impedance,
and meteorological conditions), the orientation of the sound source, and the type of sound
source outside the building (e.g. point source, line source).
Microphone positions relative to the façade and the ground often differ depending upon whether
the primary aim is to measure the façade sound insulation, or measure/predict the environmental noise near the façade. In the latter case, the measurements/predictions are often used at a
later point in time to estimate sound transmission into the building via the façade; it is clearly
advantageous if the microphone positions are the same or the levels can be accurately converted. For field measurements of façade sound insulation, the microphone is usually attached
to the surface of the façade at variable heights that depend upon the building element that is
being measured, or positioned 2 m in front of the façade at a height of 1.5 m above the floor of
the receiving room (ISO 140 Part 5). Environmental noise measurements are taken at a variety
of different positions; often at a height of 1.2, 1.5, or 4 m above floor level, and at distances
between 1 and 2 m in front of the façade (ISO 1996 Part 1).
In practice we often need to convert sound pressure levels near the building façade to freefield levels in the absence of the façade and vice versa. This section therefore looks at the
difference between the external sound pressure level with the façade to the level without the
façade (i.e. the change in level due to the presence of the façade).
1.4.1 Point sources and semi-infinite façades
For façade sound insulation measurements made with a loudspeaker and some environmental
noise sources it is appropriate to consider a point source. We therefore start by looking at the
sound field generated by a point source in the vicinity of a façade. By creating the image sources
for this situation as shown in Fig. 1.58, we see that the sound travels from the source (S) to
the receiver (R) via four different paths: the first path is the direct path from the source to the
receiver, the second path involves a single reflection from the ground, the third path involves
a single reflection from the ground and a single reflection from the façade, and the fourth path
involves a single reflection from the façade. The path lengths in terms of the distance, d, from
the source, or image source, to the receiver are also indicated in this diagram.
We will assume that: (a) the source emits spherical waves, (b) the ground and façade
are perfectly reflecting with no phase change upon reflection (c) all reflections are specular
97
S o u n d
I n s u l a t i o n
Façade
R
d4
d1
S
d2
d3
Ground
Figure 1.58
External façade sound pressure level measurements. Source (S) and receiver (R) orientation with image sources (✰) for the
different propagation paths.
(d) the façade has dimensions that are very large compared to the wavelength (i.e. a semiinfinite plate), and (e) there are no other façades nearby that significantly affect the sound
field. Therefore we will not concern ourselves with diffraction from the edges of the wall or with
different impedances for the ground and the façade. The assumption of specular reflection is
reasonable for this situation, particularly below 1000 Hz; it can generally be assumed that real
façades have small scattering coefficients (Ismail and Oldham, 2005).
We are interested in the difference between the sound pressure level in front of the façade and
the free-field level without the façade. This requires the ratio of the total mean-square sound
pressure, p2 , to the mean-square sound pressure, (p1 + p2 )2 ; the latter term corresponds to
the combination of the direct path between source and receiver (path length d1 ), and the path
in which the sound is reflected directly from the ground to the receiver (path length d2 ). The
sound pressure for spherical waves at single frequencies is taken from Eq. 1.22, hence the
required ratio is
exp(−ikd1 ) exp(−ikd2 ) exp(−ikd3 ) exp(−ikd4 ) 2
+
+
+
p2 t
d1
d2
d3
d4
=
(1.191)
2
2
(p1 + p2 ) t
exp(−ikd1 ) exp(−ikd2 )
+
d
d
1
2
Now we can calculate the change in level due to the presence of the façade for different
distances of the receiver from the façade. For façade sound insulation measurements, the
external microphone is usually at a distance of 2 m from the façade or on the surface of the
façade (ISO 140 Part 5); hence we will use these to define the minimum and maximum distances for the range of interest. To illustrate the effect of intermediate distances we will look at
300 mm and 1 m.
For measurements on the surface of the façade there are usually physical limitations that determine how close the microphone can be positioned to the surface. For a half-inch microphone
(12.7 mm diameter) attached to the façade with the axis of the microphone parallel to the plane
of the façade, we can assume that the façade-receiver distance is 6.35 mm (i.e. the distance
from the façade surface to the centre of the microphone diaphragm). Figure 1.59 shows the
98
Chapter 1
Change in SPL due to the presence of a façade (dB)
9
6
3
0
⫺3
⫺6
⫺9
⫺12
⫺15
⫺18
⫺21
⫺24
Façade–receiver distances
6.35 mm
300 mm
1m
2m
⫺27
⫺30
⫺33
⫺36
0
100
200
300
400
500
600
Frequency (Hz)
R
700
800
900
1000
d
1.2 m
S
20 m
0.5 m
Figure 1.59
Change in the sound pressure level due to the presence of the façade for receiver positions at four different distances, d, from
the façade (single frequencies from 0.25 to 1000 Hz). Source-receiver-façade geometry is indicated in the sketch.
calculated level difference for four different façade-receiver distances at frequencies up to
1000 Hz. For a half-inch microphone attached to the surface of a façade there is a constant
level difference of 6 dB, this is referred to as pressure doubling. As the microphone is moved
further away from the façade we see that there are interference minima in the spectrum due
to destructive interference between the different propagation paths. These occur due to the
different distances travelled by the sound waves along each of the different paths. For the various combinations of paths, the path difference in metres corresponds to a phase difference
in radians. Destructive interference occurs where the path length difference, dpq , between
paths p and q, corresponds to a phase difference of an odd number of π radians,
2π
dpq
= (2n + 1)π
λ
(1.192)
where n = 0, 1, 2, 3, etc.
99
S o u n d
I n s u l a t i o n
The upper frequency shown in this example has been limited to 1000 Hz because at higher
frequencies, turbulent air in the outdoor environment tends to reduce the coherence between
the waves that travel along the different propagation paths (Attenborough, 1988; Quirt, 1985).
As a result, this simple model is no longer appropriate, and sharp minima in the spectrum due
to destructive interference are less likely to occur above 1000 Hz.
In practice we usually deal with frequency bands rather than single frequencies. For frequency
bands the same ratio can be calculated from the band centre frequency using
2 2 2
d1
d1
d1
1+
+
+
d2
d3
d4
+
p2 t
=
(p1 + p2 )2 t
+
2d1
2d1
2d1
R(d12 ) +
R(d13 ) +
R(d14 )
d2
d3
d4
2d12
2d12
2d12
R(d23 ) +
R(d24 ) +
R(d34 )
d2 d3
d2 d4
d3 d4
2
d1
2d1
+
R(d12 )
1+
d2
d2
(1.193)
where the autocorrelation function, R(dpq ) for each path length difference (magnitude), dpq ,
is (Delany et al., 1974)
2πdpq
2πBL dpq
λ
cos
(1.194)
sin
R(dpq ) =
2πBL dpq
λ
λ
for which λ is the wavelength corresponding to the band centre frequency, and BL is calculated
from the lower and upper frequency limits of the band, fl and fu , using BL = (fu − fl )/(fu + fl ).
BL = 0.115 for one-third-octave-bands.
For one-third-octave-bands between 50 and 1000 Hz the change in level due to the presence
of the façade is shown in Fig. 1.60 (source and receiver positions are the same as in Fig. 1.59).
The change in level is 6 dB for a half-inch microphone attached to the surface of the façade.
One advantage of using surface measurements is that if the microphone is positioned very
close to the surface, we can avoid interference minima in the building acoustics frequency
range, although there will be small departures from pressure doubling in the high-frequency
range. This allows us to make the convenient assumption of pressure doubling. In contrast, the
façade–receiver distance of 300 mm provides an example of the variation that can be introduced
when measurements are not made on the surface of the façade. With this particular combination
of source–receiver–façade geometry there is an interference dip around the 315 Hz band. If,
for example, we were to change the façade-receiver distance from 300 to 200 mm, we would
shift the interference dip into a different frequency band. This dependence on the specific
geometry of each situation illustrates the importance of well-defined measurement positions for
comparative measurements of the sound field near façades. For a façade-receiver distance of
1 or 2 m there are dominant interference minima in the low-frequency range. However the onethird-octave-bands get wider as the frequency increases and the interference effects begin to
average out. For façade-receiver distances of 1 or 2 m in the mid-frequency range, the change
in level tends towards 3 dB; this is referred to as energy doubling.
For a point sound source, such as a loudspeaker in the low-frequency range, the above discussion indicates that the sound pressure level will also vary over the surface of a façade due to
the different interference patterns that occur with different source–receiver–façade geometries.
In the mid- and high-frequency ranges, loudspeakers tend to become increasingly directional
100
Change in level due to the presence of a façade (dB)
Chapter 1
9
6
3
0
⫺3
Façade–receiver distances
6.35 mm
⫺6
300 mm
1m
⫺9
2m
⫺12
50
63
80
100 125 160 200 250 315 400 500
One-third-octave-band centre frequency (Hz)
630
800 1000
Figure 1.60
Change in the sound pressure level due to the presence of the façade for receiver positions at different distances from the
façade.
and no longer act as point sources, hence the variation in sound pressure level over the façade
is also affected by the directionality of the loudspeaker.
Although several assumptions have been made in this basic model, it adequately illustrates
the general trends. In practice there are other factors that affect the depth and frequency of the
interference minima. The finite impedance of the ground causes a phase change upon reflection
from the ground, so to improve the model it is necessary to incorporate measurements of the
ground impedance (e.g. see Ogren and Jonasson, 1998). Compared to the ground, relatively
little information is available on the impedance of façades. However, façade surfaces are rarely
highly porous and tend to have low-absorption coefficients (typically less than 0.1 in the lowand mid-frequency ranges). For this reason, the assumption of a perfectly reflecting surface is
often reasonable.
1.4.1.1 Effect of finite reflector size on sound pressure levels near the façade
As real façades are of finite size, we need to look at the effects of diffraction from the edges of
a façade. The sound field in front of finite size reflectors can be considered as the combination
of the four geometrical wave paths (as previously considered for the semi-infinite reflector),
combined with edge or boundary diffraction waves. To assess diffraction we will look at indoor
scale-model measurements because it is awkward to control all the relevant parameters with
outdoor measurements near real buildings. Results are taken from scale model experiments in
a semi-anechoic chamber with a concrete floor, and a 30 mm thick square reflector (varnished
board) to represent the façade. Good agreement between these measurements and predictions
using Integral Equation Methods (IEM) allow conclusions to be drawn purely by using measured
data (Hopkins and Lam, 2008). A 1:5 scale model was used for the measurements, but all the
results shown and discussed in this section are scaled-up to the situation for real façades (i.e.
full-size). The source was a small loudspeaker positioned in the vertical plane perpendicular to
101
S o u n d
I n s u l a t i o n
the center line of the reflector. The receiver position was offset from this plane by one-twelfth of
the reflector dimension to avoid perfect symmetry in the set-up that might be unrepresentative
of the situation in practice. Five square reflectors were tested that represented full-size façades
with side dimensions of 2, 3, 4, 5, and 6 m.
Figure 1.61 shows the change in level due to two square reflectors (6 × 6 m and 2 × 2 m) with
a façade–receiver distance of 2 m. Measured data is shown alongside the prediction for a
semi-infinite reflector (Eq. 1.191). Compared to the semi-infinite reflector, diffraction from the
edges of the finite reflector affects the frequency of the peaks and troughs as well as their
values. As one would expect, this is more pronounced for the smaller reflector. For small
reflectors, the receiver will be relatively close to the edges and the edge diffracted pressure
can significantly change the interference pattern in comparison to the semi-infinite reflector.
The 6 × 6 m reflector can be taken as being representative of the façade of a detached house,
and diffraction can be considered to have negligible effect on measured levels above 100 Hz.
For the 2 × 2 m reflector, diffraction can have a significant effect below 1000 Hz; in practice,
most façades are much larger than this, but it is used here to represent small square protruding
sections of a building (e.g. bay window, entrance hall, enclosed balcony).
For practical purposes we need to assess the difference between finite size reflectors and a
semi-infinite reflector in one-third-octave-bands; this is done using the difference between the
measured and the predicted (Eq. 1.193) change in level due to the presence of the façade.
Figure 1.62 shows this level difference for a façade–receiver distance of 2 m. For square
reflectors with side dimensions between 3 and 6 m, the level differences are generally less
than 3 dB in the low-frequency range. The differences are larger with the 2 × 2 m reflector,
particularly at 63 Hz, but they are generally less than 3 dB across the low- and mid-frequency
ranges. Environmental noise measurements are often taken using a façade–receiver distance
between 1 and 2 m. Figure 1.63 shows the level difference for 11 different façade–receiver
distances in 0.1 m steps from 1 to 2 m for each of four different square reflectors (side dimensions between 3 and 6 m). In the low-frequency range there are significant differences between
the semi-infinite and the finite reflectors due to diffraction. In the mid-frequency range these
differences are negligible and these finite reflectors can be treated as semi-infinite. The level
differences for the 2 × 2 m reflector are shown separately in Fig. 1.63; these indicate that it is
not appropriate to treat this small reflector as semi-infinite in both the low- and mid-frequency
ranges.
In practice there are so many permutations of source–receiver–façade geometry that it is
difficult to make a definitive statement about the conditions in which diffraction effects will
be negligible. As a rule-of-thumb for a point source near the ground, and façade–receiver
distances between 1 and 2 m, diffraction effects are only likely to be significant in the lowfrequency range for façades with dimensions <5 m.
1.4.1.2
Spatial variation of the surface sound pressure level
Façade elements such as windows or doors often have lower airborne sound insulation than
the wall around them. Hence, field measurement of the apparent sound reduction index is
often needed for these elements. This requires measurement of the average surface sound
pressure level over the element. A spatial average is needed for all elements regardless of
the source–receiver–façade geometry; usually between three and ten microphone positions
on the surface of the element (ISO 140 Part 5).
102
Change in SPL due to the presence of a façade (dB)
8
6
4
2
0
⫺2
⫺4
⫺6
⫺8
⫺10
⫺12
⫺14
⫺16
⫺18
⫺20
⫺22
⫺24
⫺26
Change in SPL due to the presence of a façade (dB)
Chapter 1
8
6
4
2
0
⫺2
⫺4
⫺6
⫺8
⫺10
⫺12
⫺14
⫺16
⫺18
⫺20
⫺22
⫺24
⫺26
Predicted
(semi-infinite reflector)
Measured
(6 ⫻ 6 m reflector)
0
100
200
300
400
500
600
Frequency (Hz)
700
800
900
1000
800
900
1000
Predicted
(semi-infinite reflector)
Measured
(2 ⫻ 2 m reflector)
0
100
200
300
400
500
600
Frequency (Hz)
R
700
2m
1.2 m
S
14.5 m
0.5 m
Figure 1.61
Comparison of measured and predicted data for the change in the sound pressure level due to the presence of the façade.
Source–receiver–façade geometry is indicated in the sketch. Measured data reproduced with permission from Hopkins and
Lam (2008).
103
S o u n d
I n s u l a t i o n
Measured minus predicted change in SPL
due to the presence of a façade (dB)
15
Finite reflector size (m)
12
2⫻2
3⫻3
9
4⫻4
6
5⫻5
6⫻6
3
0
⫺3
⫺6
⫺9
50
63
80
100 125 160 200 250 315 400 500 630 800 1000
One-third-octave-band centre frequency (Hz)
Figure 1.62
Difference between the measured (finite reflector) and predicted (semi-infinite reflector) change in level due to the presence of
the façade. Façade–receiver distance of 2 m. Measured data are reproduced with permission from Hopkins and Lam (2008).
For protruding or recessed building elements, the spatial variation over the surface can be
affected by a combination of diffraction, shielding, and, within a recess, the existence of a
sound field that partly resembles a two-dimensional reverberant field (sometimes referred to
as a niche effect). It is quite common for windows to be installed in a recess. Figure 1.64 shows
the effect of measuring the surface sound pressure level within a 200 mm deep frame (1 × 1 m)
attached to the surface of a masonry façade (Quirt, 1985). A single measurement within the
frame is seen to be unrepresentative of the average from eight positions. Measurements on a
1.2 × 1.2 m window with recess depths of 120 and 320 mm indicate that the spatial variation
over the surface of a window is larger with a deeper recess (Jonasson and Carlsson, 1986). To
get a more accurate estimate of the average surface sound pressure level, more microphone
positions may be needed with deep recesses (≈300 mm), than with shallower ones (≈100 mm).
1.4.2 Line sources
Façade sound insulation is often assessed using road traffic noise, which can be represented
by a line source. The details of a model for a line source are not discussed here, but the basic
principle involves approximating a line source by a line of closely spaced incoherent point
sources. For a line source comprising many incoherent point sources, air absorption starts to
become significant towards the ends of the line source and therefore needs to be included in the
model (ISO 9613 Part 1). An overview of a suitable spherical wave propagation model for each
point source which incorporates the ground impedance can be found from Attenborough (1988).
To gain a practical insight into the sound field near façades with a line source it is more useful
to look at the statistics of measured data. Hall et al. (1984) took measurements at houses
on 33 different sites with road traffic as the sound source to assess the level measured at
a distance of 2 m from the façade using a microphone on the façade surface (one position
104
Chapter 1
Measured minus predicted change in SPL
due to the presence of a façade (dB)
15
Finite reflector size (m)
12
3⫻3
4⫻4
9
5⫻5
6
6⫻6
3
0
⫺3
⫺6
⫺9
50
63
80
100 125 160 200 250 315 400 500 630 800 1000
One-third-octave-band centre frequency (Hz)
15
Measured minus predicted change in SPL
due to the presence of a façade (dB)
Finite reflector size (m)
12
2⫻2
9
6
3
0
⫺3
⫺6
⫺9
50
63
80
100 125 160 200 250 315 400 500 630 800 1000
One-third-octave-band centre frequency (Hz)
Figure 1.63
Difference between the measured (finite reflector) and predicted (semi-infinite reflector) change in level due to the presence
of the façade. For each reflector size the 11 curves correspond to façade–receiver distance in 0.1 m steps from 1 to 2 m
inclusive. Measured data are reproduced with permission from Hopkins and Lam (2008).
only). The microphone height above ground level was unspecified although it was the same
for both the surface and the 2 m measurement. The results are shown in Fig. 1.65. By assuming
pressure doubling (6 dB) for the façade microphone, the assumption of energy doubling (3 dB)
for the microphone that is 2 m from the façade can be assessed by comparing the difference
between these two microphone positions with a value of 3 dB. In the low-frequency range the
assumption of energy doubling is invalid due to large fluctuations caused by the interference
pattern. In the mid- and high-frequency ranges the assumption of energy doubling is reasonable when we consider the mean of many measurements; however, from the minimum and
maximum values in Fig. 1.65 we see that this assumption is not always valid for an individual
105
S o u n d
I n s u l a t i o n
6
Change in SPL due to frame (dB)
5
Average of eight positions within the frame
4
Example of a single position within the frame
3
2
1
0
⫺1
⫺2
⫺3
⫺4
⫺5
⫺6
100
160
250
400
630
1000 1600 2500
One-third-octave-band centre frequency (Hz)
Figure 1.64
Surface SPL minus
SPL with a façade-receiver distance of 2 m (dB)
Change in the surface sound pressure level on a wall due to the addition of a 200 mm deep frame (1 × 1 m). The loudspeaker
was placed on the ground at a distance of 25 m from the mid-point of the frame, with sound incident upon the surface at an
angle of 60◦ . NB: The angle prescribed for façade insulation measurements with a loudspeaker in ISO 140 Part 5 is 45◦ ± 5◦
rather than 60◦ . Measured data are reproduced with permission from Quirt (1985) and the National Research Council of
Canada.
12
11
10
9
8
7
6
5
4
3
2
1
0
⫺1
⫺2
⫺3
⫺4
Mean with 95%
confidence intervals
Minimum and
maximum levels from
33 different sites
20
31.5
50
80 125 200 315 500 800 1250 2000 3150 5000
One-third-octave-band centre frequency (Hz)
Figure 1.65
Sound pressure levels measured at houses on 33 different sites to assess the level that is measured 2 m from the façade with
road traffic as the sound source. Measurements were made with a microphone on the façade surface, and at a distance of
2 m from the façade. Measured data are reproduced with permission from Hall et al. (1984).
106
Chapter 1
measurement. This presents a problem if we need to accurately convert individual measurements in frequency bands from the 2 m microphone position to a different microphone position
near the façade. This will rarely be possible due to the uncertainty in the many factors that
affect the sound propagation paths. Usually we can only make reasonable estimates when
we want to convert the mean value of many measurements for either frequency bands, or an
A-weighted level.
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Chapter 2
Vibration fields
2.1 Introduction
W
hen airborne sound from the human voice is transmitted from one room to another,
the resulting vibration on the walls and floors in the receiving room is at a sufficiently
low level that it can rarely be perceived with our fingertips. Ears are rather more
sensitive, so if the background noise level in the receiving room is low, and the airborne sound
insulation is quite low too, we can detect the sound pressure radiated by these vibrating walls
and floors. This chapter looks at the theory describing vibration fields and relates it to plates
and beams that are used to build walls and floors in buildings. An understanding of vibration
fields, power input into structures, and sound radiation from vibrating structures is essential to
the study of sound insulation. Compared to sound fields the degree of complexity increases
due to the existence of different wave types as well as a wide range of wavelengths over the
building acoustics frequency range.
The main structural building components for which we need to describe the vibration fields are
beams and plates (see Fig. 2.1). Rooms are bounded by plates in the form of walls and floors;
hence plates play the key role in both sound radiation and structure-borne sound transmission.
Many walls and floors, or their constituent parts, can be represented as solid plates hence these
form a fundamental building element. In practice there are many plates with more complicated
forms, such as walls built from masonry blocks with large internal voids, profiled concrete
floors, and studwork walls; however, the solid plate forms an important benchmark against
which other plates can be assessed. Beams, such as columns and joists, play a minor role in
terms of sound radiation but play an important role in the transmission of structure-borne sound
between plates. Buildings typically contain a variety of solid beams (e.g. concrete columns,
timber joists) along with more complex profiles (e.g. resilient metal channels in lightweight walls
and floors).
2.2 Vibration
To describe the vibration of structures we make use of three vector quantities: (i) displacement
(η), (ii) velocity (v), and (iii) acceleration (a). Their time signals are related to each other by
differentiation or integration using,
η(t) =
v(t) dt
v(t) =
dη(t)
dt
v(t) =
a(t) dt
a(t) =
dv(t)
dt
(2.1)
For a sinusoid or narrow bandwidths with an angular centre frequency, ω, the amplitudes of
these quantities (η̂, v̂, â) are related to each other by
η̂ =
v̂
â
= 2
ω
ω
(2.2)
111
S o u n d
I n s u l a t i o n
(a) Rectangular and circular cross-section beams
z
z
y
y
hy
2r
Lx
hz
Lx
x
x
(b) Plates
z
Lx
x
Ly
h
y
Figure 2.1
General coordinate conventions and dimensions used for beams and plates.
2.2.1 Decibels and reference quantities
The decibel scale (dB) is commonly used to deal with the wide range of values encountered
with vibration. Table 2.1 gives the reference quantities that are used for all figures in the book.
2.3 Wave types
For sound in air and other gases there is only one wave type that needs to be considered, namely longitudinal waves. In contrast, there are three wave types for vibrations on
homogeneous beams and plates in buildings.
For beams the three important wave types are bending, quasi-longitudinal, and torsional waves.
For plates they are bending, quasi-longitudinal, and transverse shear waves. We are primarily
112
Chapter 2
Table 2.1. Vibration – definitions of levels in decibels
Level
Displacement
Velocity#
Acceleration#
Energy
Structure-borne
sound power
Loss factors (Internal,
Definition
η
Ld = 20 lg
η0
where η is the rms
displacement
v
Lv = 20 lg
v0
where v isthe
rms velocity
a
La = 20 lg
a0
where a isthe rms
acceleration
E
LE = 10 lg
E
0
W
LW = 10 lg
W0
Lηii /Lηij /Lηj = 10 lg
Coupling, Total)
Impedance∗
Mobility∗
#
∗
η
η0
Z
Z
0
Y
LY = 20 lg
Y0
LZ = 20 lg
Reference quantity
η0 = 10−12 m
v0 = 10−9 m/s
a0 = 10−6 m/s2
E0 = 10−12 J
W0 = 10−12 W
η0 = 10−12
Z0 = 1 Ns/m
Y0 = 1 m/(Ns)
Indicates use of preferred SI reference quantities for acoustics (ISO 1683).
Applies to real part, imaginary part, or magnitude.
interested in sound radiation from plates, for which bending waves are the most important;
although all three types play a role in structure-borne sound transmission. For brevity, quasilongitudinal and transverse shear waves are sometimes referred to as in-plane waves. It is
not usually necessary to consider in-plane waves for direct sound transmission across a wall,
floor, door, or window. However when flanking transmission involves several connected walls
and floors in the mid- and high-frequency range, in-plane waves start to play an important role.
Compared to sound in gases it is clearly more complicated to have three different wave types
to consider, as well as some differences between the wave motion on beams and plates.
There are two practical options for modelling walls and floors as simple plates over the building
acoustics frequency range; either bending waves are considered to be the only important wave
type (thin and/or thick plate theory), which is the simpler and often perfectly adequate option,
or all three wave types are considered together because quasi-longitudinal and transverse
shear waves always co-exist. However, it is better to avoid the temptation to jump straight
to the section on bending waves. There is a logical flow in the sequential derivation of quasilongitudinal waves, followed by transverse shear waves, and finally, bending waves. In addition
it is easier to grasp the concepts of the different wave types on beams before moving on to
look at plates. It is important to note that the range of wavelengths for structural waves is very
different to sound in air; some examples for different plates are shown in Fig. 2.2.
For a thorough review of structure-borne sound the reader is referred to the book by Cremer
et al. (1973).
113
S o u n d
I n s u l a t i o n
100
Glass: cL ⫽ 5200 m/s, ⫽ 0.24
Plasterboard: cL ⫽ 1490 m/s, ⫽ 0.3
Concrete: cL ⫽ 3800 m/s, ⫽ 0.2
Wavelength (m)
10
1
Bending: 6 mm glass
Bending
0.1
Quasi-longitudinal
12.5 mm plasterboard
Transverse shear
Bending
Quasi-longitudinal
140 mm concrete
Transverse shear
0.01
50
80
125
200
315
500
800
1250 2000
One-third-octave-band centre frequency (Hz)
3150
5000
Figure 2.2
Wavelengths for bending, quasi-longitudinal, and transverse shear waves on different plates (thin plate theory).
2.3.1 Quasi-longitudinal waves
For longitudinal waves, propagation is in the same direction as the particle motion. In Chapter 1
we looked at longitudinal waves for sound in air (or other gases) in rooms and cavities. This
pure form of longitudinal wave motion can also occur in structures too, but only when all the
dimensions of the structure are much larger than the wavelength of the longitudinal wave.
Over the building acoustics frequency range, most beams and plates in buildings are ‘thin’
in terms of their thickness when compared to the wavelength. For this reason the majority
of beams and plates support a different type of longitudinal wave, a quasi-longitudinal wave.
As with pure longitudinal waves propagating in the x-direction, quasi-longitudinal waves have
in-plane displacements in the x-direction, ξ, which give rise to longitudinal strains. However
these displacements not only cause longitudinal strains, but also lateral strains (see Fig. 2.3).
The out-of-plane displacements that are associated with these lateral strains are ζ and η for
the y- and z-directions respectively.
In contrast to bending waves, the lateral strains associated with quasi-longitudinal waves produce small lateral displacements. For this reason, sound radiation from quasi-longitudinal
114
Chapter 2
z
η
(maximum)
Propagation direction
λL
ξ
(maximum)
x
Figure 2.3
Quasi-longitudinal wave on a beam or plate.
waves is usually insignificant compared to bending waves, as is airborne excitation of these
quasi-longitudinal waves from a reverberant sound field (Heckl, 1981). However, quasilongitudinal waves do play an important role in the transmission of structure-borne sound
between connected plates where they can be generated by bending waves that are incident
upon a junction with other plates.
For a quasi-longitudinal wave propagating in the x-direction, the lateral strains, εy and εz ,
resulting from the longitudinal strain, εx , are determined from the Poisson’s ratio of the material,
ν. The general set of equations relating Young’s modulus and Poisson’s ratio to stress and
strain in Cartesian coordinates is
E=
σx − ν (σy + σz )
εx
E=
σy − ν (σx + σz )
εy
(2.3)
σz − ν (σx + σy )
εz
For a beam with a quasi-longitudinal wave propagating in the x-direction there are no constraints on the sides of the beam; hence substituting the conditions σy = 0 and σz = 0 into
Eq. 2.3 gives
σx
(2.4)
E=
εx
E=
and
εy = εz = −νεx
(2.5)
The in-plane displacements and stresses on a beam are shown in Fig. 2.4 for a small rectangular element with a width, dx, for which the longitudinal strain is written in terms of the
displacement, ξ, as
∂ξ
∂x
The stress is related to the displacement using the equation of motion,
∂2 ξ
∂σx
dx − σx = ρ dx 2
σx +
∂x
∂t
εx =
(2.6)
(2.7)
hence,
∂2 ξ
∂σx
=ρ 2
∂x
∂t
(2.8)
115
S o u n d
I n s u l a t i o n
ξ⫹
ξ
∂ξ
dx
∂x
σx
∂σx
dx
∂x
σx ⫹
x
dx
Figure 2.4
Small rectangular element of a beam undergoing quasi-longitudinal wave motion in the x-direction. The position at rest is
shown in dashed lines along with the altered shape and position in solid lines due to the wave motion.
where ρ is the density.
For a beam, the longitudinal force, Fx , on the cross-sectional area of the element, S, is defined
such that a compressive stress, −σx , gives a positive force, and a tensile stress, σx , gives a
negative force:
Fx = −Sσx
(2.9)
It is now possible to determine the wave equation; from Eqs 2.8 and 2.9 we find that
∂Fx
∂2 ξ
= −ρS 2
∂x
∂t
(2.10)
∂2 ξ
∂Fx
= −ES
∂t
∂x∂t
(2.11)
and from Eqs 2.4, 2.6 and 2.9 we have
Equations 2.10 and 2.11 therefore define the wave equation for quasi-longitudinal waves on a
beam as
∂2 ξ
∂2 ξ
E 2 −ρ 2 =0
(2.12)
∂x
∂t
Having established the basic equations for a beam, they can now be adapted to a plate.
For a quasi-longitudinal wave propagating along a plate in the x-direction the cross-section is
constrained by the plate material in the y-direction, but unconstrained in the z-direction, hence
substituting the conditions εy = 0 and σz = 0 into Eq. 2.3 gives
σx
E
=
(1 − ν2 )
εx
where the strain is written in terms of the displacement, ξ, as
∂ξ
εx =
∂x
116
(2.13)
(2.14)
Chapter 2
For a plate it is necessary to reclassify Fx as the longitudinal force per unit width, to give
(2.15)
Fx = −hσx
where h is the plate thickness.
Equations 2.8 and 2.13–2.15 therefore give
and
∂Fx
∂2 ξ
= −ρh 2
∂x
∂t
(2.16)
∂Fx
−Eh ∂2 ξ
=
∂t
(1 − ν2 ) ∂x ∂t
(2.17)
From Eqs 2.16 and 2.17 the wave equation for quasi-longitudinal waves on a plate is
E
∂2 ξ
∂2 ξ
−
ρ
=0
(1 − ν2 ) ∂x 2
∂t 2
(2.18)
For a quasi-longitudinal wave propagating along a beam or plate in the x-direction, the
displacement in the x-direction, ξ, can be described by
ξ(x, t) = ξ̂ exp(−ikx x) exp(iωt)
(2.19)
where ξ̂ is an arbitrary constant, and kx equals the quasi-longitudinal wavenumber, kL = ω/cL .
The phase velocity, cL , for quasi-longitudinal waves is determined by inserting Eq. 2.19 into
the appropriate wave equation for beams or plates. Hence the phase velocity is
E
cL,b =
for beams
(2.20)
ρ
and
cL,p =
E
ρ(1 − ν2 )
for plates
(2.21)
Note that the subscript b for beams and p for plates is used in this chapter to make a distinction
between them when it is considered necessary or helpful; later chapters predominantly discuss
plates so the subscript is dropped and reference is simply made to cL .
For most homogeneous materials used to build walls and floors in buildings, quasi-longitudinal
phase velocities are >1400 m/s (see the Appendix, Table A2 for examples of material properties). We will soon see that the phase velocities for other wave types can be calculated from
the quasi-longitudinal phase velocity; this makes it a very useful property.
The group velocity, cg , is the velocity at which the wave energy propagates across the beam
or plate. For quasi-longitudinal waves where the wavelength is much greater than the beam or
plate thickness, the group velocity cg(L) , is the same as the phase velocity, cL .
2.3.1.1 Thick plate theory
The phase velocity can be assumed to be independent of frequency, i.e. non-dispersive, when
the wavelength is much greater than the plate thickness. However, the dispersive nature
of quasi-longitudinal waves can no longer be ignored in the high-frequency range with thick
masonry/concrete plates that are sometimes used in buildings. An error of X % in the phase
117
S o u n d
I n s u l a t i o n
Quasi-longitudinal wave thin plate limit (Hz)
10 000
9000
8000
7000
6000
5000
Plate thickness
4000
100 mm
3000
200 mm
300 mm
2000
1000
1500
400 mm
2000
2500
3000
3500
4000
Quasi-longitudinal phase speed, cL,p (m/s)
4500
5000
Figure 2.5
Thin plate limits for quasi-longitudinal waves.
velocity can be used to define a limiting frequency, fL(thin) ; this frequency can be considered
as a thin plate limit above which the plate can no longer be considered as a thin plate for
quasi-longitudinal waves (Cremer et al., 1973).
!
" X%
cL,p "
# 100
fL(thin) =
(2.22)
π2
ν 2
h
6
1−ν
For an error of 3% and a Poisson’s ratio of 0.3, this corresponds to the frequency at which
λL ≈ 3h.
Figure 2.5 shows the thin plate limit for a common range of phase velocities, assuming an
error of 3% and a Poisson’s ratio of 0.2. The thin plate limits tend to be in or above the highfrequency range. Due to other aspects which limit accurate prediction of structure-borne sound
transmission on building elements (particularly at high frequencies where many plates cannot
be considered as homogeneous) it is usually sufficient to use thin plate theory over the entire
building acoustics frequency range.
2.3.2 Transverse waves
Solids in the form of beams and plates can support transverse waves due to their ability to
support shear stresses. On beams these waves are referred to as torsional waves, whereas
those on plates are referred to as transverse shear waves.
2.3.2.1 Beams: torsional waves
Torsional waves are generated in beams where a time-varying moment is applied via an axis
that passes through the axis of the beam. As a torsional wave propagates along a beam in the
x-direction, the cross-sections rotate about the axis of the beam by an angle, θ. It is therefore
118
Chapter 2
x
Propagation
direction
λT
Figure 2.6
Torsional wave on a beam.
z
ζ
η
hz
y
θ
hy
Figure 2.7
Cross-section of a rectangular beam in the yz plane undergoing torsional wave motion. The position at rest is shown in dashed
lines along with the altered shape and position in solid lines due to the wave motion. Example displacements (exaggerated)
are indicated by for one point on the surface of the beam.
apparent that for beams with circular cross-sections there will be no displacement outside of the
cross-section. These waves are easier to visualize for a beam with a rectangular cross-section
as shown in Fig. 2.6. For rectangular cross-sections the displacements in the y- and z-directions
are ζ and η respectively (see Fig. 2.7), and vary depending on the distance from the axis,
ζ = −θz
η = θy
(2.23)
119
S o u n d
I n s u l a t i o n
The moment about the x-axis, Mx , for any shape of cross-section is (Cremer et al., 1973)
Mx = T
∂θ
∂x
(2.24)
This moment is used to determine the wave equation that governs the propagation of torsional
waves in the x-direction and is defined as (Cremer et al., 1973)
∂2 θ
ρIθ ∂2 θ
−
=0
2
∂x
GJ ∂t 2
(2.25)
where Iθ is the polar moment of inertia (per unit length of the beam) about the x-axis, J is the
torsional moment of rigidity, and G is the shear modulus of the material which is related to
Young’s modulus by
G=
E
2(1 + ν)
(2.26)
For a torsional wave propagating along a beam in the x-direction, the angular torsional
displacement, θ, can be described by
θ(x, t) = θ̂ exp(−ikx x) exp(iωt)
(2.27)
where θ̂ is an arbitrary constant, and kx equals the torsional wavenumber, kT = ω/cT,b .
Substitution of Eq. 2.27 into the wave equation gives the phase velocity as
GJ
T
=
cT,b =
ρIθ
ρIθ
(2.28)
The product GJ equals the torsional stiffness, T . The product ρIθ is sometimes referred to as
the mass moment of inertia about the x-axis.
For a solid beam with a circular cross-section of radius, r , the torsional stiffness is
T =G
πr 4
2
(2.29)
and the polar moment of inertia about the x-axis is
Iθ =
πr 4
2
(2.30)
For a solid beam of rectangular cross-section, the torsional stiffness can be calculated using
(Timoshenko and Goodier, 1970)
T =G
hz hy3
3
1−
192hy
πhz
tanh
π 5 hz
2hy
(2.31)
where hz ≥ hy , and the polar moment of inertia about the x-axis is
Iθ =
(hy hz3 + hz hy3 )
12
(2.32)
Torsional waves are non-dispersive, so the group velocity, cg(T) , is the same as the phase
velocity, cT,b .
120
Chapter 2
2.3.2.2
Plates: transverse shear waves
For transverse shear waves propagating in the x-direction, the in-plane displacement, ζ, takes
place in the y-direction, i.e. perpendicular to the direction of propagation (see Fig. 2.8). Because
the only motion of the plate surface is tangential to the adjacent air (or other gas) these waves
are not able to radiate sound, or to be excited by airborne sound that is incident upon the
plate. So, as with quasi-longitudinal waves, their primary role is in structure-borne sound
transmission.
Transverse motion is assessed using a small rectangular element on the plate which lies in
the xy plane (see Fig. 2.9) with an area of dxdy (Cremer et al., 1973). For a wave propagating
in the x-direction there will be different displacements on the two sides of the element that are
parallel to the y-axis, namely ζ and ζ + (∂ζ/∂x)dx. These displacements cause the element to
change shape from its rectangular form to that of a parallelogram. Therefore the sides of the
element that were parallel to the x-axis at rest have moved through the small angle, γxy , (where
tan γxy ≈ γxy ) although the volume of the plate element has remained the same. This angle is
Propagation direction
y
ζ
(maximum)
x
λT
Figure 2.8
Transverse shear wave on a plate.
y
τyx⫹
∂τyx
dy
∂y
τxy⫹
∂τxy
dx
∂x
τxy
dy
γxy
τyx
ζ⫹
∂ζ
dx
∂x
ζ
x
dx
Figure 2.9
Small rectangular element in the xy plane of a plate undergoing transverse shear wave motion. The position at rest is shown
in dashed lines along with the altered shape and position in solid lines due to the wave motion.
121
S o u n d
I n s u l a t i o n
referred to as the shear strain and is related to the displacement by
γxy =
∂ζ
∂x
(2.33)
The displacements result in shear stresses: τxy (a stress acting in the y-direction on the plane
that lies perpendicular to the x-axis, i.e. the yz plane) and τyx (a stress acting in the x-direction
on the plane that lies perpendicular to the y-axis, i.e. the xz plane). These shear stresses are
of equal magnitude when the displaced element is at equilibrium in terms of the moments that
act upon the yz and xz planes; they are also proportional to the shear strain, hence
(2.34)
τxy = τyx = Gγxy
where G is the shear modulus, the ratio of the shear stress to the shear strain. The shear
modulus is related to Young’s modulus by
E
2(1 + ν)
(2.35)
∂τxy
∂2 ζ
=ρ 2
∂x
∂t
(2.36)
G=
The equation of motion for the element is
The wave equation governing the propagation of transverse shear waves in the x-direction can
now be found from Eqs 2.33, 2.34, and 2.36 as
G
∂2 ζ
∂2 ζ
−ρ 2 =0
2
∂x
∂t
(2.37)
For a transverse shear wave propagating along a plate in the x-direction, the displacement in
the y-direction, ζ, can be described by an equation of the form
ζ(x, t) = ζ̂ exp(−ikx x) exp(iωt)
(2.38)
where ζ̂ is an arbitrary constant, and kx equals the transverse shear wavenumber, kT = ω/cT,p .
The phase velocity, cT,p for transverse shear waves is determined by inserting Eq. 2.38 into
the wave equation; hence it is related to the shear modulus, which can be calculated from the
quasi-longitudinal phase velocity,
cT,p =
G
=
ρ
E
= cL,p
2ρ(1 + ν)
1−ν
2
(2.39)
Transverse shear waves are non-dispersive hence the group velocity cg(T) , is the same as
the phase velocity, cT,p . Plates that are thin compared to the wavelength do not affect the
transverse shear waves which propagate in the plane of the plate. So, unlike bending and
quasi-longitudinal waves there is no limiting frequency for the phase velocity relating to the
assumption of a thin plate where the wavelength is large compared to the plate thickness.
122
Chapter 2
Propagation direction
η
(maximum)
z
x
λB
Figure 2.10
Bending wave on a beam or plate.
2.3.3 Bending waves
Pure bending waves occur where the bending wavelength is large compared to the beam or
plate thickness; hence these waves only occur on ‘thin beams’ and ‘thin plates’. A propagating
bending wave causes both rotation and lateral displacement of the beam or plate elements
(see Fig. 2.10). When compared with in-plane waves, bending waves have large lateral
displacements; hence they play the main role in the radiation of sound.
The wave equation for a beam is determined by considering a bending wave propagating in
the x-direction (Cremer et al., 1973). For the purpose of this derivation the beam is positioned
so that the x-axis runs along the centroid of the cross-section. We will choose the z-direction
as the direction in which the lateral displacement occurs but note that a beam of rectangular
cross-section can support a bending wave in both the y- and z-directions. Figure 2.11 shows
that a small rectangular element on the beam will undergo lateral displacement, η, in the
z-direction as well as rotation through the small angle, β (where tan β ≈ β). Hence we need to
consider the lateral velocity, vz = ∂η/∂t, in the z-direction, as well as the angular velocity, ωy ,
about the y-axis. The former is defined as positive in the positive z-direction, and the latter is
positive for anti-clockwise rotation. The angle of rotation is related to the displacement by
β=
∂η
∂x
(2.40)
from which the angular velocity is related to the lateral velocity by
ωy =
∂2 η
∂vz
∂β
=
=
∂t
∂t ∂x
∂x
(2.41)
In comparing the element before and after bending deformation we see that the line on the
upper surface is shortened due to compression (i.e. AD becomes A′ D′ ), and that the line on
the lower surface is elongated due to tension (i.e. BC becomes B′ C′ ). Between the upper and
lower surfaces there is a neutral axis along which there is neither compression nor tension;
therefore the element length, dx, along this axis is unchanged. The magnitude of the stresses
and strains are zero on the neutral axis and linearly increase towards the largest values on the
upper and lower surfaces of the element (see Fig. 2.12).
For a homogeneous beam of rectangular or circular cross-section, the neutral axis lies at the
mid-height of the beam. The neutral axis on the deformed element lies on an arc of radius, R0 ,
hence
dx = R0 θ
(2.42)
123
S o u n d
I n s u l a t i o n
z
R0
β
θ
A⬘
D⬘
β⫹θ
tral
Neu
axis
β
B⬘
C⬘
η
η⫹
D
∂η
dx
∂x
A
Neutral
axis
C
x
B
dx
Figure 2.11
Small rectangular element of a beam undergoing bending wave motion with displacement in the z-direction. The position at
rest is shown in dashed lines along with the altered shape and position in solid lines due to the wave motion.
124
Chapter 2
θ
z
z
Displacement on a surface
that lies a distance, z, from the
neutral axis, resulting in a
strain, εx = –zθ/dx
Stress, σx, on a surface
that lies a distance, z,
from the neutral axis
A⬘ A
Compression
(positive z)
z
z
x
Tension
(negative z)
B
B⬘
dx
Figure 2.12
Displacements (left) and stresses (right) on the beam element using line AB as an example.
It is assumed that plane surfaces in the beam remain plane when they are rotated by the small
angle, β. This important assumption means that lines such as AB or CD on the element at
rest, which are perpendicular to the neutral axis, will remain normal to the neutral axis on the
element during bending wave motion. Therefore lines AB and CD rotate in opposite directions
such that the resulting angle, θ, between A′ B′ and C′ D′ on the deformed element is
θ=
∂β
∂2 η
dx = 2 dx
∂x
∂x
(2.43)
and from Eq. 2.42, the radius of the neutral axis is related to the lateral displacement by
1
θ
∂2 η
=
= 2
R0
dx
∂x
(2.44)
Hence we can describe the strain in the x-direction on any surface that lies a distance, z, from
the neutral axis (see Fig. 2.12) using
εx =
[(R0 − z)θ] − R0 θ
z
∂2 η
=−
= −z 2
R0 θ
R0
∂x
(2.45)
The accompanying stress for the unconstrained cross-section of a beam is calculated using
the conditions σy = 0 and σz = 0 in Eq. 2.3, and using Eq. 2.45 for the strain to give
σx = Eεx = −Ez
∂2 η
∂x 2
(2.46)
125
S o u n d
I n s u l a t i o n
The stresses from compression and tension result in a bending moment, My about the
y-axis. This moment rotates in the same direction as the angular velocity, ωy . The individual moments that result in the bending moment are the product of the surface stress and the
distance of the surface, z, from the neutral axis (see Fig. 2.12). This bending moment is calculated by summing up the individual moments over the cross-sectional area of the element,
S, using
∂2 η
My = σx z dS = −EIb 2
(2.47)
∂x
S
where the moment of inertia for a beam, Ib , is a measure of its resistance to bending. The
moment of inertia is defined about an axis. In this case we need the moment of inertia about
the y-axis which is
Ib = z 2 dS
(2.48)
S
To relate the bending moment to the angular velocity, Eq. 2.47 is differentiated with respect
to time,
∂My
∂ωy
∂3 η
= −EIb
= −EIb
∂t
∂t ∂x 2
∂x
for which the bending stiffness of a solid beam, Bb , is defined as
2
Ib
Bb = EIb = ρcL,b
(2.49)
(2.50)
For solid circular section beams, Ib = π(2r )4 /64, where r is the radius. Solid beams with rectangular sections can support bending waves in both the y- and z-directions. Therefore it is
necessary to identify the direction of the lateral displacement because the moment of inertia
depends upon the orientation of the beam. For lateral displacement in the z-direction as we
have used in this derivation, the moment of inertia of the cross-sectional area about the y-axis
is Ib = hy hz3 /12; whereas for lateral displacement in the y-direction the moment of inertia about
the z-axis would be Ib = hz hy3 /12.
In addition to the stresses within the beam element due to compression and tension, there will
also be shear stresses due to a shear force, Fz , acting in the z-direction (see Fig. 2.13). The
shear force is found in terms of the lateral displacement using the equation of motion
∂Fz
∂2 η
F z − Fz +
dx = ρl dx 2
(2.51)
∂x
∂t
which gives
∂2 η
∂Fz
= −ρl 2
∂x
∂t
(2.52)
where ρl is the mass per unit length.
Balancing the shear forces and moments on the element at equilibrium (see Fig. 2.13) yields,
∂My
Fz dx = My − My +
dx
(2.53)
∂x
hence the shear force is related to the bending moment by
Fz = −
126
∂My
∂x
(2.54)
Chapter 2
z
vz
Fz
ωy
My ⫹
My
Fz ⫹
∂Fz
∂x
∂My
∂x
dx
dx
x
dx
Figure 2.13
Moments, shear forces, angular and transverse velocities for the beam element undergoing bending wave motion.
Note that this equilibrium equation has ignored the rotatory inertia of the element. This is a
valid assumption for pure bending wave motion on thin beams and thin plates (Cremer et al.,
1973). For thick beams the rotatory inertia is accounted for by replacing Fz with Fz + ρIb (∂ωy /∂t)
in Eq. 2.54.
The one-dimensional wave equation for bending waves on a thin homogeneous beam is
determined from Eqs 2.49, 2.52, and 2.54,
Bb
∂4 η
∂2 η
+ ρl 2 = 0
4
∂x
∂t
(2.55)
The bending wave equation is different to the other wave equations in that it contains a fourth
order (rather than second order) derivative term for the displacement. This results in a more
complex solution to the wave equation. For a sinusoidal bending wave propagating along a
beam in the x-direction we initially describe the displacement in the z-direction by an equation
of the form
η(x, t) = η̂ exp(−ikx x) exp(iωt)
(2.56)
where η̂ is an arbitrary constant and kx is a constant relating to the wavenumber.
Substituting Eq. 2.56 into the wave equation for the beam gives the bending wavenumber,
kB , as
kx4 = kB4 =
ω 2 ρl
Bb
(2.57)
127
S o u n d
I n s u l a t i o n
which has four roots: +kB , −kB , +ikB , and −ikB . Hence for bending wave motion in the
x-direction the general solution for the displacement is written in the form
η(x, t) = [η̂+ exp(−ikB x) + η̂− exp(ikB x) + η̂n− exp(−kB x) + η̂n+ exp(kB x)] exp(iωt)
(2.58)
where η̂+ , η̂− , η̂n− , and η̂n+ are constants.
The first and second terms within the square brackets represent a bending wave propagating
in the positive and the negative x-directions respectively. The third and fourth terms represent
vibration fields called nearfields that decay away exponentially in the positive and negative
x-directions respectively. An exponential decay means that the nearfield contribution to the
overall displacement is usually only significant at positions close to the point at which the
nearfields are generated.
From Eq. 2.57, the phase velocity for the propagating bending waves on a solid beam is
2 2 2
2
4 4π f cL,b Ib
4 ω Bb
cB,b =
(2.59)
=
ρl
S
In contrast to the phase velocity for quasi-longitudinal waves on thin plates, transverse shear
waves on plates, and longitudinal sound waves in air, the phase velocity for bending waves is
frequency-dependent. Bending waves are therefore described as dispersive. The dispersive
nature of bending waves means that the group velocity, cg(B),b , at which bending wave energy
travels is not the same as the phase velocity and is calculated from
cg(B),b =
dω
= 2cB,b
dk
(2.60)
The one-dimensional wave equation for a thin beam can be adapted to apply to a bending wave
propagating in the x-direction on a thin homogeneous isotropic plate. As we have already seen
with quasi-longitudinal waves on plates, E must be replaced by E/(1 − ν2 ) because, unlike the
beam, the plate element cross-section is constrained by the material in the y-direction, hence
Eq. 2.46 describing the internal stresses becomes
σx =
Eεx
−Ez ∂2 η
=
2
1−ν
1 − ν2 ∂x 2
(2.61)
and the bending stiffness for a thin homogeneous isotropic plate, Bp , is defined as
Bp =
2
ρcL,p
h3
EIp
Eh3
=
=
1 − ν2
12(1 − ν2 )
12
(2.62)
where Ip is the moment of inertia per unit width, h3 /12.
Equation 2.52 is adapted by temporarily reclassifying Fz as the shear force per unit width of
the plate, and by replacing the mass per unit length with the mass per unit area, ρs , to give
∂Fz
∂vz
∂2 η
= −ρs
= −ρs 2
∂x
∂t
∂t
(2.63)
Hence, the one-dimensional wave equation for bending waves propagating in the x-direction
on a thin homogeneous plate is
Bp
128
∂4 η
∂2 η
+
ρ
=0
s
∂x 4
∂t 2
(2.64)
Chapter 2
z
dx
x
dy
Qy
η
h
Myy
Qx
y
Mxy ⫹
∂Mxy
∂x
dx
Myx
Mxx ⫹
Mxx
Myx ⫹
∂Mxx
∂x
ξ
dx
ωx
∂Myx
∂y
dy
Mxy
Myy ⫹
∂Myy
∂y
dy
Qx ⫹
ωy
Qy ⫹
∂Qx
∂x
dx
∂Qy
∂y
dy
ζ
Figure 2.14
Bending and twisting moments per unit width, shear forces per unit width, displacements and angular velocities for an element
on a plate undergoing bending wave motion.
The phase velocity on a thin plate is determined using the same approach as for beams, where
the wavenumber is
ω 2 ρs
kB4 =
(2.65)
Bp
which gives the phase velocity as
2
2 2 2
2πfhcL,p
4 ω Bp
4 4π f h E
cB,p =
=
=
√
ρs
12ρ(1 − ν2 )
12
(2.66)
The group velocity, cg(B),p , which describes the transport of bending wave energy is determined
from Eq. 2.65, where
dω
cg(B),p =
= 2cB,p
(2.67)
dkB
In practice there will be bending waves on a plate propagating in the x- and y-directions
simultaneously. Therefore we need to derive the two-dimensional wave equation for a thin
plate that resists bending deformation about the x- and y-axes and has a twisting moment. We
start by positioning the plate so that the x-axis is aligned at mid-height of the cross-section.
A rectangular element in the xy plane of the plate with the dimensions, dx by dy, undergoes
displacements, ξ, ζ, and η in the x-, y- and z-directions respectively. For the element in
equilibrium the moments and shear forces per unit width are shown in Fig. 2.14.
129
S o u n d
I n s u l a t i o n
The shear forces per unit width are now defined as Qx and Qy . The bending moments per
unit width are Mxy and Myx where the moment acts perpendicular to the axis denoted by the
first subscript letter, in the direction denoted by the second subscript letter. The twisting (or
torsional) moments per unit width are Mxx and Myy . The direction of rotation for the moments
is found by taking a viewpoint looking directly into the arrowhead, along the line of the arrow;
the rotation from this viewpoint is in an anti-clockwise direction. There are now two angular
velocities: ωx , about the x-axis, and ωy , about the y-axis, defined as positive in the anticlockwise direction as shown in Fig. 2.14. We have already formulated ωy in Eq. 2.41 for an
angle of rotation, β, in the anti-clockwise direction about the y-axis. To determine ωx the plate
element rotates anti-clockwise about the x-axis by an angle, α, such that
∂η
∂y
(2.68)
∂2 η
∂vz
∂α
=−
=−
∂t
∂t ∂y
∂y
(2.69)
α=−
and the angular velocity is
ωx =
To find the bending moments it is necessary to determine the stresses. The stress–strain
relationships are found by substituting the condition σz = 0 into Eq. 2.3 to give
E=
σx − νσy
εx
and
E=
σy − νσx
εy
(2.70)
from which we obtain each stress in terms of the strains
σx =
E
(εx + νεy )
1 − ν2
and σy =
E
(εy + νεx )
1 − ν2
(2.71)
These strains can now be written in terms of the lateral displacement, η, using the same
approach as for a beam element (see Eq. 2.45) as
εx = −z
∂2 η
∂x 2
and
εy = −z
∂2 η
∂y 2
(2.72)
to give the stresses
σx = −
Ez
1 − ν2
∂2 η
∂2 η
+
ν
∂x 2
∂y 2
and σy = −
Ez
1 − ν2
∂2 η
∂2 η
+
ν
∂y 2
∂x 2
(2.73)
To find the twisting moments we need to determine the shear stress, τxy , in terms of the lateral
displacement. The shear stress is determined from the shear strain,γxy , which is a result of the
in-plane displacements as shown in Fig. 2.15. The sides of the element at rest that were parallel
to the x- and y-axes have moved through the small angles, ∂ζ/∂x and ∂ξ/∂y respectively. The
sum of these angles equals the shear strain, therefore the shear stress is
∂ξ
∂ζ
(2.74)
τxy = Gγxy = G
+
∂y
∂x
From Eq. 2.72 we already have the strains in terms of the lateral displacement; hence these
can be equated to the following strains in terms of the in-plane displacement,
εx =
130
∂ξ
∂x
and εy =
∂ζ
∂y
(2.75)
Chapter 2
y
ξ⫹
∂ξ
∂y
dy
∂ξ
∂y
dy
∂ζ
ξ
∂x
ζ
ζ⫹
∂ζ
∂x
dx
x
dx
Figure 2.15
Small rectangular element in the xy plane of a plate undergoing shear deformation as part of bending wave motion. The
position at rest is shown in dashed lines along with the altered shape and position in solid lines due to shear deformation.
for which the appropriate differentiation and integration gives
∂ξ
∂ζ
∂2 η
=
= −z
∂y
∂x
∂x ∂y
(2.76)
The shear stress can now be written in terms of E and ν by using Eq. 2.35 for the shear modulus
to give
τxy = −
Ez(1 − ν) ∂2 η
Ez ∂2 η
=−
1 + ν ∂x ∂y
1 − ν2 ∂x ∂y
(2.77)
We now assume that the shear strains, γxz and γyz are so small as to be negligible. This is
consistent with the assumption that any point on the element which is normal to the neutral
axis when at rest, remains normal to the neutral axis after bending deformation.
The bending moments per unit width are calculated by summing up the individual moments
over the plate thickness,
Mxy =
h/2
−h/2
σx z dz
and Myx = −
h/2
σy z dz
(2.78)
−h/2
The rotation of Myx about the x-axis is defined as being in the same direction as the rotation of
the vertical edge of the plate element that lies in the xz plane due to compression and tension.
This is opposite to the convention used for Mxy , and therefore a negative sign is included in
the equation for Myx above.
131
S o u n d
I n s u l a t i o n
The moment of inertia for a plate is,
Ip =
h/2
z 2 dz =
h3
12
(2.79)
−h/2
hence these moments are
∂2 η
∂2 η
+
ν
∂x 2
∂y 2
2
∂2 η
∂ η
+
ν
= Bp
∂y 2
∂x 2
Mxy = −Bp
and
Myx
(2.80)
(2.81)
A similar approach is used to find the twisting moments from the shear stress, where τxy = τyx .
Taking account of the different moment directions these are
h/2
∂2 η
Mxx = −Myy = −
(2.82)
τxy z dz = Bp (1 − ν)
∂x ∂y
−h/2
The two equilibrium equations for the moments of all forces acting on the element about the
y-axis and the x-axis are
∂Mxy
∂Qx
dx (x + dx) dy + Mxy − Mxy +
dx dy
Qx x − Qx +
∂x
∂x
(2.83)
∂Myy
dy dx = 0
+ Myy − Myy +
∂y
∂Myx
∂Qy
dy (y + dy) − Qy y dx + Myx − Myx +
dy dx
Qy +
∂y
∂y
∂Mxx
(2.84)
+ Mxx − Mxx +
dx dy = 0
∂x
For the shear force terms, the plate is orientated in the coordinate system so that x = y = 0, and
the term involving ∂Qx /∂x can be assumed to be negligible. Hence, substituting the moments
(Eqs 2.80, 2.81, and 2.82) into the above equations results in
∂Myy
∂Mxy
∂ ∂2 η
∂2 η
−
= Bp
+
Qx = −
(2.85)
∂x
∂y
∂x ∂x 2
∂y 2
∂Myx
∂Mxx
∂ ∂2 η
∂2 η
+
= Bp
+
Qy =
(2.86)
∂y
∂x
∂y ∂x 2
∂y 2
For thin plates, as with thin beams, it is appropriate to ignore the rotatory inertia of the element
in these equilibrium equations. For thick plates, rotatory inertia is accounted for by replacing
Qx with Qx + ρIp (∂ωy /∂t) in Eq. 2.85 and replacing Qy with Qy + ρIp (∂ωx /∂t) in Eq. 2.86.
The shear force is now described in terms of the lateral displacement using the equation of
motion,
∂Qy
∂Qx
∂2 η
−
dxdy −
dydx = ρs dx dy 2
(2.87)
∂x
∂y
∂t
The two-dimensional wave equation for bending waves on a thin homogeneous isotropic plate
is now given by substituting Eqs 2.85 and 2.86 into Eq. 2.87 to yield
4
∂4 η
∂4 η
∂2 η
∂ η
+
2
+
+ ρs 2 = 0
Bp
(2.88)
4
2
2
4
∂x
∂x ∂y
∂y
∂t
132
Chapter 2
In Eq. 2.58 a general solution was derived for bending wave motion in the x-direction which
involved both bending waves and nearfields. In a similar way to the representation of sound
fields using plane waves, we can assume a plane wave field for bending waves on a plate
without nearfields. The displacement can therefore be described by an equation of the form
η(x, y, t) = η̂ exp(−ikx x) exp(−iky y) exp(iωt)
(2.89)
where η̂ is an arbitrary constant, and kx and ky are constants relating to the wavenumber.
The relationship between the wavenumber and the constants, kx and ky is determined by
inserting Eq. 2.89 into the wave equation, which gives
kB2 = kx2 + ky2
(2.90)
In a similar way to sound fields in rooms and cavities, Eq. 2.90 can be used to determine the
plate modes.
2.3.3.1 Thick beam/plate theory
At frequencies where the bending wavelength is not large compared to the beam or plate
thickness, pure bending waves no longer occur and account needs to be taken of the shear
deformation and rotatory inertia that occurs with thick plates (Mindlin, 1951). For thin plates
it is assumed that plane surfaces in the plate remain plane when they are rotated through
the small angles, α and β. For thin beams this only applies to the angle, β. This assumption
becomes invalid as the bending wavelength decreases with increasing frequency due to shear
deformation of the cross-section of the element. Referring back to the beam element in Fig. 2.11
this shear deformation means that lines A′ B′ and C′ D′ in the deformed element are no longer
straight lines. For homogeneous beams and plates the effect of shear deformation tends to be
more significant than rotatory inertia (Cremer et al., 1973).
The theory for thick beams and plates is more complex than that for thin beams or plates.
In addition, it does not allow simple calculation of the mode frequencies, although estimates
can be made for the modal density of thick beams and plates (Lyon and DeJong, 1995).
This complicates matters for masonry/concrete structures because we will soon see that the
crossover frequency from thin to thick beam/plate theory usually occurs in the building acoustics
frequency range. Hence it is not always possible to justify using pure bending wave theory
across the entire frequency range. This transition from thin to thick can be observed in practice
when measuring the modal response of a single beam or plate when it is uncoupled from other
structural elements. However a building only consists of coupled beams and plates, and the
point at which it is essential to switch from thin to thick beam/plate theory in the prediction of
sound transmission is not clear-cut. For some masonry/concrete buildings, thin plate theory
can be used across the building acoustics frequency range without incurring errors that are
larger than those from other assumptions. The simplest approach is to use thin beam/plate
theory and always calculate a crossover frequency to indicate the frequency above which there
is increased uncertainty in the use of thin plate theory.
For beams and plates, a general rule for the validity of pure bending wave motion is that
the bending wavelength must be much larger than the thickness. We will focus on thick plates
(rather than thick beams) as these tend to be more important in determining the sound insulation
in many buildings. A specific crossover frequency for plates can be calculated by considering
the percentage difference in the phase velocity between pure bending waves on thin plates,
133
S o u n d
I n s u l a t i o n
2800
Bending wave
thin plate limit (Hz)
2600
Plate thickness
2400
100 mm
2200
200 mm
2000
300 mm
1800
400 mm
1600
1400
1200
1000
800
600
400
200
1500
2000
2500
3000
3500
4000
Quasi-longitudinal phase speed, cL,p (m/s)
4500
5000
Figure 2.16
Thin plate limits for bending waves.
and bending waves on thick plates (Cremer et al., 1973). This frequency is described as the
thin plate limit for pure bending waves, fB(thin) , and can be calculated for an X % difference in
phase velocity using
2.4(1 + ν) −1
X % 21.6cL,p
1
+
(2.91)
fB(thin) =
100 π2 h
1 − ν2
Cremer et al. (1973) propose that the bending wave correction terms for shear deformation
and rotatory inertia can be important when the difference in the phase velocity is >10%. This
corresponds to the frequency at which λB = 6h. Most building materials have a Poisson’s ratio
between 0.2 and 0.3, hence the thin plate limit can be estimated using
fB(thin) ≈
0.05cL,p
h
(2.92)
The frequency at which λB = 6h can be used to define a thin beam limit, hence Eq. 2.92
also applies to rectangular section beams where h corresponds to the direction of lateral
displacement, hy or hz , and cL,p is replaced with cL,b .
The thin plate limit for a single sheet of 12.5 mm plasterboard is above the building acoustics
frequency range. However, for masonry/concrete walls and floors it is quite common for the
thin plate limit to occur in the mid- or high-frequency range; the thin plate limit for a common
range of quasi-longitudinal phase velocities is shown in Fig. 2.16 for X = 10% and a Poisson’s
ratio of 0.2.
The bending wave equation can be modified for thick plates to include terms that account for
shear deformation and rotatory inertia (Mindlin, 1951). The bending wave effectively becomes
a combination of pure bending waves and transverse shear waves. In practice we need to
calculate the phase velocity and group velocity across a frequency range in which the plate
is either described as thin or thick depending upon the frequency. To cover such a frequency
134
Chapter 2
range with a single smooth curve, a single phase velocity, cB,thick,p , is given by (Rindel, 1994)
cB,thick,p =
1
3
cB,p
1
+ 3 3
γ cT,p
− 13
(2.93)
where γ is defined for different values of the Poisson’s ratio as γ = 0.689 for ν = 0.2, and
γ = 0.841 for ν = 0.3 (Craik, 1996a).
The corresponding group velocity is (Craik, 1996)
cg(B,thick),p =
2
cB,thick,p
2
cB,p
cg(B),p
+
2
cB,thick,p
3
γ 3 cT,p
−1
(2.94)
2.3.3.2 Orthotropic plates
For bending wave motion on a homogeneous plate it is necessary to establish whether the plate
should be classified as isotropic (uniform stiffness), orthotropic (different stiffness in the x- and
y-directions – assuming that the axes of stiffness align with the plate axes), or anisotropic
(different stiffness in multiple directions). The majority of homogeneous plates in buildings can
be considered as isotropic or orthotropic. The transmission of sound and vibration involving
orthotropic plates is more complex than with isotropic plates, but if the properties in the two
orthogonal directions are not very different, it is often possible to treat them as isotropic plates.
The wave equation for bending waves on a thin homogeneous orthotropic plate is (Cremer
et al., 1973)
Bp,x
∂4 η
∂4 η
∂4 η
∂2 η
+ 2Bp,xy 2 2 + Bp,y 4 + ρs 2 = 0
4
∂x
∂x ∂y
∂y
∂t
(2.95)
where Bp,x and Bp,y are the bending stiffness in the x- and y-directions respectively, and
Bp,xy =
2Gxy h3
νxy Exy h3
+
12
12
(2.96)
The elastic properties for orthotropic materials needed to calculate Bp,xy (Eq. 2.96) are not
always known. However, Bp,xy is approximately equal to the geometric mean of the bending
stiffness in the two orthogonal directions. This is referred to as the effective bending stiffness,
Bp,eff , and is given by (Cremer et al., 1973)
Bp,xy ≈ Bp,eff =
Bp,x Bp,y
(2.97)
Use of the effective bending stiffness often allows an orthotropic plate to be approximated by
an isotropic plate. Equation 2.62 can then be used to calculate an effective quasi-longitudinal
phase velocity, cL,eff , if required for other calculations.
Masonry blocks or bricks are usually rectangular in cross-section; hence a wall will have significantly different numbers of joints per metre in the horizontal and vertical directions. For solid
block/brick walls with both the horizontal and vertical joints mortared, the bending stiffness is
often up to 20% higher in the horizontal direction than the vertical direction. However, in most
calculations of airborne and structure-borne sound transmission it is reasonable to treat these
plates as isotropic.
135
S o u n d
I n s u l a t i o n
Beam component
Beam components forming a profiled plate
dR
Figure 2.17
Profiled plate and one of its constituent beams, where the repetition distance is dR .
For flat plates (isotropic or orthotropic) it is appropriate to refer to either the bending stiffness
or the quasi-longitudinal phase velocity, but for profiled, corrugated or ribbed plates it is only
appropriate to refer to the bending stiffness in the two orthogonal directions. The bending
stiffness in the two orthogonal directions for profiled, corrugated and ribbed plates is calculated
in the following sections on the assumption that the bending wavelength is much larger than
the profile, corrugation, or rib spacing.
2.3.3.2.1
Profiled plates
Many orthotropic plates have profiles or corrugations running in one direction. When viewed in
cross-section, there is usually a single profile that repeats at regular intervals with a repetition
distance, dR . Hence a plate can effectively be built-up from beams that have this profile, where
the beams are connected along their lengths and each beam has a width, dR (see Fig. 2.17).
To determine the bending stiffness of such a plate we need to determine the moment of inertia
for one of these beams. We will continue to use the coordinate conventions in Fig. 2.1, hence
the cross-section of each beam lies in the yz plane with its length running along the x-axis.
This means we need to determine the moment of inertia about the y-axis. For many profiled
plates, such as steel cladding, the cross-section of each beam can be represented by six solid
homogenous rectangular elements (Bies and Hansen, 1988). A generalized cross-section for
a beam formed by N solid homogenous rectangular elements (for which N ≥ 2) is shown in
Fig. 2.18. Although rectangular elements are most commonly used to form the cross-section
(see Fig. 2.19), other shapes of element could also be used in the following analysis.
For any shape of element, the moments of inertia about the y- and z-axes (Iy and Iz ,
respectively) are
(2.98)
Iy = z 2 dS and Iz = y 2 dS
S
S
These are centroidal moments of inertia and are calculated using the axes convention shown in
Fig. 2.20. Hence for a rectangular element (base length, b, and height, h), Iy can be calculated
from Eq. 2.98 yielding the well-known equation,
Iy =
h/2 b/2
−h/2 −b/2
136
z 2 dy dz =
bh3
12
(2.99)
Chapter 2
Z
b3
θ3
b2
b4
θ2
z3
b1
z2
z0
θ1
Neutral
axis
z4 θ4
b5
z1
z5
θ6 = 0°
b6
θ5
z6 = h/2
Y
dR
Figure 2.18
Generalized profiled beam section formed from rectangular elements (N = 6) of thickness, h, with a repetition distance, dR .
dR
dR
Figure 2.19
Examples of profiled plates with a repeating pattern that can be built up from a beam consisting of rectangular elements.
For other shapes, the centroidal moments of inertia can be calculated in a similar way, or
found in standard textbooks. To create a plate profile, some of its constituent elements need
to be rotated. For elements that are inclined at an angle, θ, the moment of inertia needs to
be transformed to axes (y ′ , z ′ ) that are inclined at an anti-clockwise angle, θ, from the (y, z)
axes (see Fig. 2.20). This is done by replacing y with y ′ , where y ′ = y cos θ + z sin θ, and
replacing z with z ′ , where z ′ = z cos θ − y sin θ. Hence the transformed moment of inertia about
the y ′ -axis is
Iy ′ = Iy cos2 θ + Iz sin2 θ − Iyz sin 2θ =
Iy − Iz
Iy + Iz
+
cos 2θ − Iyz sin 2θ
2
2
(2.100)
(2.101)
where Iyz is the product of inertia given by
Iyz =
yz dS
S
Note that when either the y- or z-axis is an axis of symmetry of the element, then Iyz = 0.
137
S o u n d
I n s u l a t i o n
Z
z
z⬘
y⬘
θ
h
C
y
dS
Y
b
Figure 2.20
Axes used for calculating the moment of inertia of an element; this example shows a rectangular element. (Y, Z) are the
global rectangular axes. (y, z) are rectangular axes with their origin at the centroid (C) of the element. (y′ , z′ ) correspond to
rectangular axes (y, z) rotated anti-clockwise by an angle, θ , whilst keeping the origin at the centroid.
Returning to the generalized cross-section for the beam (Fig. 2.18), the neutral axis of the
beam cross-section is at a height, z0 , which is calculated using
N
Sn z n
z0 = n=1
N
n=1 Sn
(2.102)
where Sn is the area of each element (hbn ), and zn is the z-coordinate of the centroid for each
element.
The moment of inertia, Iy , for each of the N individual elements now needs to be transferred
from the y-axis to the neutral axis for the beam cross-section; this gives Iy0 . As these two axes
are parallel, this is done using the parallel axis theorem,
Iy0 = Iy + Sdn2
(2.103)
where dn is the distance of the centroid of each element from the neutral axis and is given by
dn = |z0 − zn |
(2.104)
For N rectangular elements, the moment of inertia for the beam cross-section can now be
calculated using Eq. 2.103 to give
Ib =
N
n=1
Iy + Sdn2 = h
N
n=1
dn2 +
h2 + bn2
h2 − bn2
+
cos 2θn
24
24
(2.105)
As the profiled plate is formed from beams running in the x-direction, the moment of inertia, Ip,x ,
for the profiled plate is Ip,x = Ib /dR . We can now calculate the bending stiffness for the profiled
138
Chapter 2
dc
y
h
dR
Figure 2.21
Cross-section of corrugated plate.
plate under the assumption that λB ≫ dR . So the plate bending stiffness in the x-direction (along
the profiles) is
Bp,x =
Ib
E
(1 − ν2 ) dR
(2.106)
and the plate bending stiffness in the y-direction (perpendicular to the profiles) is the same as
an isotropic plate but takes account of the effective increase in plate width due to the profiles,
dR
Eh3
12(1 − ν2 ) N
n=1 bn
Bp,y =
2.3.3.2.2
Corrugated plates
(2.107)
For corrugated plates of thickness, h, (see Fig. 2.21) with the profiles running in the x-direction,
the bending stiffness in the x- and y-directions is given by (Cremer et al., 1973)
⎤
⎡
Bp,x =
Bp,y =
2.3.3.2.3
Ehdc2 ⎢
⎣1 −
2
Eh3
12(1 − ν2 )
0.81
1 + 2.5
dc
2dR
1
1+
πdc
2dR
⎥
2⎦
for λB ≫ dR
(2.108)
2
Ribbed plates
For a plate of thickness, h, formed from a single material with rectangular ribs running in one
direction (see Fig. 2.22) the bending stiffness in the x and y directions is given by (Cremer
et al., 1973)
dy
dR 2
Bp,x = E
[C1 − (C1 − h)2 ] +
[(C1 − h)2 + (dz − C1 )2 ]
3
3
Bp,y
dR
Eh3
=
12 d − d 1 −
R
y
where C1 =
for λB ≫ dR
(2.109)
h3
dz3
1 dR dz2 + (dR − dy )h2
2 dR dz + (dR − dy )h
In Section 2.8.3.3, calculation of the driving-point mobility on ribbed plates requires the moment
of inertia about the y-axis for the repeating T-shape beam; this is given by
Iy =
%
1$
dy zC3 + dR (dz − zC )3 − (dR − dy )(dz − zC − h)3
3
(2.110)
139
S o u n d
I n s u l a t i o n
h
C
dz
dR
dy
y
zC
dR
Repeating T-shape
beam used for moment of
inertia calculation
Figure 2.22
Cross-section of ribbed plate with ribs running in the x -direction.
where
zC = dz −
dy dz2 + h2 (dR − dy )
2(hd R + dy dz − hd y )
2.4 Diffuse field
A diffuse field for vibration can be defined in a similar way to that for sound. However, it
is necessary to account for the fact that with some structure-borne sound waves, such as
bending waves, the intensity is not constant throughout the cross-section of the structure. For
this reason it is useful to refer to imaginary lines on the surface that run through the thickness
dimension of the structure. A diffuse field therefore has uniform intensity per unit angle for all
possible angles when waves are incident upon any such line within the structure. Waves that
are incident upon this line must have random phase with equal probability of a wave arriving
from any of the possible directions.
On a beam, a wave can only arrive from two possible directions, and as the boundary conditions
do not change with time, the phase is not random. For this reason there is no need to define
a diffuse field for beams; the modal overlap determines whether the wave field on the beam
can be considered as reverberant. On a sufficiently large plate, waves can potentially arrive
with uniform intensity from many directions in the plane within which it lies. Hence there is
the potential for a diffuse field when there are diffuse reflections from the plate boundaries.
Most walls and floors in buildings are rectangular with plate boundaries that are assumed to be
uniform, so the likelihood of diffuse reflections from these boundaries could be considered as
rather low. In practice the impedance along these boundaries will vary due to different material
properties and workmanship, and in walls there will be additional boundaries formed by the
perimeter of windows and doors. It can therefore be assumed that there will be a degree of
non-specular reflection. However, the assumption of diffuse vibration fields on walls and floors
is often harder to justify than with diffuse sound fields in rooms.
2.4.1 Mean free path
For beams or plates, the mean free path, dmfp , is the average distance travelled by a structureborne sound wave between two successive reflections from the boundaries. It is important
because it is used to determine the power that is incident upon the junction of connected
beams and plates.
140
Chapter 2
For a beam, the mean free path is equal to its length,
dmfp = Lx
(2.111)
For a plate, the mean free path is the same as for a two-dimensional sound field in a cavity.
Hence assuming that there are diffuse reflections from the plate boundaries
dmfp =
πS
U
(2.112)
where U is the perimeter of the plate (U = 2Lx + 2Ly for a rectangular plate).
2.5 Local modes
In a similar manner to sound waves in enclosed spaces, local modes occur when a structureborne sound wave travels around a beam or plate and, after reflection from various boundaries,
returns to the starting point travelling in exactly the same direction as when it first left, whilst
achieving phase closure. When calculating local modes for sound waves in rooms and cavities
it is usually sufficient to consider only one idealized boundary condition, namely rigid surfaces.
For structure-borne sound waves there are a variety of idealized boundary conditions: free
(i.e. no constraint), simply supported (also called pinned), clamped (also called fixed, or a rigid
constraint), and guided boundaries. The magnitude of the reflected wave from these idealized
boundaries is the same as the incident wave (i.e. a perfect reflection) but the boundary can
introduce phase shifts in the reflected wave. The actual boundary condition of a beam or plate
in a building is often unknown. This makes it difficult to gain accurate estimates for their mode
frequencies; but a rough estimate is often sufficient. Mode frequencies may only be needed
to indicate the frequency range in which it is reasonable to adopt a statistical approach to
the prediction of sound transmission. Fortunately at frequencies above the tenth local mode
(approximately), estimates of the statistical modal density are not significantly affected by
different boundary conditions.
We will only calculate the mode frequencies for solid homogeneous, isotropic beams and plates
with a limited set of relevant boundary conditions. Formulae to calculate the mode frequencies
for other combinations of boundary conditions can be found from Blevins (1979) and Leissa
(1973). For beams and plates with complex shapes, such as lightweight steel components
with cut-outs, the mode frequencies can be predicted using Finite Element Methods (FEM)
(Zienkiewicz, 1977) or measured using modal analysis. For laminated plates, details on mode
shapes and mode frequencies can be found from Qatu (2004).
2.5.1 Beams
To determine the conditions for phase closure that define the local modes on a beam it is
convenient to follow the journey of a propagating wave as it is reflected from each end of
the beam. For bending waves this also provides some insight into the vibration field near the
boundaries which is more easily obtained for wave propagation in one-dimension (beams) than
in two-dimensions (plates). The same approach will be useful when we look at vibration fields
near plate boundaries.
141
S o u n d
I n s u l a t i o n
Simply supported ends
Clamped ends
(a) Bending wave leaves an arbitrary starting point (•) on the beam and travels towards x ⫽ 0
η̂⫹
x ⫽⫺Lx
η̂⫹
x ⫽⫺Lx
x=0
(b) Bending wave is reflected at x ⫽ 0,
generating a nearfield
(b) Bending wave is reflected at x ⫽ 0
η̂⫺
(c) Bending wave is reflected at x ⫽ ⫺Lx
Phase change
of ⫺π on
reflection
η̂⫹
x=0
η̂⫺
Phase change
of ⫺π on
reflection
η̂n⫹ Exponentially
decaying
nearfield
Phase change
of ⫺π /2 on
reflection
(c) Bending wave is reflected at x ⫽ ⫺Lx
generating a nearfield
Exponentially
η̂n⫺
decaying
η̂⫹
nearfield
Phase change
of ⫺π/2 on
reflection
Figure 2.23
Bending wave modes on beams. A mode occurs when the bending wave returns to the starting point travelling in exactly the
same direction as when it first left, whilst achieving phase closure.
2.5.1.1 Bending waves
The effect of the boundaries can be seen by positioning the beam so that one end is at x = −Lx ,
and the other is at x = 0. We will assume that lateral displacement is in the z-direction. At some
arbitrary starting point near the middle of the beam we consider a bending wave, for which the
lateral displacement is described by η̂+ exp (−ikB x), propagating in the positive x-direction
towards the boundary at x = 0 (see Fig. 2.23). We start by assuming that the ends of the beam
are simply supported; for this boundary condition the lateral velocity, ∂η/∂t, and the bending
moment, My (Eq. 2.47) at these ends must be zero. Hence the boundary conditions are
∂η
= 0 and
∂t
∂2 η
=0
∂x 2
at x = 0 and x = −Lx
(2.113)
Using the boundary conditions at x = 0 in the general solution for bending wave motion on a
beam (Eq. 2.58) gives
142
η̂+ + η̂− + η̂n+ = 0
(2.114)
−η̂+ − η̂− + η̂n+ = 0
(2.115)
Chapter 2
where η̂− corresponds to the reflected bending wave travelling in the negative x-direction, and
η̂n+ corresponds to the nearfield that decays in the negative x-direction.
Combining Eqs 2.114 and 2.115 yields two pieces of information about what happens at the
boundary. For the propagating bending waves, η̂− = −η̂+ , which means that there is a phase
shift of −π when the bending wave is reflected at the end. As η̂n+ = 0, there is no nearfield
accompanying the reflected wave that travels in the negative x-direction. The fact that no
nearfield is generated at a simply supported boundary is very convenient as it simplifies the
analysis, although real boundary conditions for beams and plates often generate nearfields. We
now know that when this reflected wave reaches x = −Lx , it will undergo another phase shift of
−π and that there will not be a nearfield accompanying the wave that returns to the starting point.
Hence modes occur at frequencies where phase closure is achieved for the propagating bending wave after it has travelled a distance of 2Lx , during which there will have been two phase
shifts of −π. Recalling the analysis of room modes, we see that we have a one-dimensional
version of the three-dimensional situation that was used for rooms with perfectly reflecting
boundaries. As the phase shifts due to reflection amount to −2π, phase closure occurs when
kx (2Lx ) = p(2π)
(2.116)
where p takes positive integer values 1, 2, 3, etc., and the relationship to the wavenumber is
kx = kB =
pπ
2π
=
λB
Lx
(2.117)
Substituting kB from Eq. 2.57 into Eq. 2.117 gives the bending mode frequency, fp(B) , for thin
beams with simply supported ends as
π Bb p 2
fp(B) =
(2.118)
2 ρl L x
For beams connected to other building elements at both ends, such as structural (load-bearing)
columns, it can be useful to calculate the mode frequency for both simply supported and
clamped ends; in practice, the modal response often occurs somewhere between these two
calculated frequencies. At clamped ends the lateral velocity, ∂η/∂t, and the angular velocity,
ωy (Eq. 2.41) must be zero. Hence the boundary conditions are
∂η
= 0 and
∂t
∂2 η
=0
∂t∂x
at x = 0 and x = −Lx
(2.119)
As before we take a starting point somewhere in the middle of the beam and consider a bending wave propagating in the positive x-direction towards the boundary at x = 0 (see Fig. 2.23).
Substitution of the boundary conditions into the general solution at x = 0 gives
η̂+ + η̂− + η̂n+ = 0
(2.120)
−i η̂+ + i η̂− + η̂n+ = 0
(2.121)
Equations 2.120 and 2.121 can be combined to provide information on the reflected bending
wave and the nearfield that return from this clamped end. For the bending waves, η̂− = −i η̂+ ,
which means that there has been a phase shift of −π/2 on reflection. At x = 0 the magnitude of
the nearfield relative to the incident bending wave is given by η̂n+ = (i − 1)η̂+ . So the reflected
bending wave that travels in the negative x-direction is accompanied by a nearfield which
is exponentially decaying from x = 0 towards x = −Lx . The existence of the nearfield means
143
S o u n d
I n s u l a t i o n
that defining the modes through phase closure will be more awkward than with simply supported boundaries. We now have to follow a reflected bending wave and a nearfield towards
the boundary at x = −Lx from which we will again have bending waves and accompanying
nearfields to follow on their way back to the starting point. However, due to the exponential decay of the nearfield it is reasonable to assume that when the beam is sufficiently long,
the nearfield will be negligible by the time it reaches x = −Lx . A ‘sufficiently long’ beam is a
frequency-dependent description because the distance over which the nearfield decays to a
negligible level will decrease with increasing frequency. However, this assumption allows us to
ignore the nearfield and find the mode frequencies where phase closure is achieved after the
bending wave has travelled a distance of 2Lx , during which it will have undergone two phase
shifts of −π/2. As the phase shifts due to reflection amount to −π, phase closure occurs when
kx (2Lx ) = (2p − 1)π
(2.122)
where p can only take positive integer values 2, 3, 4, etc.
The bending mode frequency, fp(B) , for thin beams with clamped ends can be estimated from
fp(B)
π
≈
8
Bb
ρl
2p − 1
Lx
2
(2.123)
Hence a beam with clamped ends will have higher mode frequencies than the same beam with
simply supported ends.
Another idealized boundary condition is the free boundary where the bending moment, My (Eq.
2.47) and the shear force, Fz (Eq. 2.54) must be zero. However, this condition is less relevant
to beams in buildings because most beams which have a free end will be supported at various
points along their length; hence it will be the combination of these supports along with the free
ends that defines the modes. It is therefore sufficient to be aware that reflection of a bending
wave from a free boundary also gives rise to a nearfield. In passing, it is worth noting that a
thin beam with free ends has the same mode frequencies as a thin beam with clamped ends
and can be estimated using Eq. 2.123.
2.5.1.2 Torsional waves
For torsional waves propagating along a beam in both directions, the angular torsional
displacement is described by
θ(x, t) = [θ̂+ exp (−ikx x) + θ̂− exp (ikx x)] exp (iωt)
(2.124)
For torsional waves the idealized boundary conditions that are most common are free or
clamped ends. At free ends the moment, Mx (Eq. 2.224), must be zero. At clamped ends
the angle of rotation, θ, must be zero. Clamped ends are more relevant to beams in buildings,
for which the boundary condition at x = 0 yields the relation, θ̂− = −θ̂+ . This means there will be
a phase shift of −π upon reflection from each end. Hence we have exactly the same situation
as with bending waves on a beam with simply supported ends, and phase closure is defined
by Eq. 2.116, for which the torsional wavenumber is
kx = kT,b =
144
pπ
2πf
=
cT,b
Lx
(2.125)
Chapter 2
Therefore the torsional mode frequency, fp(T) , for beams with both ends clamped is
fp(T) =
pcT,b
2Lx
(2.126)
where p takes positive integer values 1, 2, 3, etc.
2.5.1.3 Quasi-longitudinal waves
For quasi-longitudinal waves propagating along a beam in both directions, the displacement is
described by
ξ(x, t) = [ξ̂+ exp(−ikx x) + ξ̂− exp(ikx x)] exp(iωt)
(2.127)
As with torsional waves there are two idealized boundary conditions that are commonly considered, free or clamped ends. For free ends the longitudinal force, Fx (Eq. 2.9) must be zero.
For clamped ends the in-plane displacement, ξ, must be zero. Both conditions give rise to the
same mode frequencies, but with different mode shapes. We will look at the clamped condition,
which is more relevant to structural beams and columns in buildings. Substitution of the boundary conditions into the solution at x = 0 therefore gives ξ̂− = −ξ̂+ , which means a phase shift
of −π upon reflection from each end. Phase closure is therefore defined by Eq. 2.116. Hence
the quasi-longitudinal mode frequency, fp(L) , for beams with both ends clamped (or free) is
fp(L) =
pcL,b
2Lx
(2.128)
where p takes positive integer values 1, 2, 3, etc.
2.5.1.4 Modal density
The modal density of a beam is calculated by arranging the eigenvalues along a line. This is a
one-dimensional version of the mode lattice used for rooms. The length associated with each
eigenvalue is a line of length π/Lx . An example mode line is shown in Fig. 2.24 for bending
modes on a beam with simply supported ends.
Below a specified wavenumber, k, the number of modes is equal to the number of eigenvalues that lie upon a line of length, k. Therefore the number of modes, N(k), below the
wavenumber, k, is
N(k) =
kLx
π
(2.129)
The modal density can then be calculated from N(k) and the group velocity using Eq. 1.57.
Strictly speaking, N(k) needs a correction factor for the various different boundary conditions
(Lyon and DeJong, 1995). However, Eq. 2.129 can be used to give a reasonable estimate
k
0
kx
π/Lx
Figure 2.24
Mode line for bending modes on a beam with simply supported ends. A line of length, k, encloses eigenvalues below
wavenumber, k.
145
S o u n d
I n s u l a t i o n
in most cases as there is always a degree of uncertainty in the actual boundary conditions.
Hence the modal densities can be calculated using
nB,b (f ) =
Lx
cB,b
nT,b (f ) =
nL,b (f ) =
2Lx
cL,b
for bending waves (thin beam)
(2.130)
2Lx
cT,b
for torsional waves
(2.131)
for quasi-longitudinal waves
(2.132)
The statistical mode count, Ns , in a frequency band is calculated from the modal density using
Eq. 1.63. It is instructive to look at differences between the statistical mode count, and the
mode count that is determined by counting the number of modes that fall within each band. To
do this we will look at the modes for a common beam found in buildings, a timber stud from a
timber frame wall.
This idealized timber stud will be assumed to have simply supported ends for the bending
modes, and clamped ends for the torsional and quasi-longitudinal modes. It is also assumed
that any fixings along the length of the stud that would occur in a real wall have a negligible
effect on the mode frequencies. The mode count and the statistical mode count in one-thirdoctave-bands for this beam are shown in Fig. 2.25. The statistical mode count for each wave
type is only plotted for frequency bands at and above the band that contains the estimated
fundamental mode. The frequencies for the thin beam limit are both close to the 5000 Hz onethird-octave-band so it is reasonable to calculate the bending modes using thin beam theory
(Eq. 2.118) across the entire building acoustics frequency range.
Figure 2.25 serves as a reminder that there can be bending wave modes with motion in both
the y- and z-direction, although the type of excitation and the direction in which it is applied
will determine which, if any, are excited. For studs in walls, or joists in floors, airborne or
impact sources usually excite the bending wave modes that have motion perpendicular to the
surface of the wall or floor. Both of the fundamental bending modes, f1(B) , are below the 50 Hz
band. However, it is clear from the mode count in Fig. 2.25a that this hardly results in an
abundance of bending modes in and above the 50 Hz band. For a beam with simply supported
ends the bending mode frequencies, fp(B) , (see Eq. 2.118) above the fundamental mode (i.e.
where p > 1) are multiples of f1(B) with the factor p2 . Hence there is a relatively wide frequency
interval between adjacent modes which results in low mode counts. In each frequency band
there are zero, one, or two modes for each of the four different types: bending (y-direction),
bending (z-direction), torsional, and quasi-longitudinal. These low mode counts are in stark
contrast to the large numbers of modes in rooms. The fundamental bending modes occur at
much lower frequencies than the fundamental torsional and quasi-longitudinal modes. For this
timber stud there are no torsional and quasi-longitudinal modes in the low-frequency range, so
only bending modes need to be considered in the modal response of the beam. In the mid- and
high-frequency ranges there can be torsional and quasi-longitudinal modes, which, if excited,
also need consideration.
We can now use Fig. 2.25 to compare the different mode counts. In the low- and mid-frequency
ranges the statistical mode count now takes fractional values <1. To find some physical meaning in a fractional mode count we should first recall that the mode count in Fig. 2.25a was based
on mode frequencies for idealized boundary conditions; we rarely know the actual boundary conditions. We then assumed that we had precise knowledge of the dimensions and the
146
Chapter 2
(a) Mode count, N
10
Mode count, N
Timber frame wall stud
Lx ⫽ 2.3 m
hy ⫽ 0.038 m, hz ⫽ 0.063 m
fB(thin) ⫽ 6579 Hz (y) and 3968 Hz (z)
f1(B) ⫽ 16 Hz (y) and 27 Hz (z)
Timber (soft wood)
Bending modes (thin beam) y-direction
Bending modes (thin beam) z-direction
Torsional modes
Quasi-longitudinal modes
1
0.1
50
80
125
200
315
500
800
1250
2000
One-third-octave-band centre frequency (Hz)
3150
5000
(b) Statistical mode count, Ns
10
Statistical mode count, Ns
Timber frame wall stud
Lx ⫽ 2.3 m
hy ⫽ 0.038 m, hz ⫽ 0.063 m
fB(thin) ⫽ 6579 Hz (y) and 3968 Hz (z)
f1(B) ⫽ 16 Hz (y) and 27 Hz (z)
Timber (soft wood)
Bending modes (thin beam) y-direction
Bending modes (thin beam) z-direction
Torsional modes
Quasi-longitudinal modes
1
0.1
50
80
125
200
315
500
800
1250
2000
One-third-octave-band centre frequency (Hz)
3150
5000
Figure 2.25
Beams: mode counts for a timber frame wall stud.
material properties of a single timber stud. In practice there will be uncertainty in all of these
parameters, whether they are estimated or measured. In fact, as a lightweight wall is built from
several timber studs we expect there to be natural variation in the properties of different studs;
this also gives rise to uncertainty in the mode frequencies. The mode count also ignores the
fact that each mode has a finite bandwidth due to damping; hence the 3 dB bandwidth can overlap two or more bands. For these reasons the mode count is only useful in indicating potential
issues with a lack of modes; we remain uncertain as to which band we can attribute each mode
in the actual construction. Therefore the statistical modal density tends to be of more practical
use. To try and ensure that the statistical modal density is only used in frequency bands above
the fundamental mode, this mode can be estimated by assuming idealized boundary conditions.
147
S o u n d
I n s u l a t i o n
(a) Timber floor joist
10
Statistical mode count, Ns
Timber floor joist
Lx ⫽ 4 m
hy ⫽ 0.05 m, hz ⫽ 0.225 m
fB(thin) ⫽ 5000 Hz (y) and 1111 Hz (z)
f1(B) ⫽ 7 Hz (y) and 32 Hz (z)
Timber (soft wood)
Bending modes (thin beam) y-direction
Bending modes (thin beam) z-direction
Torsional modes
Quasi-longitudinal modes
1
0.1
50
80
125
200
315
500
800
1250
2000
One-third-octave-band centre frequency (Hz)
3150
5000
(b) Structural concrete column
10
Statistical mode count, Ns
Structural column
Lx ⫽ 3 m
hy ⫽ 0.4 m, hz ⫽ 0.6 m
fB(thin) ⫽ 475 Hz (y) and 317 Hz (z)
f1(B) ⫽ 77 Hz (y) and 115 Hz (z)
In situ concrete
Bending modes (thin beam) y-direction
Bending modes (thin beam) z-direction
Torsional modes
Quasi-longitudinal modes
1
0.1
50
80
125
200
315
500
800
1250
2000
One-third-octave-band centre frequency (Hz)
3150
5000
Figure 2.26
Beams: statistical mode counts for a timber floor joist and a structural concrete column.
The statistical mode count for the timber stud can be compared with two other common
beams found in buildings, a timber floor joist and a structural column as shown in Fig. 2.26.
These beams are much thicker than the timber stud and the thin beam limit occurs within the
building acoustics frequency range. Note that the structural column has a very low statistical
mode count.
2.5.2 Plates
To simplify the calculation of specific mode frequencies and to derive the modal density in rooms
and cavities we assumed box-shaped spaces. In a similar way we will now use rectangular
148
Chapter 2
plates to calculate the specific mode frequencies. This is convenient because most plates in
buildings are rectangular, but it will also allow us to determine statistical modal densities for
any shape of plate.
For beams we looked at the specific phase change associated with different boundary conditions and the generation of nearfields that can occur when bending waves impinge upon certain
boundary conditions. There are buildings where each side of a plate that forms a wall or floor has
significantly different boundary conditions; however there will always be uncertainty in describing each of these conditions. From our previous analysis of modes on beams, a practical
solution for plates in buildings is to assume that all boundaries are simply supported for bending waves so that there is no need to consider nearfields, and all boundaries are clamped for
in-plane waves. This assumption is usually reasonable when the boundaries of a plate or beam
are rigidly connected to other beams or plates. One exception is a plate that forms a floating
floor (e.g. a floating floor screed on a resilient material) for which the plate boundaries can be
assumed to be free; in this case the fundamental bending mode will be lower than for simply
supported boundaries and can be calculated from tabulated data if required (see Blevins, 1979;
Leissa, 1973).
2.5.2.1 Bending waves
On a thin plate with simply supported boundaries there will be a plane bending wave field without
nearfields. With this plane wave field we have already used the wave equation to find the relationship between the wavenumber and the constants, kx and ky in Eq. 2.90. To determine these
constants we can compare a rectangular plate with simply supported boundaries to two beams
with simply supported ends. We will orientate these beams with lengths, Lx and Ly , along the xand y-axis respectively. For the beam aligned along the x-axis, we already have the equation
that describes kx (Eq. 2.116) for phase closure. For the beam aligned along the y-axis it therefore follows that phase closure is achieved when ky (2Ly ) = q(2π) where q takes positive integer
values 1, 2, 3, etc. Hence from Eq. 2.90, phase closure on the plate must be satisfied when
2 2
qπ
pπ
+
(2.133)
kB2 =
Lx
Ly
The bending mode frequencies for a simply supported, isotropic, rectangular plate can now be
calculated from Eqs 2.133, 2.65, and 2.66, to give
&
2 '
2
cB,p
p 2
q
q
πhcL,p
p 2
fp,q(B) =
= √
+
+
(2.134)
2
Lx
Ly
Lx
Ly
4 3
where p and q take positive integer values 1, 2, 3, etc.
Similarly, for a simply supported, orthotropic, rectangular plate,
&
&
2
2
2 '
2 '
π
p
q
p
q
πh
cL,p,x
fp,q(B) = √
+ Bp,y
= √
+ cL,p,y
Bp,x
2 ρs
Lx
Ly
Lx
Ly
2 12
(2.135)
2.5.2.2 Transverse shear waves
For transverse shear waves on a rectangular plate with clamped boundaries, we have
the boundary condition that the in-plane displacement, ζ, must be zero at the boundaries.
149
S o u n d
I n s u l a t i o n
Hence we use the wavenumber relationship kT2 = kx2 + ky2 where kx (2Lx ) = p(2π) and
ky (2Ly ) = q(2π). This gives the transverse shear mode frequencies as
2
2
cT,p
q
q
cL,p 1 − ν
p 2
p 2
fp,q(T) =
+
=
+
(2.136)
2
Lx
Ly
2
2
Lx
Ly
where p and q take positive integer values 1, 2, 3, etc.
2.5.2.3 Quasi-longitudinal waves
For quasi-longitudinal waves on a beam with clamped ends, we have already seen that the
phase shift upon reflection is the same as for bending waves on a beam with simply supported ends. Hence we use the same expressions for kx and ky as with bending waves. For
quasi-longitudinal waves on a rectangular plate with clamped boundaries, the wavenumber
relationship kL2 = kx2 + ky2 gives the quasi-longitudinal mode frequencies as
fp,q(L)
cL,p
=
2
p
Lx
2
+
q
Ly
2
(2.137)
where p and q take positive integer values 1, 2, 3, etc.
2.5.2.4 Modal density
The modal density is calculated in a similar way to cavities where the eigenvalues are arranged
in a two-dimensional lattice as shown in Fig. 2.27. It is assumed that the plate boundaries
are simply supported (bending waves) or clamped (in-plane waves) such that p and q only
take positive integer values; hence there are no eigenvalues on the x- and y-axes. The area
associated with each eigenvalue is a rectangle with an area of π2 /Lx Ly , which is equal to π2 /S.
The number of modes below a specified wavenumber, k, is equal to the number of eigenvalues
that are contained within one-quarter of the area of a circle with radius, k (see Fig. 2.27). Hence
the number of modes, N(k), below the wavenumber, k, is
N(k) =
k 2S
4π
(2.138)
where k is kB , kT or kL for bending, transverse shear, or quasi-longitudinal waves, respectively.
For each of the three wave types the modal density is calculated using Eq. 1.57 where the
group velocity, cg is equal to 2cB for bending waves, cL for quasi-longitudinal waves, and cT
for transverse shear waves. Hence the modal densities are
√
S 3
for bending waves (thin plates)
(2.139)
nB,p (f ) =
hcL,p
150
nT,p (f ) =
2πfS
4πfS
= 2
2
cT,p
cL,p (1 − ν)
nL,p (f ) =
2πfS
2
cL,p
for transverse shear waves
(2.140)
for quasi-longitudinal waves (thin plates)
(2.141)
Chapter 2
ky
k
k2,2
π/Ly
0,0
k
π/Lx
kx
Figure 2.27
Mode lattice for a two-dimensional plate with simply supported boundaries. The vector corresponding to eigenvalue, k2,2 , is
shown as an example. The quadrant with radius, k, encloses eigenvalues below wavenumber, k.
Note that for bending waves on orthotropic plates, the effective bending stiffness can be used
to calculate an effective quasi-longitudinal phase velocity for use in Eq. 2.139.
The statistical mode count, Ns , is calculated from the modal density using Eq. 1.63 and is
shown for four different plates in Fig. 2.28: a sheet of plasterboard, a concrete floor, and
two different masonry walls. It is assumed that the plate boundaries are simply supported for
bending waves, and clamped for transverse shear and quasi-longitudinal modes. For each
wave type the statistical mode count is only plotted for frequency bands at and above the band
that contains the estimated fundamental mode.
As with beams, the fundamental bending modes on plates are at a lower frequency than the
fundamental in-plane modes. For these plates this means that the in-plane modes only need
to be considered in the mid- and high-frequency ranges.
In the low-frequency range the concrete floor and masonry walls have <5 bending modes
in each band. In contrast, a sheet of plasterboard has many more bending modes than
masonry/concrete walls and floors. However, sheets of plasterboard that form part of a wall or
floor can be fixed in many different ways. Some fixings will significantly change the bending
wave modal density from this idealized situation of simply supported boundaries. For example,
the change due to a few screw fixings onto a light steel frame will be less significant than large
dabs of adhesive that are closely spaced over the entire surface of the sheet.
151
S o u n d
I n s u l a t i o n
(a) Sheet of plasterboard
Statistical mode count, Ns
1000
100
Sheet of plasterboard
Lx ⫽ 2.4 m, Ly ⫽ 1.2 m
h ⫽ 0.0125 m
ρs ⫽ 10.8 kg/m2
fB(thin) ⫽ 5891 Hz
f11(B) ⫽ 7 Hz
Natural gypsum
Bending modes (thin plate)
Transverse shear modes
Quasi-longitudinal modes
10
1
0.1
50
80
125
200
315
500
800
1250 2000
One-third-octave-band centre frequency (Hz)
3150
5000
3150
5000
(b) Concrete floor
Statistical mode count, Ns
1000
100
Concrete floor
Lx ⫽ 4 m, Ly ⫽ 3.5 m
h ⫽ 0.14 m
ρs ⫽ 308 kg/m2
fB(thin) ⫽ 1485 Hz
f11(B) ⫽ 35 Hz
In situ concrete
Bending modes (thin plate)
Transverse shear modes
Quasi-longitudinal modes
10
1
0.1
50
80
125
200
315
500
800
1250 2000
One-third-octave-band centre frequency (Hz)
Figure 2.28
Plates: statistical mode counts for a sheet of plasterboard, a concrete floor, and two different masonry walls. Note that values
are only shown at and above the frequency band that contains the estimated fundamental mode for each wave type.
2.5.3 Equivalent angles
For modes on beams, the waves propagate along the axis of the beam so there is no need
to consider equivalent angles. For modes on plates, the two-dimensional lattice in k-space
defines the equivalent angles for the different wave types. They are defined in the same way
as for rooms and cavities (Section 1.2.5.4), therefore the equivalent angles are related to the
mode wavenumber, kp,q , and the constants, kx and ky by
kx = kp,q sin θx
152
ky = kp,q sin θy
(2.142)
Chapter 2
(c) Masonry wall
Statistical mode count, Ns
1000
100
10
Masonry wall
Lx ⫽ 4 m, Ly ⫽ 2.5 m
h ⫽ 0.215 m
ρs ⫽ 430 kg/m2
fB(thin) ⫽ 814 Hz
f11(B) ⫽ 69 Hz
Dense aggregate
blocks (solid)
Bending modes (thin plate)
Transverse shear modes
Quasi-longitudinal modes
1
0.1
50
80
125
200
315
500
800
1250 2000
One-third-octave-band centre frequency (Hz)
3150
5000
3150
5000
(d) Masonry wall
Statistical mode count, Ns
1000
100
Masonry wall
Lx ⫽ 4 m, Ly ⫽ 2.5 m
h ⫽ 0.1 m
ρs ⫽ 80 kg/m2
fB(thin) ⫽ 1040 Hz
f11(B) ⫽ 19 Hz
Aircrete blocks (solid)
Bending modes (thin plate)
Transverse shear modes
Quasi-longitudinal modes
10
1
0.1
50
80
125
200
315
500
800
1250 2000
One-third-octave-band centre frequency (Hz)
Figure 2.28
(Continued)
and the equivalent angles for each plate mode are
θx = asin
pc
2Lx fp,q
θy = asin
qc
2Ly fp,q
(2.143)
where c is cB,p , cL,p , or cT,p for bending, quasi-longitudinal, or transverse shear waves
respectively.
The equivalent angles θx and θy are defined from lines normal to the x- and y-axis respectively;
hence, θx and θy also represent the equivalent angles of incidence upon the plate boundaries,
Lx and Ly respectively. The ratio, c/fp,q , in Eq. 2.143 is the same for all three wave types, hence
for modes with the same values of p and q, the angles θx and θy are identical.
153
S o u n d
90
I n s u l a t i o n
Sheet of plasterboard
Lx ⫽ 2.4 m, Ly = 1.2 m
80
h ⫽ 0.0125 m
ρs ⫽ 10.8 kg/m2
70
Natural gypsum
Angle (°)
60
50
θx
40
θy
30
20
10
Bending modes
0
1
10
100
Frequency (Hz)
1000
10 000
Figure 2.29
Equivalent angles for the bending modes of a sheet of plasterboard.
Examples of equivalent angles for a sheet of plasterboard and a masonry wall are shown in
Figs. 2.29 and 2.30 respectively. For the purpose of these examples we can overlook the
thin plate limits and the fact that masonry walls are not homogeneous plates with specularly
reflecting boundaries at high frequencies. The plate boundaries are assumed to be simply
supported for bending waves, and clamped for transverse shear and quasi-longitudinal modes;
hence it is not possible for the equivalent angles to be 0◦ or 90◦ . For the bending modes of the
plasterboard, there are a large number of modes and there is symmetry in the pattern about
the 45◦ line because Lx = 2Ly . The large number of bending modes for the plasterboard is in
marked contrast to the masonry wall. When the angles are considered as equivalent angles of
incidence upon each boundary, it is clear that there is a limited range of angles incident upon
the plate boundaries in the low- and mid-frequency ranges. This is relevant to the calculation of
structure-borne sound transmission between coupled plates using the wave approach where
it is assumed that waves are incident at all possible angles.
2.6 Damping
Damping is described here for vibration in a similar style to sound in Chapter 1 with emphasis
on the internal, coupling, and total loss factors that are defined for use with Statistical Energy
Analysis (SEA).
2.6.1 Structural reverberation time
The structural reverberation time, Ts , is defined in the same way as for reverberation times
in rooms. In practice we only need to measure and use the structural reverberation time for
154
Chapter 2
90
Angle (°)
Masonry wall
80 Lx ⫽ 4 m, Ly ⫽ 2.5 m
h ⫽ 0.215 m
ρs ⫽ 430 kg/m2
70
Dense aggregate
blocks (solid)
60
50
θx
40
θy
30
20
10
Bending modes
0
10
100
1000
10 000
1000
10 000
1000
Frequency (Hz)
10 000
Frequency (Hz)
90
80
70
Angle (°)
60
50
θx
40
θy
30
20
10
Transverse shear modes
0
10
100
Frequency (Hz)
90
80
70
Angle (°)
60
50
θx
40
θy
30
20
10
0
10
Quasi-longitudinal modes
100
Figure 2.30
Equivalent angles for bending, transverse shear, and quasi-longitudinal modes of a 215 mm dense aggregate wall.
155
S o u n d
I n s u l a t i o n
bending wave fields although we can separately calculate structural reverberation times for
bending, quasi-longitudinal, and transverse shear waves. In comparison with reverberation
times in rooms, structural reverberation times of building elements can be quite short. As with
rooms, non-diffuse fields on plates also cause curvature of the decay curve when the decaying
modes have very different decay times. For the same reasons as discussed for rooms (Section
1.2.7.5.2), T10 , T15 , or T20 are used to relate the decay time to reverberant vibration levels rather
than T30 or T60 .
2.6.2
Absorption length
For rooms, the absorption area is a convenient way to describe all the absorption in a room.
For bending waves on a plate it is possible to use a similar concept with an absorption length,
a, in metres. This can be used to describe all the absorption at the plate boundaries by using
a single length which is totally absorbing.
The absorption length is defined as the ratio of the power absorbed by the plate boundaries,
to the intensity incident upon them. The power absorbed is
Wabs = ωηE = ω
2.2
E
f Ts
(2.144)
where η is the total loss factor, and E is the plate energy.
Assuming a diffuse vibration field, the intensity that is incident upon the boundaries (in W/m) is
Iinc =
cg(B),p E
2cB,p E
=
dmfp U
dmfp U
(2.145)
where dmfp is given by Eq. 2.112.
The bending phase velocity can be written in terms of the critical frequency of the plate, fc ,
using Eqs 2.66 and 2.201 to give
f
(2.146)
cB,p = c0
fc
Hence, the ratio of Eq. 2.144 to Eq. 2.145 defines the absorption length as
2.2π2 S fc
c0 Ts
f
(2.147)
Equation 2.147 is only suitable for homogeneous plates where the critical frequency is known.
This is slightly awkward because a single critical frequency is not well-defined for nonhomogeneous and/or orthotropic plates. Therefore to generalize the definition for a plate, i,
an equivalent absorption length, ai , is defined by using a reference frequency rather than the
critical frequency (Gerretsen, 1996; also see EN 12354 Parts 1 & 2)
2.2π2 Si fref
π 2 Si η i
ai =
=
fref f
(2.148)
c0 Ts,i
f
c0
where fref is a reference frequency of 1000 Hz. Note that the equivalent absorption length only
corresponds to a totally absorbing length for a plate which actually has a critical frequency
of 1000 Hz, and that measured structural reverberation times or total loss factors will include
both radiation and internal losses.
156
Chapter 2
2.6.3
Internal loss factor
The internal losses describe the inherent material damping. When a beam or plate deforms
whilst undergoing wave motion, the internal losses convert vibration energy into heat; therefore high internal losses are desirable for sound insulation purposes. In comparison with other
parameters that describe the material properties, such as density or longitudinal phase velocity, the internal loss factor (denoted as ηint , or, ηii for SEA subsystem i), is more awkward
to quantify. Internal losses can vary depending upon the type of wave motion, frequency,
temperature, amplitude of vibration, and manufacturing process. This is rarely a problem
when predicting vibration transmission between adjacent rooms in buildings. Uncertainty in
the internal loss factor can often be tolerated when the sum of the coupling loss factors
is much greater than the internal loss factor; this occurs with many plates and beams in
buildings.
Measurements used to determine internal loss factors for beams and plates are discussed in
Section 3.11.3.4. Measured internal loss factors only tend to be available for bending waves;
example values for common building materials are listed in the Appendix, Table A2. Due to a
lack of data it is often assumed that the internal loss factors for in-plane waves are the same
as those for bending waves; measurements indicate that this is a reasonable assumption for
materials such as concrete or bricks (Kuhl and Kaiser, 1952).
For solid homogeneous plates formed from concrete or masonry it is reasonable to assume
a frequency-independent internal loss factor (bending or in-plane waves) over the building
acoustics frequency range. Frequency-dependence has been measured for solid concrete
covered with a granular material (e.g. sand) and hollow masonry/concrete units filled with sand;
with increasing frequency, the granular material tends to increase the internal losses (Kuhl and
Kaiser, 1952). The internal loss factor for laminated glass is also frequency-dependent and
tends to increase significantly with frequency (see Section 3.11.3.4). The effect of temperature
on the internal loss factor varies between materials; it is of minor importance for concrete, but
very important for laminated glass.
If the internal loss factor changes with the amplitude of vibration, the internal damping is
described as non-linear. This has been observed in the measurement of internal damping with
both bricks and hollow blocks that were mortared together; increasing the vibration amplitude
resulted in higher internal loss factors (Kuhl and Kaiser, 1952; Watters, 1959). It can be tentatively assumed that it applies to most mortared masonry blocks. However, this non-linearity
needs to be kept in perspective. The relatively high vibration amplitudes that can be used in
the laboratory to measure the internal loss factor do not usually correspond to the vibration
amplitudes that occur on building elements excited by typical airborne and structure-borne
sound sources. For structure-borne sound transmission in buildings it is reasonable to assume
that Hooke’s law is obeyed in the majority of situations (Cremer et al., 1973). Internal loss
factors should therefore be determined from measurements at low levels of excitation where
the internal damping can be considered as linear.
Another form of internal damping occurs due to losses at the edges of a plate caused by
bending wave motion. This is sometimes referred to as edge damping; although it should
not be confused with coupling losses to the surrounding structure along the plate edges. Edge
damping is usually specific to the edge fixing for a specific material. For panes of glass mounted
in rubber/neoprene gaskets, dissipation can occur due to frictional losses at the edges caused
by the gasket mounting (Utley and Fletcher, 1973). For metal beams riveted to metal plates,
157
S o u n d
I n s u l a t i o n
energy can be dissipated due to viscous losses of air pumping in and out of the small spaces
between the beam and the plate (Maidanik, 1966).
For two plates closely connected together over their surface, bending wave motion also causes
viscous losses in the thin layer of air between them which can result in a higher internal loss
factor that is also frequency-dependent (Trochidis, 1982).
2.6.4 Coupling loss factor
Coupling losses describe the energy losses from a plate or beam via structural connections to
other plates or beams and via sound radiation to the surrounding air (or other gas). For example,
the coupling losses for one leaf of a cavity masonry wall are due to: connections around the
perimeter to other masonry walls, connections such as wall ties, and sound radiation into a
room on one side and a cavity on the other. For masonry/concrete walls and floors, the radiation
losses are often insignificant compared to the structural coupling losses.
The role of the coupling loss factor, ηij , in quantifying vibration transmission is discussed in
Chapters 4 and 5. At this stage we simply focus on vibration transmission from one plate or
beam (referred to with subscript i) to another plate or beam (referred to with subscript j) to
which it is coupled. We will assume that these plates or beams have reverberant vibration
fields.
For any type of wave on plate or beam, i, that is incident upon the junction connecting i to j,
the transmission coefficient, τij , is
τij =
ωηij Ei
Wij
=
Winc,i
Winc,i
(2.149)
where Wij is the power transmitted from i to j, and Winc,i is the incident power on i.
The coupling loss factor can now be written as
ηij =
τij Winc,i
ωEi
(2.150)
In this section we are only concerned with quantifying Winc,i on beams and plates; the transmission coefficient for different types of junction will be discussed in Section 5.2. The incident
power upon a boundary is determined by using the mean free path to quantify the number
of times that vibrational energy is reflected from the boundaries of a beam or a plate every
second.
For a beam i that is connected at both of its ends to either a plate or a beam, the power incident
upon one of its two ends is
Winc,i =
cg,i Ei 1
dmfp,i 2
(2.151)
where dmfp is the mean free path given by Eq. 2.111 and from Eq. 2.150 the coupling loss
factor is
ηij =
158
cg,i τij
4πfLi
(2.152)
Chapter 2
where Li is the length of beam i. For a plate i with a junction length, Lij , along its perimeter, Ui ,
that connects plate i to plate j, the power incident upon the junction is
Winc,i =
cg,i Ei Lij
dmfp,i Ui
(2.153)
where dmfp is the mean free path given by Eq. 2.112 and from Eq. 2.150 the coupling loss
factor is
ηij =
cg,i Lij τij
2π2 fSi
(2.154)
2.6.5 Total loss factor
The total loss factor for subsystem i (denoted as ηi ) is the sum of the internal loss factor for
subsystem i and all the coupling loss factors from subsystem i to other subsystems (Eq. 1.106).
The coupling loss factors are due to sound radiation (Section 4.3.1.1) and/or structural coupling
to other plates/beams (Section 5.2).
Total loss factors can be calculated by predicting or measuring the internal and coupling loss
factors. If accurate prediction of the total loss factor is not possible, it is usually necessary
to measure the structural reverberation time (Section 3.11.3.3) which is simply related to the
total loss factor by Eq. 1.107. For some calculations, an estimate for the total loss factor is
adequate. For bending wave motion on masonry/concrete plates that are rigidly connected on
all sides, estimates for the total loss factor usually take the form (Craik, 1981)
X
ηi = ηii + √
f
(2.155)
The second term in Eq. 2.155 represents the sum of the coupling loss factors for structural
coupling; note that for these connected plates the coupling losses due to sound radiation are
negligible. When the plate boundaries are rigidly connected to other parts of the structure, the
following estimates can be used for X . Field measurements indicate that X = 1 is a reasonable
estimate for masonry/concrete walls and floors (Craik, 1981). The structural coupling losses in
the laboratory are not usually as high as in the field. To try and minimize variation between measurements in different laboratories, the requirement in the measurement standard is based on
X > 0.3 (ISO 140 Part 1). In the field and laboratory, 0.3 ≤ X ≤ 1 gives a reasonable indication
of the range for the total loss factor of masonry/concrete walls and floors (see Fig. 2.31).
For bending wave motion on single sheets/boards (e.g. plasterboard, chipboard) that form a
lightweight wall or floor, the measured total loss factor can be highly variable depending on
the type of frame and whether there is an absorbent material in the cavity. Rough estimates
for sheets/boards can be calculated using Eq. 2.155 with X = 0.4 for timber frames with empty
cavities, and X = 0.8 for timber or steel frames with mineral wool very close or touching the
sheet/board (100–5000 Hz). If a wall and floor has a significant decrease in vibration across it,
then the total loss factor is of questionable use.
2.6.6
Modal overlap factor
The modal overlap factor is calculated using Eq. 1.109.
159
S o u n d
I n s u l a t i o n
112
Total Loss Factor: X = 1, ηii = 0.01
Total Loss Factor: X = 0.3, ηii = 0.01
110
Internal Loss Factor: ηii = 0.01
Loss factor (dB)
108
106
104
102
100
98
50
80
125
200
315
500
800
1250 2000
One-third-octave-band centre frequency (Hz)
3150
5000
Figure 2.31
Total loss factor estimates for masonry/concrete plates that are rigidly connected on all sides along with the internal loss factor
used in this example.
2.7 Spatial variation in vibration level: bending waves on plates
Measurement and prediction of bending wave vibration on walls and floors requires an awareness of the spatial variation. The focus here is mainly on reverberant bending wave fields
although it should be noted that there will also be a net flow of bending wave energy across the
plate for which structural intensity can give an insight into the energy flow (Section 3.12.3.1.4).
2.7.1 Vibration field associated with a single mode
For a rectangular plate with simply supported boundaries the local mode shape, ψp,q , describes
the plate displacement for bending mode, fp,q(B) . By satisfying the boundary conditions and the
wave equation (Eq. 2.88) this gives
pπx
qπy
ψp,q = sin
(2.156)
sin
Lx
Ly
When at least one of the sine terms is zero, the displacement is also zero, so there will be
nodal lines of zero displacement perpendicular to the x- or y-axis. For a solid homogenous wall
or floor with simply supported boundaries we can use Eq. 2.156 to visualize the bending wave
vibration field as shown in Fig. 2.32. Real walls and floors are not all homogenous, isotropic
and rectangular plates with simply supported boundary conditions; they also have openings
(e.g. windows, doors) for which there can be a variety of boundary conditions. Uncertainty in
some or all of the boundary conditions for a wall or floor (and any openings) means that mode
shapes can rarely be accurately predicted.
The mode shapes for many real structures are too complex to calculate with analytical models
but they can be calculated using FEM (Zienkiewicz, 1977). This approach is used to calculate
the first eight local mode shapes for a wall with and without a window opening as shown in
160
Chapter 2
f11(B)
y
x
f21(B)
y
x
f12(B)
y
x
f32(B)
y
x
Figure 2.32
Mode shapes for bending modes on a plate with simply supported boundaries.
Fig. 2.33 (Hopkins, 2003b). Without an opening, this wall can be seen as representing one leaf
of a cavity separating wall between dwellings. With and without an opening this wall can be
seen as representing the inner leaf of the external flanking wall. The figure shows that introducing an opening can significantly alter the mode shape depending on the boundary conditions of
the opening. Mode shapes for plates without openings and the same boundary conditions on
all sides have simple symmetry; this no longer occurs when there is an opening. The thin strips
of wall between the wall boundaries and the boundaries of the opening tend to have very low
vibration levels when both boundaries are simply supported and the distance between them
is <λB /2.
Flanking walls contain openings such as doors and windows in many different positions with a
wide range of boundary conditions. In conjunction with the uncertainty in the wall dimensions
and material properties this also implies that a statistical approach to plate vibration is more
practical than deterministic calculations of the modal response.
2.7.2
Nearfields near the plate boundaries
When measuring bending wave vibration it is necessary to be aware of nearfields near the
boundaries of a beam or plate. In most situations we are only interested in quantifying the
reverberant vibration level due to propagating bending waves because, unlike the nearfield,
they transport energy. It is therefore necessary to exclude measurement positions close to the
161
22 Hz (f11)
43 Hz (f21)
67 Hz (f12)
78 Hz (f31)
88 Hz (f22)
123 Hz (f32)
127 Hz (f41)
142 Hz (f13)
22 Hz
47 Hz
60 Hz
73 Hz
83 Hz
117 Hz
118 Hz
148 Hz
32 Hz
50 Hz
76 Hz
91 Hz
116 Hz
142 Hz
154 Hz
170 Hz
45 Hz
89 Hz
126 Hz
154 Hz
165 Hz
172 Hz
218 Hz
246 Hz
Figure 2.33
Mode shapes for a wall with simply supported boundaries, with and without a window opening.
Row 1: Wall without an opening.
Row 2: Wall with an opening that has free boundaries.
Row 3: Wall with an opening where the top boundary is simply supported and other boundaries are free.
Row 4: Wall with an opening that has simply supported boundaries.
In each row the eight lowest mode frequencies are shown. For all modes except those in the first column, red and dark blue indicate maximum displacement in opposite directions.
Wall properties: x = 3.5 m, y = 2.4 m, h = 0.1 m, ρ = 600 kg/m3 , cL = 1900 m/s, ν = 0.2.
Reproduced with permission from Hopkins (2003).
Chapter 2
Incident wave
ηˆ⫹
ηˆ⫺
ηˆn⫹
Reflected wave
Exponentially
decaying
nearfield
B
o
u
n
d
a
r
y
x
x⫽0
x ⫽ ⫺∞
Figure 2.34
Bending waves incident upon, and reflected from the beam boundary.
boundaries that are affected by nearfield vibration. Hence it is useful to estimate the minimum
distance from the boundaries that should be used when taking measurements so that the
nearfield can be assumed to be negligible. This also needs to take into consideration the fact
that the exact boundary conditions are rarely known.
Bending waves on a plate impinge upon the boundaries from a variety of angles, but to simplify
the analysis only normal incidence is considered here. This conveniently allows us to follow
a bending wave propagating along a beam from −∞ towards the boundary at x = 0 (see
Fig. 2.34). This analysis will apply to beams if the bending stiffness for a beam is used to
determine the bending wavenumber, but we are looking at plates so the plate bending stiffness
is used. As we are restricting our attention to the end of the beam near x = 0 the relevant solution
for bending wave motion on a beam is taken from Eq. 2.58 as
η(x, t) = [η̂+ exp(−ikB x) + η̂− exp(ikB x) + η̂n+ exp(kB x)]exp(iωt)
(2.157)
where the complex wave coefficients are
η̂+ = |η̂+ | exp(iγ+ )
η̂− = |η̂− | exp(iγ− )
η̂n+ = |η̂n+ | exp(iγn+ )
(2.158)
Differentiating the real part of Eq. 2.157 with respect to time gives the velocity, which is squared
and time-averaged to give the mean-square velocity
( )
∂η 2
ω2
2
vz t =
=
(|η̂+ |2 + |η̂− |2 + |η̂n+ |2 exp(2kB x))
∂t
2
t
2
+ ω |η̂+ ||η̂− |cos(2kB x − γ+ + γ− )
+ ω2 |η̂+ ||η̂n+ |exp(kB x)cos(kB x − γ+ + γn+ )
+ ω2 |η̂− ||η̂n+ |exp(kB x)cos(kB x + γ− − γn+ )
(2.159)
From Eq. 2.159 we see that the mean-square velocity is described by the interaction between
the incident wave, the reflected wave and the nearfield. At this point we recall that the nearfield
must be considered in the solution to ensure that the wave equation and the boundary condition
are satisfied. However, to assess the effect of the nearfield on the vibration field we will ignore
163
S o u n d
I n s u l a t i o n
the nearfield components of Eq. 2.159 to give a mean-square velocity that is only valid at
distances from the boundary where the nearfield is negligible
vz2∗ t =
(
∂η
∂t
2 )
t
=
ω2
(|η̂+ |2 + |η̂− |2 ) + ω2 |η̂+ ||η̂− | cos (2kB x − γ+ + γ− )
2
(2.160)
To compare the mean-square velocities from Eqs 2.159 and 2.160 it is useful to reference each
of them to the mean-square velocity of the free-field incident bending wave,
v+2 t =
ω2
|η̂+ |2
2
(2.161)
The vibration field can now be shown using the velocity level differences,
10 lg
vz2 t
v+2 t
and 10 lg
vz2∗ t
v+2 t
where 0 dB corresponds to the level of the free-field incident wave.
To calculate Eqs 2.159 and 2.160, the wave coefficients can either be determined by assuming
perfect reflection at idealized boundary conditions or by using wave analysis to determine the
amplitudes of the reflected wave, transmitted wave and nearfields at a specific type of junction
(Craik, 1996; Cremer et al., 1973).
The wave coefficients for bending waves that are incident upon idealized boundaries can be
found in terms of the arbitrary constant, η̂+ , where (Mead, 1982a)
η̂− = −η̂+ and η̂n+ = 0 for a simply supported boundary
(2.162)
η̂− = −i η̂+ and η̂n+ = (i − 1)η̂+
for a clamped boundary
(2.163)
η̂− = −i η̂+ and η̂n+ = (1 − i)η̂+
for a free boundary
(2.164)
There is a nearfield generated at clamped and free boundaries, but not at simply supported
boundaries. Out of these three idealized boundaries it is only the free (i.e. unconnected) boundary that can be visually identified in a building. Free plate boundaries can occur along the top
or side edge of a wall, near door or window openings in a wall, and along the edges of a floating
floor. For any plate boundary that forms a junction with another beam or plate we have seen
that when calculating mode frequencies or mode shapes it is convenient to represent it as either
simply supported or clamped. In practice the boundary condition will rarely (if ever) be either
of these and will be more complex. This is partly due to the amplitude of the reflected wave
and nearfield varying with the angle at which the incident wave impinges upon the junction.
As a more realistic boundary condition we take the situation where the plate of interest (plate 1)
is connected to another plate (plate 2). For this example we will choose a right-angled corner
junction, referred to as an L-junction. The complex reflection coefficients for the bending wave
and the nearfield are R and Rn respectively, where
η̂− = R η̂+
164
and η̂n+ = Rn η̂+
(2.165)
Chapter 2
For a bending wave incident upon an L-junction the reflection coefficients are taken from wave
theory as (Cremer et al., 1973)
R =
ψ(1 − 2β2 − β1 β2 ) + χ(1 + 2β1 − β1 β2 ) + i[ψ(1 + β1 − β1 β2 ) + χ(−1 + β2 + β1 β2 )]
[ψ(−1 − β1 − 2β2 − β1 β2 ) + χ(−1 − 2β1 − β2 − β1 β2 )] + i[(ψ + χ)(1 − β1 β2 )]
(2.166)
Rn =
−1 + β2 − R(1 + β2 )
1 + iβ2
(2.167)
where
χ=
kB2
kB1
ψ=
2
Bp,2 kB2
2
Bp,1 kB1
β1 =
cB,p2 ρs2
cL,p1 ρs1
and
β2 =
cB,p1 ρs1
for plates.
cL,p2 ρs2
The minimum distance from the boundary at which the nearfield is negligible is frequencydependent. For free or clamped boundaries the magnitude of the difference between velocity
levels with and without the nearfield is <1 dB at distances >λB /3 m from the boundary, and
<0.1 dB at distances >3λB /4 m from the boundary. Measurements are usually carried out in
all frequency bands simultaneously, and the minimum distance will be largest at the lowest
frequency; a practical solution is to calculate a minimum distance based on 100 Hz.
Figure 2.35 shows the velocity level difference near a free boundary, a clamped boundary,
and an L-junction relative to the incident bending wave. These are shown for three different
plates that commonly form walls or floors. For bending waves that are reflected from free and
clamped boundaries the reflection coefficients are identical in magnitude and phase. Therefore
Eq. 2.160, which assumes that the nearfield is negligible, gives the same curve for both of these
idealized boundaries.
Close to the boundary at x = 0 the interference pattern is due to the interaction between
the incident wave, reflected wave and the nearfield. In this region, the large difference
between velocities calculated with and without the nearfield indicates the significant effect
of the nearfield. However, the nearfield has an exponential decay. Therefore it soon becomes
negligible with increasing distance from the boundary until there is no difference between the
mean-square velocity calculated using Eq. 2.159 or 2.160. For the three plates used in this
example at 100 Hz, the magnitude of this difference is <1 dB at distances between 0.2 and
0.8 m from the boundary. In practice, 0.25 m is chosen as a suitable distance for the majority
of constructions over the building acoustics frequency range (ISO 10848 Part 1). This takes
account of the fact that some walls and floors have relatively small dimensions, such that larger
distances would limit measurements to a very small area in the middle of the wall or floor.
At distances where the nearfield is negligible, there are sharp minima in the interference pattern
with the idealized boundary conditions. These are due to destructive interference between the
incident and reflected waves that have identical amplitudes. This is in contrast to the L-junction
where a fraction of the power that is incident upon the boundary is transmitted to the other plate.
Therefore the reflected wave has a lower amplitude than the incident wave which results in
minima which are shallower, and maxima that are reduced in height compared to the idealized
boundaries which are perfectly reflecting. This is a better representation of the real situation in
buildings where walls and floors are coupled together.
The focus here has been on the plate boundaries. However over the central area of many
lightweight walls and floors there are sheets of plasterboard or timber boards that are supported
165
(a) Lightweight aggregate blocks
(b) Cast in situ concrete
10
0
0
0
Only valid at distances
from the boundary
where the nearfield is
negligible
Plate details
h ⫽ 0.1 m
Lightweight aggregate blocks
⫺30
⫺40
⫺50
⫺20
Plate details
h = 0.1 m
In situ concrete
⫺30
⫺40
⫺1.75
⫺1.5
⫺1.25
⫺1
⫺0.75
⫺0.5
Distance from boundary, x (m)
⫺0.25
⫺2
⫺1.75
⫺1.5
⫺1.25
⫺1
⫺0.75
⫺0.5
⫺0.25
10
5
⫺15
⫺20
0
⫺5
Plate details
h = 0.1 m
In situ concrete
⫺10
⫺15
Forms an L-junction with a
plate of identical material
but with h = 0.2 m
⫺0.25
0
0
⫺5
Plate details
h = 0.0125 m
Plasterboard
(Natural gypsum)
⫺10
⫺15
Forms an L-junction
with a plate of identical
material and thickness
⫺20
⫺20
⫺25
⫺2
⫺1.75
⫺1.5
⫺1.25
⫺1
⫺0.75
Distance from boundary, x (m)
⫺0.5
⫺0.25
0
Velocity level difference (dB)
⫺1.25
⫺1
⫺0.75
⫺0.5
Distance from boundary, x (m)
0
5
Velocity level difference (dB)
⫺1.5
⫺0.25
10
⫺10
Forms an L-junction
with a plate of identical
material and thickness
⫺0.5
5
Velocity level difference (dB)
⫺1.75
⫺0.75
Distance from boundary, x (m)
⫺25
⫺2
⫺60
⫺1
0
10
⫺5
Only valid at distances
from the boundary
where the nearfield is
negligible
⫺40
⫺50
Distance from boundary, x (m)
Plate details
h ⫽ 0.1 m
Lightweight aggregate blocks
⫺30
⫺60
0
0
Boundary: L-junction
⫺20
Plate details
h ⫽ 0.0125 m
Plasterboard
(Natural gypsum)
⫺50
⫺60
⫺2
⫺10
Velocity level difference (dB)
Boundary: Free
⫺10
Velocity level difference (dB)
⫺20
Velocity level difference (dB)
10
⫺10
Boundary: Clamped
(c) Plasterboard
10
⫺25
⫺1
⫺0.75
⫺0.5
⫺0.25
0
Distance from boundary, x (m)
Figure 2.35
Velocity levels at different distances from various boundaries at x = 0 (free boundary, clamped boundary, and an L-junction); these levels are referenced to the incident bending wave (normal
incidence) to give a velocity level difference. The plates are formed from different materials (lightweight aggregate blockwork, cast in situ concrete, and plasterboard).
Chapter 2
by a frame of studs or joists. This frame can represent a discontinuity to the propagating wave
on the plate and also generate nearfields.
2.7.3 Diffuse and reverberant fields
As with diffuse sound fields in rooms, diffuse vibration fields on plates represent the ideal
situation rather than the reality. In practice we often need to assume that there are diffuse
fields in order to simplify the calculations for sound transmission. It is difficult to make general
statements about the likelihood of finding close approximations to diffuse vibration fields on
walls and floors without appearing to be overly pessimistic. This is partly because of the wide
variety of materials, plate dimensions, and plate geometries. It is also because the vibration
field can rarely be described as diffuse across the entire building acoustics frequency range.
At frequencies above the lowest mode it is usually more appropriate to refer to the vibration
field as reverberant rather than diffuse.
2.7.4 Reverberant field
For practical purposes, measurements are usually taken at random positions over the plate
surface to calculate the spatial average vibration level in frequency bands. An example is used
here to illustrate various aspects that give rise to the spatial variation.
Figure 2.36 shows vibration contour plots from measured velocity levels on a masonry wall.
The masonry wall is 100 mm thick with a 13 mm plaster finish on one side. The side with the
plaster finish faces into a reverberant room and is directly excited by the sound field in this
room where the sound source is broad-band noise from a loudspeaker. The left and right
hand side boundaries of the wall form corner junctions with other masonry walls. The lower
boundary forms an in-line junction with a similar masonry wall below but there are built-in
timber floor joists along this junction line. Vibration measurements were taken with different boundary conditions for the top boundary of the wall; firstly with a free (unconnected)
boundary, and secondly after the masonry wall was extended upwards to form an in-line
junction.
In the 50 Hz one-third-octave-band (Fig. 2.36a) there is evidence of the f21 mode shape when
there are junctions at all four boundaries. The mode frequency for the f21 mode is calculated
to be 37 Hz which is just outside this band but we should note that it is not unusual to predict to
an accuracy of plus or minus one-third-octave-band. The vibration level at 50 Hz varies by up
to 17 dB over the surface. This large variation needs to be borne in mind when determining the
number of measurement positions. At 1000 Hz (Fig. 2.36b) where M ≥ 1 and N ≥ 5 for bending
modes we see that regardless of the top boundary condition, the vibration levels are relatively
uniform in the central area of the wall. The vibration levels vary by < 8 dB. In the 3150 Hz band
(Fig. 2.36c) we might expect to find greater uniformity due to the large number of modes and
high modal overlap. Whilst this is generally true, the existence of very minor damage to the
plaster finish, and positions where the plaster did not bond strongly to the masonry result in
high vibration levels at a few points. In addition there is variation in fixing the accelerometers
with bees wax in the high-frequency range. Hence despite the increased uniformity of the
vibration field at high frequencies it is still necessary to average measurements from a number
of different positions to ensure that the spatial average value is representative.
167
S o u n d
I n s u l a t i o n
(a) 50 Hz one-third-octave-band (M ≈ 0.5, N ≈ 1)
Free boundary
14
98
11
15
13
9
7
9 10
4
6
12
14
3
10
Junction
13
3
3
14
4
1
2
11
12
12
12
8
7
13
10
9
15
13
9
4
10
11
4
12
5
7 6
8
12
10
3
4
9
9
13
14
13
8
12
11
8
9
11
9 10
7
8
9
11
8
10
12
7
9
11
8
Junction
12
11
10
8
12
13
14
10
15
12
10
7
5
7
8
6
9
9
8
6
7
10
1112
9
13
14
16
15
Junction
Junction
9
8
3
2 456
7
8
9
11
10
9
8
10
6
5
11
10
12
13
7
10
8
10
2
11
12
11
11
13
9
10
6
12
11
8
10
5
11
4
32
5
3
9
10
8
7
6
2
Junction
Junction
11
4
5
7
11
10
4
6
7
8 9
11
6
7
9
10
9
10 8
6
5
9
7
8
10
Junction
Figure 2.36
Vibration contour plot produced from a grid of measurements over the surface of a 100 mm masonry wall (x = 5.45 m,
y = 2.45 m, ρs = 166 kg/m2 , cL = 2690 m/s, built from solid blocks mortared together on all sides, 13 mm plaster finish). The
lower, left, and right wall boundaries are rigidly connected to other masonry walls. Two different boundary conditions are
shown for the top wall boundary: free boundary (upper plot) and a rigidly connected boundary forming an in-line junction to
an identical masonry wall (lower plot). The contours indicate the velocity levels in decibels relative to the lowest level that was
measured on the wall in that one-third-octave-band. The estimated modal overlap factor (M) and mode count (N) for bending
modes are given for each one-third-octave-band.
Measured data from Hopkins are reproduced with permission from ODPM and BRE.
2.7.5 Direct vibration field
For point excitation of bending waves on a plate there will be a direct vibration field near
the excitation point. In a similar way to rooms we can calculate the distance, rrd , at which the
168
Chapter 2
(b) 1000 Hz one-third-octave-band (M ≈ 2.5, N ≈ 20)
Free boundary
6
6
5
7
6
6
8
6
6
7
7
6
5
4
5
6
6
6
Junction
Junction
6
6
6
5
5
5
6
5
7
6
6
7
5
6
5
7
6
6
5
6
5
5
4
4 3
2
1
Junction
Junction
3
4
3
5
3
5
4
5
2
4
4
4
3
2
3
Junction
Junction
1
3
4
4
4
4
4
2
4
3
5
3
4
5
4
4
3
3
2
1
Junction
Figure 2.36
(Continued)
energy density in the direct field equals that in the reverberant field (Lyon and DeJong, 1995). A
cylindrical wave radiates out from the excitation position, and attenuation with distance occurs
due to both geometrical spreading of the wavefront and internal damping (see Fig. 2.37).
Internal damping, as described by the internal loss factor, ηint , causes an exponential decay
of vibration with distance. The decrease in bending wave amplitude due to internal damping
after travelling a distance, d, is determined by the factor (Cremer et al., 1973)
−ωηint d
exp
cg(B),p
(2.168)
169
S o u n d
I n s u l a t i o n
(c) 3150 Hz one-third-octave-band (M ≈ 5, N ≈ 62)
Free boundary
5
5
4
5
5
5
7
6
7
4
6
8
5
6
6
7
6
5
8
Junction
6
7
5
7
6
8
6
9
7
5
7
7
6
9
6
10
6
8
5
6
8
6
7
4
4
4
5
5
7
6
7
8
5
7
Junction
5
4
5
5
6
5
4
4
4
4
4
Junction
Junction
1
1
2
2 1
1
3
2
2
2
2 3
4
5
2
3
45
6
6
4
5
5
6
3
2
6
7
10
8 9
6
3
6
3
5
4
2
4
6
2
4
5 6
4
5
3
6 4
2
5
3
4
5
Junction
(Continued)
Figure 2.37
Cylindrical wavefronts from point excitation of bending waves on a thin plate.
3
4
2
1
2
2
Figure 2.36
170
3
2
3
Junction
Junction
3
5
4
2
3
4
2
2
2
3
3
1
2
3
Chapter 2
The power input due to point excitation is W . At a distance, d, from the excitation point, the
energy density of the direct field due to both geometrical spreading and internal damping is
W
−ωηint d
2
wd = ρs vd t =
exp
(2.169)
2πdcg(B),p
cg(B),p
On a finite plate, the direct wave is reflected from the plate boundaries. Ideally we need to
assume that there are irregularly shaped boundaries so that there is no coherence between the
reflected wave field and the direct wave field. In practice the boundaries of walls and floors are
usually regular but we can assume that that there are a sufficiently large number of reflections
to ensure that any coherence is lost. The mean free path is defined as the average distance
travelled by a wave after leaving one boundary and striking the next boundary (Eq. 2.112).
To determine the reverberant power input, we can assume that the average distance from the
excitation position to a plate boundary is approximately equal to half the mean free path.
The reverberant power is calculated by assuming that the direct field is attenuated with distance across the plate and then undergoes partial reflection at a plate boundary which has an
absorption coefficient, α. As we are currently interested in the vibration field on the plate we
refer to an absorption coefficient, whereas in Section 5.2 we will look at it from the perspective
of vibration transmission to other connected plates, the same value will then be referred to as
the transmission coefficient. The energy density of the reverberant field is therefore dependent
on the total loss factor, η, and is calculated from
W (1 − α)
−ωηint dmfp
wr = ρs vr2 t =
exp
(2.170)
ωηS
2cg(B),p
As an example, Fig. 2.38 shows the energy density at 1000 Hz due to the direct and the reverberant fields for a 10 m2 masonry wall at distances up to 1 m from the excitation position where
the total loss factor is representative of that in a transmission suite. The velocity level due to
Direct field
Reverberant field
Energy density (dB)
f ⫽ 1000 Hz
S ⫽ 10 m2
h ⫽ 0.1 m
cL ⫽ 2200 m/s
ηint ⫽ 0.01
η ⫽ ηint ⫹ 0.3f ⫺0.5
3 dB
rrd
0
0.2
0.4
0.6
Distance, d (m)
0.8
1
Figure 2.38
Energy density due to the direct and reverberant fields on a plate at distances up to 1 m from the excitation point.
171
S o u n d
I n s u l a t i o n
the direct field decreases by 3 dB every time the distance is doubled. The distance from the
excitation point at which the energy density in the direct field (Eq. 2.169) equals that in the
reverberant field (Eq. 2.170) is the reverberation distance, rrd . Due to the dispersive nature of
bending waves, rrd varies with frequency. The reverberation distance can either be calculated
by using a numerical solution to find the distance at which wd = wr , or, by using thin plate theory and assuming that there is perfect reflection at the plate boundaries to give the following
approximation
rrd ≈
ωηS
4πcB,p
(2.171)
Over the building acoustics frequency range, walls and floors with surface areas <20 m2 that
act as homogenous, isotropic plates typically have reverberation distances <0.75 m. When
measuring the reverberant velocity level a practical choice for the minimum distance between
an accelerometer and the excitation position is usually taken as 1 m (ISO 10848 Part 1).
2.7.6 Statistical description of the spatial variation
In a similar way to a room excited by a point source, the spatial variation of the mean-square
velocity on a plate excited by a point force can also be described by a gamma probability
distribution (Lyon, 1969). However, we are also interested in wave fields on plates that have
been excited by sound fields, and/or by structural waves that impinge upon the plate junctions
from connected plates or beams. This means that the mean-square velocity at any point on the
plate is determined by a large number of variables. If these are independent random variables
(or not too strongly associated), then, by invoking the central limit theorem, the probability
distribution of a sum or average of many small random quantities can be approximated by a
normal (Gaussian) distribution. It can therefore be assumed that the spatial variation of the
mean-square velocity will have a log-normal probability distribution, and the velocity level in
decibels will have a normal probability distribution (Lyon and DeJong, 1995).
Compared with sound pressure in rooms, it is more difficult to calculate reasonable estimates
for the standard deviation for the spatial variation of vibration on walls and floors. We will
assume that the uncertainty due to time-averaging is negligible. In frequency bands above the
fundamental mode of the plate, the standard deviation of the velocity level in decibels, σdB , is
dependent upon the mode count in the frequency band, N, and the modal overlap factor, M,
(Lyon and DeJong, 1995)
!
&
"
'
"
N −1
3
#
1+
(2.172)
1+
σdB ≈ 43 lg 1 +
πM
πM
For masonry/concrete walls or floors, Eq. 2.172 tends to overestimate the standard deviation
at low- and mid-frequencies, and underestimate it at high frequencies. For masonry walls with
or without a plaster finish the standard deviation usually starts to increase above 2000 Hz due
to three main factors: imperfections over the measurement surface, a non-reverberant field
where the vibration level decreases with distance across the plate, and difficulties in obtaining
a strong fixing of the accelerometer to the surface using beeswax. An example is shown
in Fig. 2.39 for airborne excitation of a 100 mm masonry wall with a 13 mm plaster finish.
For masonry/concrete walls or floors a generalized curve is included on the figure because
the curve shape can usually be described using just four points. This generalized curve for
masonry/concrete elements only indicates the general shape, not the typical values for these
172
Chapter 2
6
Measured
Standard deviation, σdB (dB)
5
Predicted
Generalized curve shape
4
3
2
1
0
50
80
125 200 315 500 800 1250 2000 3150 5000
One-third-octave-band centre frequency (Hz)
Figure 2.39
Comparison of measured and predicted standard deviations for the velocity level (airborne excitation) on a 100 mm masonry
wall with a 13 mm plaster finish.
Measured data from Hopkins are reproduced with permission from ODPM and BRE.
four points; these depend on the specific type of masonry (e.g. solid bricks, hollow blocks),
surface finish, material properties, wall dimensions and type of excitation. For measurement
purposes it is useful to be aware of the increase in the standard deviation in the high-frequency
range. The increase in the mode count and modal overlap at these high frequencies does not
necessarily lead to a reduction in the standard deviation.
2.7.7 Decrease in vibration level with distance
With structural excitation there can be a significant decrease in vibration across a wall or
floor. This has implications for modeling vibration transmission between structures as well
as when measuring impact sound insulation and flanking transmission. Bending waves are
of particular importance so we will only consider the decrease that occurs for bending wave
motion perpendicular to the surface of the plate.
Any decrease in bending wave vibration with distance due to internal damping only tends
to become apparent in the high-frequency range because most materials used for walls and
floors have relatively low internal loss factors. For a plane bending wave propagating on a
plate, Eq. 2.168 is used to give the decrease in the vibration level, Lint , in decibels; this is
due to internal damping after travelling a distance, d, where
Lint =
10 ωηint d
ln 10 cg(B),p
(2.173)
As a plane bending wave travels across a plate it will also be attenuated each time it impinges
upon a plate boundary. This will be referred to as excess attenuation, i.e. the attenuation that
occurs in addition to the internal damping. We will take the situation where bending waves
on a plate are excited along one of its boundaries due to wave transmission from another
plate. For bending wave transmission across the plate in the direction away from the junction,
173
S o u n d
I n s u l a t i o n
the plate can be considered as a series of one-dimensional (beam-like) elements that are
orientated parallel to the junction line; each element is of length, L. Absorption occurs at both
ends of these elements as the waves are partially absorbed by the plate boundaries due to
wave transmission to other connected plates. The lower and upper boundaries of the plate
have absorption coefficients, αL and αU respectively, hence the average value is (αL + αU )/2.
For these one-dimensional elements, energy impinges upon the boundaries, cg(B),p /dmfp times
every second, where the mean free path is equal to L. By considering the power absorbed by
the boundaries, Wabs , a loss factor can be determined for the excess attenuation, ηexcess using
Wabs = E
cg(B),p (αL + αU )
= ωηexcess E
L
2
(2.174)
The decrease in the vibration level in decibels due to excess attenuation, Lexcess , after travelling a distance, d, is
10 (αL + αU )
ωηexcess d
Lexcess = 10 lg exp
d
(2.175)
=
cg(B),p
ln 10
2L
The total decrease in the vibration level in decibels, Ltotal , due to internal damping and excess
attenuation is then given by
Ltotal = Lint + Lexcess
(2.176)
Significant decreases usually only occur at frequencies above the thin plate limit for bending
waves. In these situations the group velocity for bending waves on thick plates, cg,B(thick) , needs
to be used instead of the thin plate group velocity, cg,B . An example is shown in Fig. 2.40 for
a 100 mm masonry wall in the 3150, 4000, and 5000 Hz one-third-octave-bands that are all
above the thin plate limit (Hopkins, 2000). For the 4000 and 5000 Hz bands, the attenuation
due to internal damping does not account for all the decrease in vibration; hence it is necessary
to include the excess attenuation. The excess attenuation is important here because the upper
plate boundary is a straight junction to an identical plate for which αU is estimated to be unity.
However, for thick plates it is difficult to obtain reasonable estimates for the power absorbed at
the plate boundaries. Figure 2.41 shows the vibration contour plot for the 5000 Hz one-thirdoctave-band to illustrate the decrease in vibration level across the surface of the same wall
with and without a window (Hopkins, 2000).
Although Lint and Lexcess can account for the decrease in vibration level across homogeneous walls and floors, the modular and periodic nature of many building elements causes
other loss mechanisms. Apart from relatively homogeneous building elements such as those
built from cast in situ concrete, the majority of heavyweight walls and floors are built from
components such as bricks, blocks, beams, or slabs which are fixed together in a variety of
ways. The connections between these individual components can be highly variable, and/or
very weak. A common example of a non-homogeneous floor is a beam and block floor as
shown in Fig. 2.42. For this particular floor the blocks are not bonded to the beams or to adjacent blocks so the decrease in vibration level is shown using measurements on the beams.
In practice, the vibration level on the beams does not keep decreasing with distance because
flanking paths to the more distant beams become more important with increasing distance.
Large floors (≈200 m2 ) built from individual concrete slabs with a screed finish can also show
significant attenuation with distance (Steel et al., 1994).
A decrease in vibration with distance can also occur with spatially periodic plates. An example
is a ribbed plate where the cross-section repeats with a certain repetition distance. In buildings,
174
Chapter 2
10
9
Decrease in vibration level (dB)
8
Decrease in
vibration level
One-third-octave-bands
3150 Hz
4000 Hz
5000 Hz
Predicted ∆Lint
Predicted ∆Lint ⫹ ∆Lexcess
Measured
7
6
Excitation
5
4
3
2
1
0
0.29
0.58
0.86
1.15
1.44 1.73 2.02
Distance (m)
2.30
2.59
2.88
3.17
Figure 2.40
Decrease in vibration level with distance across a 100 mm masonry wall (Lx = 4.04 m, Ly = L = 2.38 m, ρs = 70 kg/m2 ,
cL = 2370 m/s, ηint = 0.0125, solid aircrete blocks with mortar on each side, 13 mm plaster finish). Each measured value
corresponds to the average of seven measurement points along a line parallel to the junction. To calculate a decrease in level
these values are referenced to the line that is 0.58 m from the junction. The wall was excited via vibration transmission across
a corner junction (junction line: Ly ) using hammer excitation of the other masonry wall that formed the junction. Note that both
walls were rigidly connected to other walls/floors on all sides.
Measured data are reproduced with permission from Hopkins (2000).
ribbed plates produced from a single material can be found in the form of profiled cast in situ
concrete floors. Lightweight walls and floors also form periodic plates where sheet materials
are connected to a framework of beams; examples include timber joist floors or studwork walls.
These may be periodic in one or two directions. In addition, the beams and the plates usually
have different material properties.
Here we will briefly consider plates that are periodic in a single direction and look at the propagation of bending waves in the direction perpendicular to the ribs/beams. This is the direction that
usually has the highest attenuation with distance. When the bending wavelength is much larger
than the repetition distance, a periodic plate can be modelled as an orthotropic plate (Section
2.3.3.2). When the wavelength is similar, or smaller than the repetition distance, propagation
of bending waves becomes increasingly complex. At these frequencies, periodic structures
can generally be considered as one of two types: a precise periodic structure where there
is exact periodicity with exactly the same material properties in each repeating section, or a
non-precise periodic structure where there is irregularity in the periodicity, dimensions, boundary conditions, and/or material properties. For precise periodic structures, structure-borne
sound propagation is characterized by ‘pass bands’ and ‘stop bands’ at different frequencies
(Cremer et al., 1973; Mead, 1982b). As implied by their names, waves either propagate freely
across the ribs/beams, or they are highly attenuated. However, high levels of attenuation can
only occur across a relatively small number of ribs/beams. This is due to flanking transmission when the plate is connected to other walls and floors and/or conversion between wave
175
S o u n d
I n s u l a t i o n
y (m)
2.1
⫺17.5
⫺27.5
⫺25
⫺22.5
1.8
⫺20
⫺22.5
⫺25
Junction
⫺25
1.5
⫺20
1.2
⫺22.5
⫺25
⫺25
⫺27.5
⫺30
⫺20
0.9
⫺27.5
⫺25
0.6
3.46
0.3
3.74
⫺25
⫺22.5
⫺17.5
⫺20
⫺17.5
0.29
0.58
⫺20
0.86
1.15
1.44
1.73
2.02
2.30
2.59
2.88
3.17
2.1
⫺22.5
⫺25
⫺27.5
⫺27.5
1.8
⫺27.5
1.5
Junction
⫺30
1.2
⫺30
⫺27.5
⫺32.5
0.9
⫺30
⫺32.5
⫺30
⫺20
⫺15
0.6
⫺27.5
⫺30
⫺25
⫺17.5
0.29
0.58
⫺25
⫺22.5
0.86
1.15
1.44
1.73
2.02
x (m)
2.30
2.59
2.88
3.17
3.46
0.3
3.74
Figure 2.41
Vibration contour plot for the 5000 Hz one-third-octave-band produced from a grid of measurements over the surface of a
100 mm masonry wall. The wall is connected on all sides but only excited along one junction line (indicated on the diagram) by
mechanical excitation of the connected wall. Results are shown for the wall with and without a window opening. The decrease
in vibration level with distance is shown by plotting −Dv,ij (where Dv,ij is the velocity level difference), i.e. the −15 dB contour
line corresponds to vibration levels on this wall that are 15 dB lower than the spatial average vibration level on the connected
wall that is being excited. Refer to Fig. 2.40 for wall properties.
Measured data are reproduced with permission from Hopkins (2000).
types at each junction of the plate and the beams. We are usually interested in the situation
where bending waves are excited on a plate. These are converted to in-plane waves at the
junction with each beam or rib. Point excitation of bending waves gives a rapid decrease
in the bending wave vibration away from the excitation point. This initially steep gradient
176
Chapter 2
9
One-third-octave-bands
6
Vibration relative to beam A (dB)
3
0
⫺3
125
250
500
1000
2000
4000
⫺6
⫺9
⫺12
⫺15
⫺18
A
B
C
D
E
F
⫺21
⫺24
⫺27
⫺30
⫺33
A
B
C
D
E
F
Beam
Figure 2.42
Decrease in vibration level with distance across the beams of a beam and block floor. The propagation direction in the
measurements is perpendicular to the 150 mm thick solid concrete beams (≈ 85 × 50 mm, ρl = 31 kg/m). The 150 mm deep
blocks are solid masonry (ρ = 2000 kg/m3 ) with a width of 440 mm – except between beams D and E where the width is
215 mm. Each measured value corresponds to the average of four points on a beam. To show the decrease in level the vibration
on each beam is referenced to beam A which is closest to the junction. The floor was excited via vibration transmission across
the junction using hammer excitation on one of the masonry walls that formed the junction.
Measured data from Hopkins are reproduced with permission from ODPM and BRE.
becomes much shallower with increasing distance due to conversion from bending to inplane waves at the beam/rib junctions; these are converted back to bending waves at more
distant beam/rib junctions (Heckl, 1964; Mead and Markus, 1983). For non-precise periodic
structures, there is potential for a localization effect which confines the highest vibration levels to the vicinity of the excitation (Hodges, 1982). For the above reasons the change in
vibration level across a periodic plate is rarely a simple function of frequency; an example
is shown in Fig. 2.43 for a timber raft floating floor that represents a non-precise periodic
plate.
Figure 2.44 shows measurements on a timber joist floor from Nightingale and Bosmans (1999).
There is a rapid decrease in the vibration level with distance up to the butt joint between adjacent
boards of OSB. After this the vibration level becomes more uniform with increasing distance.
These features are indicative of a periodic plate as discussed above. The fact that the plate
surface is formed from individual boards means that there will almost always be a joint or
discontinuity between boards at a short distance from any excitation point. Different types of
connection between sheets or boards will alter the transmission of bending wave across these
joints as well as the conversion to other wave types. In the direction parallel to the joists the
attenuation with distance is much lower; the only significant attenuation occurs in the highfrequency range across tongue and grooved joints between adjacent boards (Nightingale and
Bosmans, 1999).
177
S o u n d
I n s u l a t i o n
1
A
2
3
4
5
450 mm
30
1
Vibration level relative to batten A (dB)
27
2
24
3
4
21
5
18
15
12
9
6
3
0
⫺3
50
80
125
200
315
500
800
1250
2000
One-third-octave-band centre frequency (Hz)
3150
5000
Figure 2.43
Decrease in vibration level with distance across a timber raft floating floor. Measurements used impulse excitation with a
plastic headed hammer at two different positions above batten A. Two random positions were used for the accelerometers on
the chipboard above each batten.
Timber raft details: 18 mm tongue and grooved chipboard (ρs = 10.6 kg/m2 , cL = 2410 m/s) screwed at 300 mm centres to
45 × 45 mm timber battens (ρl = 0.9 kg/m, cL = 5490 m/s), floating on mineral wool on 150 mm concrete slab.
Measured data from Hopkins are reproduced with permission from ODPM and BRE.
2.8 Driving-point impedance and mobility
To quantify the power input into a plate or beam from a mechanical source it is necessary to
know its impedance or mobility. The driving-point impedance or mobility that applies to point
force excitation of bending waves is of particular importance. Point force excitation occurs with
many structure-borne sound sources in buildings including the hammers of the ISO tapping
machine. For impact sound insulation we need to consider random excitation positions in frequency bands and it is the spatial average value over the floor surface that is of particular
interest. For other applications, such as the positioning of machinery or other equipment it is
sometimes necessary to look at the driving-point impedance or mobility at specific positions
and at single frequencies.
A force can be considered as being applied at a single point when the excitation area is
much smaller than the structural wavelength. For excitation by a point force, the driving-point
impedance, Zdp , is the ratio of the complex force to the complex velocity at the excitation point,
Zdp =
178
F
v
(2.177)
Chapter 2
ISO tapping
machine
A
Butt joint
between the
two boards
200 200
mm mm
9
One-third-octave-bands
Vibration relative to position A (dB)
6
125 Hz
250 Hz
500 Hz
1000 Hz
2000 Hz
4000 Hz
3
0
⫺3
⫺6
⫺9
⫺12
⫺15
⫺18
⫺21
⫺24
⫺27
⫺30
⫺33
0.0 0.2 0.4 0.6 0.8 1.0 1.2 1.4 1.6 1.8 2.0 2.2 2.4 2.6 2.8 3.0 3.2
Distance from position A (m)
Figure 2.44
Decrease in vibration level with distance across the walking surface of a timber joist floor. The propagation direction for the
measurements is perpendicular to the solid timber joists (38 × 235 mm, joist spacing: 400 mm centres) and includes a butt
joint between two adjacent boards of 16 mm OSB (screwed to joists at 300 mm centres). Each gridline corresponds to a
measurement point on the OSB that is either above or midway between the joists. Excitation using the ISO tapping machine.
Measured data are reproduced with permission from Nightingale and Bosmans (1999) and the National Research Council
of Canada.
and the driving-point mobility, Ydp , is defined as
Ydp =
v
1
=
Zdp
F
(2.178)
For a point force, the structural power input can be written in terms of either the impedance or
the mobility,
*
+
1
1
2
∗
2
2
(2.179)
Re{Ydp }
= Frms
Win = Fvt = Re{Fv } = vrms Re{Zdp } = Frms Re
2
Zdp
where ∗ denotes the complex conjugate. This is only applicable to a source which has negligible
impedance compared to the structure itself; discussion of this point with regards to the hammers
of the ISO tapping machine is in Section 3.6.3.
For finite reverberant structures, the driving-point mobility is determined by the damped modes
of the structure. Before looking at a multi-modal structure such as a plate it is useful to consider
179
S o u n d
I n s u l a t i o n
F(t)
m
η(t)
R
k
Figure 2.45
Mass–spring–dashpot system.
a lumped element model for a linear mass–spring–dashpot system (see Fig. 2.45). This single
degree-of-freedom system allows us to look at the main features of the driving-point mobility
for a single mode of vibration.
A sinusoidal (simple harmonic) force, F(t), with angular frequency, ω, is applied to the single
degree-of-freedom system,
F(t) = F̂ exp(i(ωt + φ))
(2.180)
where φ is an arbitrary phase term.
The equation of motion for this system is given by
F(t) = m
∂2 η
∂η
+R
+ kη
∂t 2
∂t
(2.181)
where η is the displacement, m is the mass, R is the damping constant, and k is the spring
stiffness.
For simple harmonic motion, Eq. 2.181 can be written in terms of the velocity as
k
F(t) = R + i ωm −
v(t)
ω
(2.182)
If the system has no damping (i.e. R = 0), Eq. 2.182 indicates that the system response is a
maximum at the undamped resonance frequency, f0 , where
k
1
ω0
=
(2.183)
f0 =
2π
2π m
At this point, η is no longer needed to represent the displacement. Therefore we can use it to
describe the damping in terms of the loss factor. This is more practical than using the damping
constant as we almost always discuss the modal response of plates and beams with reference
to the various loss factors. The damping constant is related to the loss factor, η, by
√
(2.184)
R = η km = ηω0 m
The driving-point mobility is now given by Eqs 2.182 and 2.184 as
1
v
=
Ydp =
F
ηω0 m + i ωm − ωk
(2.185)
From Eqs 2.179 and 2.185, the power input into the system is given by
2
2
Win = Frms
ηω0 m|Ydp |2
Re{Ydp } = Frms
(2.186)
Hence the real part and the magnitude of the driving-point mobility are needed for practical purposes (see Fig. 2.46). The response at resonance is characterized by a peak in the
180
Driving-point mobility, 20 lg (Y ) (dB)
Chapter 2
Real part, Re {Ydp}
Damping control within the
half-power bandwidth, ∆f3dB
Stiffness
control
Magnitude,Ydp
Mass
control
Approximation for the
magnitude in the
stiffness control region
Approximation for the
magnitude in the mass
control region
f0
Frequency (Hz)
Figure 2.46
Driving-point mobility for a mass–spring–dashpot system.
driving-point mobility. When the damping is sufficiently low (η < 0.3), the peak in either the real
part or the magnitude of the driving-point mobility occurs at the undamped resonance frequency
given by Eq. 2.183. From Eq. 2.186 we see that by calculating the magnitude in decibels using
20 lg|Ydp |, the 3 dB down points from the peak will give the half-power bandwidth, f3dB . Note
that the curve for |Ydp | is much broader than for Re{Ydp }. Within the half-power bandwidth, |Ydp |
is under damping control. At lower frequencies |Ydp | tends towards pure stiffness control and
at higher frequencies towards pure mass control. Approximations for the driving-point mobility
under stiffness or mass control are found from Eq. 2.185, giving
ω
under stiffness control
k
1
under mass control
≈
iωm
Ydp ≈ i
(2.187)
Ydp
(2.188)
Characteristics shown by this single degree-of-freedom system form a useful background for
the analysis of multi degree-of-freedom systems such as beams and plates.
In this section the focus has been on the driving-point mobility due to its relevance to the power
input. If the velocity is measured at a point on the structure that is different to the excitation
point, it is referred to as the transfer mobility, Ytr .
2.8.1
Finite plates (uncoupled): Excitation of local modes
Earlier sections discussed local modes on plates with idealized boundary conditions that were
uncoupled and isolated from any other structure. A plate has an infinite number of modes; hence
it has an infinite number of degrees-of-freedom. However, the single degree-of-freedom system
that was discussed in the previous section can be used to represent each mode of vibration.
Assuming linearity, the response of the uncoupled plate can be found by superposing the
response of all the local modes.
181
S o u n d
I n s u l a t i o n
Of particular interest for impact sound insulation is the excitation of bending modes on a plate by
a point force acting perpendicular to the surface. The principle of superposition can be used to
give an analytic solution for bending vibration of a thin rectangular plate. This can then be used
to give the driving-point mobility at a position (x, y) on a rectangular plate (Cremer et al., 1973)
Ydp =
∞
∞
2
ψp,q
(x, y)
i4ω
2
ρs S
[ωp,q (1 + iη)] − ω2
(2.189)
p=1 q=1
where the local bending mode shape, ψp,q (x, y) for angular mode frequency, ωp,q , is given by
Eq. 2.156 for a rectangular plate with simply supported boundaries.
Figure 2.47 shows the real part of the driving-point mobility for a concrete slab. Equation 2.189
is initially used to calculate the mobility at two chosen positions, A and B. Position A in the centre
of the plate excites f11 and f22 as indicated by the two peaks in the mobility spectrum; it does not
excite f21 and f12 because position A lies on the nodal line of these modes. In contrast, position
B excites all of the first four modes as seen by the peaks in the mobility spectrum. The fact
that the peaks correspond to excited modes is important; for a given mean-square force input,
peaks in the driving-point mobility will correspond to peaks in the power input (Eq. 2.179).
Equation 2.189 is then used to calculate a spatial average value for comparison with the
measured mobility; this shows that the local mode assumption gives a reasonable estimate of
the mobility spectrum. In general, the fluctuations significantly decrease with frequency and
the mobility tends towards a single value. This value corresponds to the driving-point mobility
for an infinite plate; it is shown in Fig. 2.47 for comparison with the measured values because
it will soon be introduced in Section 2.8.3.
2.8.2
Finite plates (coupled): Excitation of global modes
Local modes are extremely useful in the analysis of sound and vibration, but the local mode
viewpoint deliberately takes a blinkered view by ignoring the interaction with other parts of
the structure to which the plate or beam is coupled. For example, if we take a rectangular
separating wall in a building, its four boundaries will be connected to flanking walls and floors.
The local modes can be calculated for each of the uncoupled and isolated walls and floors.
However, the system of coupled walls and floors also has its own natural modes of vibration;
these can be referred to as global modes. For complex built-up structures, the global mode
frequencies and mode shapes can be determined using analytic solutions or FEM. Taking the
global mode approach to extremes will result in a model of the entire building. Whilst this may
be appropriate for very low-frequency vibration in buildings (below 20 Hz), it is not necessary
for the prediction of sound insulation where we deal with smaller wavelengths in the building
acoustics frequency range. Local modes remain a very convenient way of modelling vibration;
in Chapters 4 and 5 we will look at the prediction of sound insulation based on SEA, which
relies on the local mode assumption. However, it is worth noting that natural modes of vibration
can be modelled from both a local and a global viewpoint.
Figure 2.48 illustrates the effect of global modes on the driving-point mobility (real part) of five
connected walls that form an H-block (Hopkins, 2003a). Below 100 Hz, the separating wall only
has two local modes (assuming three simply supported boundaries and one free boundary);
hence the spatial average mobility for the uncoupled separating wall only has two peaks. In contrast to this there are 22 global modes of the H-block below 100 Hz. These global modes result
in the mobility curve for the coupled separating wall in the H-block having many more peaks
182
Chapter 2
Lx
A
Ly /2
B
Ly /3
⫺90
Ly
140 mm cast in situ concrete slab
(rigidly connected to masonry walls along all edges)
Lx ⫽ 4.2 m, Ly ⫽ 3.6 m, ρs ⫽ 345 kg/m2,
cL ⫽ 3800 m/s, η ≈ 0.005 ⫹ 0.3f ⫺0.5
fB(thin) ⫽ 1485 Hz, f11 ⫽ 32 Hz
Measured
< Re{Ydp} >s : spatial average of 10 random positions
that were ≥0.5 m from the boundaries (2 Hz FFT lines)
Lx /5
Lx /2
Infinite
plate
⫺110
20 lg (Re{Ydp}) (dB)
⫺130
Analytic model for a finite plate (local modes)
< Re{Ydp}>s : spatial average of 100 random positions
that were ≥0.5 m from the boundaries
⫺90
⫺110
⫺130
Re{Ydp} at position A
Re{Ydp} at position B
⫺90
⫺110
⫺130
f11
f21 f12
f22
⫺150
10
100
Frequency (Hz)
1000
Figure 2.47
Real part of the driving-point mobility for a finite plate. The symbol next to the x -axis indicates the local mode frequencies
for the uncoupled plate with simply supported boundaries.
Measured data from Hopkins are reproduced with permission from ODPM and BRE.
than the uncoupled wall. The existence of many peaks is also seen in the measured curve,
although only the general trend is followed, not the precise details. When the local mode density is low, the predicted mobility has deep, wide troughs between the modes that are not under
damping control. The abundance of global modes tends to reduce the depth and width of these
troughs. Similar features are seen for the flanking wall. In this case the fundamental local mode
(assuming two simply supported boundaries and two free boundaries) is just below the lowest
global mode. The global modes cause the first peak in the mobility to be shifted upwards in
frequency; this is in agreement with the measured data. If there is uncertainty in classifying the
183
S o u n d
I n s u l a t i o n
(a) Separating wall
⫺80
Measured
(10 random positions that were
≥0.5 m from the boundaries,
2 Hz FFT lines)
⫺100
20 lg ( <Re{Ydp}>s ) (dB)
⫺120
Separating
wall
⫺80
FEM
(16 random positions
that were ≥0.5 m
from the boundaries)
<Re{Ydp}>s
⫺100
FEM model of uncoupled separating wall
⫺120
FEM model of H-block
Local modes of uncoupled separating wall
⫺140
Global modes of the H-block
⫺160
⫺180
1
10
Frequency (Hz)
100
Figure 2.48
Effect of global modes on the spatial average driving-point mobility for coupled plates. This example uses an H-block of
five rigidly connected masonry walls which are free-standing on a wide 300 mm thick concrete floor. The 215 mm separating
wall (Lx = 4.5 m, Ly = 2.5 m) and the four 100 mm flanking walls (Lx = 3.6–4.1 m, Ly = 2.5 m) were all built from solid dense
aggregate blocks (ρ = 2000 kg/m3 ). The walls were slightly orthotropic; this was modelled in the FEM model using the
measured longitudinal wavespeed for each wall in the two orthogonal directions. Local and global mode frequencies were
also calculated using the FEM model.
Reproduced with permission from Hopkins (2003).
boundary conditions of a plate it would be useful if the peaks in the mobility could be used to estimate local mode frequencies for comparison with different idealized boundary conditions. This
example indicates that this approach is prone to error; in some cases a rigidly connected boundary might be classified as clamped rather than simply supported due to the effect of the global
modes. When masonry/concrete plates are rigidly connected on all sides, the total loss factor of
each connected wall will be quite high and the global modes will not usually be prominent in the
measured driving-point mobility. An example of this was seen with the concrete slab in Fig. 2.47.
The above example for coupled masonry walls indicates that even under laboratory conditions,
with measurement of the quasi-longitudinal phase velocity in both directions, and plates with
some indisputable boundary conditions (i.e. free), it is still not possible to predict the fine
structure of the driving-point mobility using deterministic models. This provides an incentive to
try and use statistical models (wherever possible) to describe the vibration of building structures.
Such models often make use of the fact that for the purpose of calculating the power input many
finite plates can be treated as infinite plates.
184
Chapter 2
(b) Flanking wall
⫺60
Measured
(10 random positions that were ≥0.5 m from the boundaries,
2 Hz FFT lines)
⫺80
<Re{Ydp }>s
20 lg ( <Re{Ydp}>s ) (dB)
⫺100
⫺70
FEM model of uncoupled
flanking wall
⫺90
FEM
(16 random positions that were
≥0.5 m from the boundaries)
FEM model of H-block
Local modes of uncoupled
flanking wall
⫺110
Global modes of the
H-block
⫺130
⫺150
1
10
Frequency (Hz)
100
Figure 2.48
(Continued )
2.8.3 Infinite beams and plates
In many situations it can be assumed that the structure is of infinite or semi-infinite extent, such
that waves travelling out from the excitation point do not return to that point. A simple link can
therefore be made to excitation of a finite plate that has completely absorbing boundaries so that
no waves return to the excitation point. In practice, waves emanating from the excitation point
are only partly absorbed at the boundaries and the reflected waves can return to the excitation
point. However, the infinite plate assumption is still appropriate when these reflected waves are
incoherent with the excitation signal and at a significantly lower level than at the excitation point.
Excitation of bending waves is now considered for the two situations shown in Fig. 2.49: (a) in
the central part of an infinite plate or beam to represent positions on finite structures that are
far away from the boundaries and (b) at a position on the edge of a semi-infinite plate or beam.
2.8.3.1 Excitation in the central part
For a thin plate of infinite extent, the driving-point impedance for excitation of bending waves
in the central part is real and given by (Cremer et al., 1973)
Zdp = 8 Bp ρh = 2.3ρcL,p h2
(2.190)
(Note: For an orthotropic plate, the effective bending stiffness (Eq. 2.97) can be used.)
185
S o u n d
I n s u l a t i o n
(a)
(b)
Figure 2.49
Excitation of bending waves: (a) in the central part of an infinite plate or beam and (b) at the edge of a semi-infinite plate or
beam.
For a thin beam of infinite extent, the driving-point impedance for excitation of bending waves
in the central part is complex and given by (Cremer et al., 1973)
Zdp = (1 + i)2ρScB,b = (1 + i)2.67ρS cL,b hf
(2.191)
where S is the cross-sectional area of the beam.
When predicting the impact sound insulation of floors it is usually the low and mid-frequency
range that are of most importance. At these frequencies most floors or floating floors act as
thin plates. The driving-point mobility for a thick plate is complex rather than real (Cremer et al.,
1973). However, it is the real part that determines the power input. As the high-frequency range
is usually of lesser importance, any errors incurred using thin plate theory above the thin plate
limit can usually be neglected.
We have already seen that peaks in the driving-point mobility occur due to excited modes.
As these peaks are under damping control, the infinite plate assumption becomes more
appropriate as the plate damping increases and the mode spacing decreases.
Figure 2.50 compares infinite plate theory (Eq. 2.190) with the measured real part of the
mobility on different finite plates in buildings that can be considered as homogeneous. The
measured data is shown in one-third-octave-bands as these are relevant to impact sound
insulation measurements. In general, the assumption of a thin infinite plate becomes more
appropriate with increasing frequency (assuming f < fB(thin) ). The largest differences between
finite and infinite plates tend to occur in the low-frequency range where there are relatively few
bending modes. These trends can be seen with the 140 mm concrete slab; this is the same
slab as previously seen in Fig. 2.47. The fundamental local mode is below 50 Hz, and there
are only two local modes between 50 and 100 Hz (f21 and f12 ). The resulting deep trough and
high peaks between 50 and 100 Hz causes the measured mobility to be significantly different
to the infinite plate.
Measurements on the two floating sand–cement screed floors illustrate a number of points.
Firstly we note that these plates are square shaped. Square isotropic plates have pairs of
186
Chapter 2
⫺100
Measured values with 95% confidence intervals
Infinite plate
⫺106
⫺112
⫺118
⫺124
215 mm dense aggregate masonry walls (fair-faced)
(three similar walls attached to different masonry flanking walls)
Lx ≈ 4.5 m, Ly ≈ 2.4 m, ρs ⫽ 430 kg/m2, cL ≈ 3130 m/s, fB(thin) ≈ 797 Hz
⫺130
20 lg (<Re{Ydp}>s) (dB)
⫺58
⫺64
⫺70
22 mm tongue & grooved chipboard (floating floor)
on 45 mm foam (s⬘ ⫽ 4 MN/m3)
Lx ⫽ 4.2 m, Ly ⫽ 3.6 m, ρs ⫽ 15 kg/m2, cL ⫽ 2200 m/s, fB(thin) ⫽ 4942 Hz
⫺94
⫺100
⫺106
⫺112
65 mm sand-cement screed (floating floor)
on 5 mm closed-cell foam (s⬘ ⫽ 115 MN/m3)
on 25 mm, 36 kg/m3 mineral wool (s⬘ ⫽ 11 MN/m3)
Lx ⫽ Ly ⫽ 2 m, ρs ≈ 142 kg/m2, cL ≈ 3975 m/s, fB(thin) ⫽ 3346 Hz
⫺106
⫺112
140 mm cast in situ concrete slab
(rigidly connected to masonry walls along all edges)
Lx ⫽ 4.2 m, Ly ⫽ 3.6 m, ρs ⫽ 345 kg/m2, cL ⫽ 3800 m/s,
fB(thin) ⫽ 1485 Hz
⫺118
⫺124
⫺130
50
63
80
100 125 160 200 250 315 400 500 630 800 1000 1250
One-third-octave-band centre frequency (Hz)
Figure 2.50
Spatial average, one-third-octave-band values for the real part of the driving-point mobility. Comparison of measurements with
infinite plate theory for different plates that form walls and floors.
Measured data from Hopkins are reproduced with permission from ODPM and BRE.
modes with identical mode frequencies and identical mode shapes. When different modes
have identical mode frequencies the plates are described as having degenerate modes. These
can cause pronounced peaks and troughs in the plate response when the mode spacing is
wide and damping is low. Secondly, the resilient material that is under the screed affects the
plate mobility. The dynamic stiffness of the closed-cell foam is much higher than the mineral
187
S o u n d
I n s u l a t i o n
wool; hence the structural coupling losses from the screed to the base floor are much higher
with the foam than the mineral wool. This means that the damping in terms of the total loss
factor is higher for the screed on the closed-cell foam; this higher damping reduces the modal
fluctuations. Thirdly, there are variations in the screed properties due to workmanship and
materials. Despite a nominally identical sand–cement mix for both screeds, the screed on the
mineral wool had a more variable thickness and density over its cross-section. This is due to
manual laying of screeds on resilient materials that have significantly different compression
under a static load. All of these factors cause the mobility to be significantly different in the lowfrequency range for what might be considered to be nominally identical plates. However, as
the mode count increases in the mid-frequency range, both screeds are adequately modelled
by an infinite plate.
The 22 mm chipboard floating floor is formed from individual boards. If this floor acted as a
single plate, all frequency bands between 50 and 1250 Hz would have mode counts greater
than five. With distributed damping over its surface due to the resilient material this floor it could
be expected to be modelled as an infinite plate in both the low- and mid-frequency ranges. The
only significant departure occurs around 80 Hz; this coincides with the mass–spring resonance
frequency. However, mass–spring resonance frequencies of floating floors (timber or screed)
are not usually detectable in the measured driving-point mobility (real part).
Three similar 215 mm dense aggregate walls were built under laboratory conditions by the
same builder using the same batch of blocks. The measurements indicate that it is unlikely
that modal fluctuations in the mobility can be accurately predicted for a specific plate and
subsequently validated by measurements when the uncertainty due to the spatial variation is
so large.
These examples illustrate that for many calculations it is adequate to estimate the spatial
average driving-point mobility by assuming an infinite plate. As it is rarely possible to predict the
modal fluctuations at low frequencies, this is often assumed over the entire building acoustics
frequency range. However, for individual positions it is possible to estimate the envelope of
the peaks in the mobility spectrum, Ŷdp , using (Skudrzyk, 1981)
Ŷdp =
X
ωηm
(2.192)
where X = 2 for beams and X = 4 for plates, η is the total loss factor and m is the mass of the
beam or plate.
Equation 2.192 shows that the envelope decreases with increasing loss factor and increasing
frequency; eventually tending towards the infinite plate equation. Figure 2.51 shows the envelope of peaks for the 140 mm concrete slab for two individual excitation positions. When the
local mode assumption is reasonable, these estimates work well for individual positions, but
for masonry/concrete plates they tend to overestimate the envelope for the spatial average
value (Moorhouse and Gibbs, 1995).
For concrete floors, static loads (e.g. furniture) do not significantly affect the driving-point
mobility. On lightweight floating floors, static loads of 20–25 kg/m2 typically change the drivingpoint mobility by up to ±3 dB in an individual one-third-octave-band. These fluctuations tend to
average out across the frequency range so the effect of the static load can usually be ignored
when the infinite plate model is appropriate for the unloaded floor.
188
Chapter 2
⫺70
Envelope of peaks, Yˆdp
⫺80
Position A
Position B
20 lg (Re{Ydp}) (dB)
⫺90
⫺100
Infinite
plate
⫺110
⫺120
⫺130
⫺140
⫺150
10
100
Frequency (Hz)
1000
Figure 2.51
Envelope of the peaks in the driving-point mobility at individual positions on the 140 mm cast in situ concrete slab described
in Fig. 2.47 (positions A and B).
2.8.3.2 Excitation at the edge
Structure-borne sound sources can sometimes be placed at the edge of a wall or floor although
it is more common for them to be in the central part.
For a thin plate of semi-infinite extent, the driving-point impedance for excitation of bending
waves at an edge is real and given by (Cremer et al., 1973)
(2.193)
Zdp = 3.5 Bp ρh = ρcL,p h2
For a thin beam of semi-infinite extent, the driving-point impedance for excitation of bending
waves at an edge is complex and given by (Cremer et al., 1973)
1
Zdp = (1 + i) ρScB,b = (1 + i)0.67ρS cL,b hf
2
(2.194)
where S is the cross-sectional area of the beam.
2.8.3.3
Finite beams and plates with more complex cross-sections
Many beams and plates have more complex cross-sections than a solid, homogeneous rectangle. For one-dimensional beams with solid, non-rectangular cross-sections, the moment of
inertia can usually be found in standard tables. This allows calculation of the phase velocity for
bending waves (Eq. 2.59) and use of the infinite beam equation to calculate the driving-point
impedance (Eq. 2.191). For point excitation in the central area of any two- or three-dimensional
isotropic wave-bearing structure (i.e. away from the boundaries), a general rule-of-thumb is
that the driving-point impedance equals the product of the angular frequency and the mass of
the structure contained within a sphere (centred at the excitation point) with a radius of one-third
189
S o u n d
I n s u l a t i o n
of the structural wavelength (Heckl, 1988). For orthotropic structures, it is the mass contained
within an ellipsoid (centred at the excitation point) where the semi-axes correspond to the
wavelengths in the different directions. For thin plates, estimates can be found by considering the mass contained below a circle (isotropic plate) or an ellipse (orthotropic plate) on the
surface. Note that for orthotropic non-homogeneous plates with complicated cross-sections, it
may only be possible to make a rough estimate for the bending wavelength.
For point excitation of bending waves, this approach can be used to estimate the driving-point
impedance in the central area of a plate (Heckl, 1988)
Zdp ≈ ωρs π
λB,p
3
2
(2.195)
or in the central part of a beam (Heckl, 1988)
Zdp ≈ ωρl
λB,b
3
(2.196)
Many walls and floors are formed from combinations of plates and beams. The resulting
driving-point mobility on the surface of the plate is often frequency-dependent over the building
acoustics frequency range. Despite this, reasonable estimates for many plates with complicated cross-sections can often be obtained by using the infinite plate equations. As an
example, Fig. 2.52 indicates that solid concrete stairs can be adequately represented by an
infinite isotropic plate with an average plate thickness, hav .
Some cast in situ concrete floors are profiled and can be modelled using the general form
of a ribbed plate, often with a solid trapezoid forming the rib cross-section instead of a solid
rectangle. Ribbed plates (as defined in Section 2.3.3.2.3) are often highly orthotropic. The
effective bending stiffness can be calculated from the stiffness in the two orthogonal directions
(Eq. 2.109). At low frequencies where λB,eff ≫ dR , the driving-point mobility can be calculated
using the effective bending stiffness in the infinite plate equation. It can also be assumed to
be the same at any point on the plate surface (i.e. on top of a rib, or in-between ribs). With
increasing frequency the bending wavelength decreases and when dR ≈ λB,eff /2, the plate
section between the ribs can be considered as vibrating independently of the entire plate. The
driving-point mobility will then be significantly different in the area between the ribs compared
to the area along the top of the ribs.
In-between the ribs the driving-point mobility tends towards that of an infinite plate where the
bending stiffness is determined only by the properties of the plate in-between the ribs. The
transition from the infinite orthotropic plate model (using the effective bending stiffness) to an
infinite isotropic plate representing the area in-between the ribs can be assumed to start when
dR ≈ λB,eff /4, and to be completed when dR ≈ λB,eff /2 (Gerretsen, 1986). A straight line between
one-third-octave-band centre frequencies is usually adequate for this transition range.
At excitation points along the top of the ribs, the driving-point mobility can be modelled using
the moment of inertia for a T-shape beam (Gerretsen, 1986). This is the repeating T-shape
cross-section that has a repetition distance, dR (refer back to Fig. 2.22). For the rectangular
T-shape beam the relevant moment of inertia for lateral displacement in the z-direction is given
by Eq. 2.110. This can be used to calculate the phase velocity for bending waves (Eq. 2.59)
followed by the driving-point impedance for a thin infinite beam (Eq. 2.191). This is a general
approach that can be used with other shapes for the repeating beam section. The model for
an infinite orthotropic plate can switch to the infinite beam model when dR ≈ λB,eff /4.
190
Chapter 2
⫺70
Solid concrete stairs ( ρ ⫽ 2000 kg/m3, cL ⫽ 3900 m/s)
Stair width: 1.1 m
Distance between landings: 2.5 m (9 stairs)
d1 ⫽ 0.16 m, d2 ⫽ 0.17 m, d3 ⫽ 0.225 m, d4 ⫽ 0.02 m
d4
⫺80
20 lg (<Re{Y dp }>s ) (dB)
⫺90
hav⫽d1 ⫹
d2 ( d3 ⫹ d4)
2
d2
d 22 ⫹d 23
A
d3
d1
B
⫺100
⫺110
⫺120
Measured on the stairs (A)
⫺130
Measured underneath (B)
Infinite plate
⫺140
50
80
125
200
315
500
One-third-octave-band centre frequency (Hz)
800
1250
Figure 2.52
Driving-point mobility (real part) on solid concrete stairs. Measured data are shown with 95% confidence intervals. Measured
data from Hopkins are reproduced with permission from ODPM and BRE.
When predicting the impact sound insulation of floors, a weighted average driving-point mobility
is usually needed at frequencies above dR ≈ λB,eff /4 (Gerretsen, 1986). This is because most
impact sources (e.g. footsteps, ISO tapping machine) excite the floor at random positions both
above and in-between the ribs. This weighted average can be calculated using
Re{Ydp }w =
dR − d y
dy
+
Re{Ydp }
Re{Ydp }
dR
dR
above
in-between
ribs
(2.197)
ribs
An example is shown in Fig. 2.53 for a 100 mm thick ribbed floor made of solid concrete; 100
and 200 mm thick solid concrete floors are also shown for comparison. Although it appears
that the 100 mm ribbed plate has a significantly lower driving-point mobility than the 100 mm
solid plate, this needs to be considered along with the modal peaks that will occur on different
finite size plates (Section 2.8.1).
Some floors incorporate beams within the cross-section; this usually results in an orthotropic
plate. A common example is a beam and block floor built from concrete beams and rows of
loose-laid masonry blocks. If the masonry blocks are not rigidly bonded together and there
are narrow gaps between some of them, the bending stiffness in the direction parallel to
the concrete beams will mainly be determined by the beams themselves. Perpendicular to
191
S o u n d
I n s u l a t i o n
⫺80
B
A
100 mm ribbed concrete plate
(Concrete: ρ ⫽ 2200 kg /m3, cL ⫽ 3800 m/s)
⫺90 h ⫽ 0.1 m
dz ⫽ 0.3 m
20 lg (Re{Ydp}) (dB)
⫺100
dR = 1 m
dy = 0.2 m
dR ⫽ λB,eff /4 at 136 Hz
⫺110
dR ⫽ λB,eff /2 at 544 Hz
⫺120
100 mm solid concrete plate
200 mm solid concrete plate
⫺130
100 mm ribbed concrete plate in-between ribs (A)
100 mm ribbed concrete plate on top of ribs (B)
Weighted average, <Re{Ydp}>w
⫺140
50
80
125
200
315
500
One-third-octave-band centre frequency (Hz)
800
1250
Figure 2.53
Driving-point mobility (real part) on ribbed concrete plates.
the beams, the bending stiffness is usually lower, but highly variable because the plates are
non-homogeneous. For floors built with solid concrete beams and solid masonry blocks the
bending stiffness can be 20–50% higher in the direction parallel to the beams compared to
perpendicular to the beams (Hopkins, 2004). Figure 2.54 shows measurements on four similar
beam and block floors with different solid blocks. Bending modes on individual solid blocks only
occur in the high-frequency range. In the low- and mid-frequency ranges the real part of the
driving-point mobility on the blocks can be highly variable and simple models cannot predict
the differences between different blocks. In contrast, the beams have a low spatial variation for
the driving-point mobility and the infinite thin beam model is reasonable over most of the lowand mid-frequency range. A thin surface finish that bonds the beams and the blocks together
may allow the floor to be modelled as an infinite plate for the driving-point mobility either on or
in-between the beams. The example shown in Fig. 2.54 indicates that the driving-point mobility
on a floor with a thin surface finish is significantly lower than on the beams or the blocks of
the same floor when fair-faced. Much thicker surface finishes, such as a bonded screed which
is typically >30 mm, do not always cause the floor to behave in the same way. For thicker
surface finishes, the driving-point mobility can be estimated using Eq. 2.195 but it is awkward
to determine the bending wavelength without measurements. If the blocks have a very low
mass per unit area (e.g. hollow clay pots or polystyrene) compared to the bonded screed then
it is sometimes possible to ignore them and use the ribbed plate model for the beams and
the screed (Gerretsen, 1986). There are a wide variety of beam and block floors of varying
complexity. Measurements of the driving-point mobility are often required to identify simple
models or to develop empirical equations for a particular type of floor.
192
Chapter 2
Four beam and block floors: Lx ≈ Ly ≈ 4.5 m, h ≈ 0.15 m, ρs ≈ 310 kg/m2
Concrete beams: ρl ≈ 31–51 kg/m, cL ≈ 4000 m/s,
width ≈ 80–150 mm, height ≈ 150 mm
Solid masonry blocks: h ≈ 0.1–0.15 m
(1) d1 ⫽ 480 mm, d2 ⫽ 240 mm, m ⫽ 35 kg, (2) d1 ⫽ 440 mm, d2 ⫽ 140 mm, m ⫽ 18 kg,
(3) d1 ⫽ 215 mm, d2 ⫽ 215 mm, m ⫽13 kg, (4) d1 ⫽ 440 mm, d2 ⫽ 190 mm, m ⫽ 25 kg
(a) On blocks (measured)
⫺82
⫺88
⫺94
⫺100
⫺106
Floor No.1
Floor No.2
Floor No.3
Floor No.4
20 lg (<Re{Y dp }>s ) (dB)
⫺112
⫺118
⫺124
B
A
d2
d1
⫺82
(b) On beams (measured and predicted)
⫺88
⫺94
⫺100
⫺106
(Black symbols are predicted values for a thin infinite beam)
⫺88
Floor No.2 (cL,eff ⫽ 2880 m/s) with a surface finish of
5 mm levelling compound/screed (measured and predicted)
⫺94
⫺100
⫺106
⫺112
⫺118
Measurements: (A) above blocks
(B) above beams
(Black symbols are predicted values for a thin infinite plate)
⫺124
⫺130
50
63
80
100 125 160 200 250 315 400 500 630
One-third-octave-band centre frequency (Hz)
800 1000 1250
Figure 2.54
Driving-point mobility (real part) on beam and block floors. Measured data are shown with 95% confidence intervals.
Measured data from Hopkins are reproduced with permission from ODPM and BRE.
193
S o u n d
I n s u l a t i o n
⫺20
Floating floor (Lx ≈ Ly ≈ 3 m):
18 mm tongue & grooved chipboard ( ρs⫽10.6 kg/m2, cL⫽2200 m/s). This walking surface is
bonded to19 mm plasterboard with ≈5 mm thick circular dabs of gypsum-based adhesive
(≈150 mm diameter at 300 mm centres) leaving areas ≈150 mm wide without adhesive
Resilient material: 30 mm mineral wool (60 kg/m3)
20 lg (<Re{Ydp}>s ) (dB)
⫺30
⫺40
Measured (spatial average on the chipboard above and in-between dabs)
⫺50
Infinite plate: 18 mm chipboard
Infinite plate: 42 mm plate with average properties of 18 mm chipboard
plus 5 mm adhesive plus 19 mm plasterboard
⫺60
Chipboard
λB/2⫽0.15 m at 798 Hz
⫺70
⫺80
50
80
125
200
315
500
One-third-octave-band centre frequency (Hz)
800
1250
Figure 2.55
Driving-point mobility (real part) on a timber platform floating floor comprised of two boards bonded together. Measured
data are shown with 95% confidence intervals. Measured data from Hopkins are reproduced with permission from ODPM
and BRE.
Lightweight walls and floors are often constructed with two or three layers of board (e.g. plasterboard, chipboard) that are fixed or bonded together using screws, nails, or dabs of adhesive.
For these closely connected plates there is only usually a thin air gap between the layers. When
point excitation is applied to one of the boards it begins to support modal vibration between the
fixing points at and above the frequency where there is half a bending wavelength between the
connections. As the frequency increases this board effectively acts independently of the other
board(s). This allows the driving-point mobility to be modelled by assuming an infinite plate for
the single layer of board that is being excited. An example is shown in Fig. 2.55 for two boards
connected with relatively large areas of adhesive leaving only small unconnected areas. Above
the frequency at which there is half a bending wavelength over each of these small areas, the
driving-point mobility is adequately predicted by using the infinite plate model. In fact as the infinite plate prediction for the two boards acting as a single plate (42 mm thick) is clearly inappropriate; the infinite plate assumption for the single board can be used to give a reasonable estimate
across the entire low- and mid-frequency range. Any differences need to be considered alongside the fact that apart from floors built in the laboratory or under factory conditions there will be
so much variation due to workmanship that the infinite plate assumption will be quite adequate.
Lightweight walls and floors are usually built from boards that are nailed or screwed to a framework of beams. We previously looked at the driving-point mobility on ribbed plates formed
194
Chapter 2
(a) Spatial average values
⫺20
Floating floor (Lx ≈ Ly ≈ 3 m, cL,eff ⫽ 2830 m/s):
18 mm tongue & grooved chipboard ( ρs ⫽ 10.6 kg/m2, cL ⫽ 2410 m/s) screwed to
45 mm ⫻ 45 mm timber battens ( ρl ⫽ 0.9 kg/m, cL ⫽ 5490 m/s)
Battens spacing: 450 mm centres. Average screw spacing: 300 mm centres
⫺30
20 lg ( <Re{Ydp}>s) (dB)
Resilient material: 30 mm mineral wool (60 kg/m3)
⫺40
Measured: (A) spatial average on the chipboard in-between the beams
Measured: (B) spatial average on the chipboard above the beams
Infinite plate: 18 mm chipboard
⫺50
⫺60
Chipboard
λB,eff /2 ⫽ 0.45 m at 114 Hz
⫺70
B
A
λB /2 ⫽ 0.45 m at 97 Hz
λB ⫽ 0.45 m at 389 Hz
⫺80
50
80
125
200
315
500
One-third-octave-band centre frequency (Hz)
800
1250
Figure 2.56
Driving-point mobility (real part) on a timber raft floating floor. Measured data are shown with 95% confidence intervals.
Measured data from Hopkins are reproduced with permission from ODPM and BRE.
from a single material, such as concrete. This model is less useful for lightweight walls and
floors when the beams and plates have significantly different material properties and there
are only point connections between them. In addition, although some frames (e.g. timber)
can significantly change the bending stiffness of the board, others (e.g. light steel) have a
lesser effect. To illustrate some of the main features, measurements on a floating floor comprising single boards screwed to solid timber battens are shown in Fig. 2.56. The battens
run in a single direction giving a bending stiffness for the floor that is ≈350% higher in the
direction parallel to the battens than perpendicular to them. Below the frequency at which
there is approximately λB,eff /2 between the battens, the spatial average mobility above or
in-between the battens is approximately the same. The relatively uniform response above and
in-between the battens can also be seen on the contour plots (region A). However, at these low
frequencies the floating floor does not simply act as an infinite orthotropic plate. The measured
mobility in region A is affected by the constraint to motion and the damping that is provided by
the resilient material and the base floor. These no longer have a significant effect when there is
approximately λB /2 between the battens and modes occur on the sub-panels that are formed
by the chipboard in-between the battens (region B). In region B there are several modes with
half a bending wavelength between the battens; the chipboard above the battens then lies
along nodal lines and the mobility is lower than in-between the battens. At frequencies where
195
S o u n d
I n s u l a t i o n
(b) Spatial variation along two measurement lines between battens
(line 1 is approximately mid-way between screws and line 2 is close to the screws)
20 lg (Re{Ydp})
(dB)
Measurement
line 1
Measurement
line 2
1225
1225
1165
1165
1105
1105
1045
1045
985
985
925
925
865
865
⫺51
⫺53
⫺55
Measurements taken at nine positions equally
spaced between mid-points of the battens
⫺57
⫺59
⫺61
805
805
745
745
685
685
625
625
565
565
505
505
⫺65
λB ⫽ 0.45 m
at 389 Hz
⫺63
445
⫺67
385
265
265
140 mm
Plan
10 mm
Measurement line 2
10 mm
300 mm
indicates screw position
205
205
145
Region:
A
85
25
100 mm
290 mm
325
⫺73
⫺75
160 mm
Measurement line 1
385
325
λB/2 ⫽ 0.45 m
at 97 Hz
⫺71
150 mm
445
B
⫺69
Frequency (Hz)
C
Section
145
85
25
Figure 2.56
(Continued )
there is at least one bending wavelength between the battens, the distinction between the
mobility above and in-between the battens starts to break down (region C). The infinite plate
model can then be used to estimate the driving-point mobility. For design purposes, different
screw spacings, different board sizes, different connections between boards (straight edge or
196
Chapter 2
tongue and grooved) as well as variation in the material properties imply that there is little to
be gained from a single deterministic model.
2.9 Sound radiation from bending waves on plates
Sound radiation concerns the coupling of structure-borne sound waves to sound waves in an
adjacent fluid. Our main interest is in bending waves radiating into air. Radiation is usually
described using the radiation efficiency, σ. This is defined as the ratio of the radiated power to
the power radiated by a large baffled piston (ka ≫ 1 where a is the piston radius) with a uniform
mean-square velocity equal to the temporal and spatial average mean-square velocity of the
plate,
σ=
W
Sρ0 c0 v 2 t,s
(2.198)
where W is the radiated sound power, and S is the surface area of the plate.
For ka ≫ 1, the piston dimensions are larger than a wavelength in air, so the air particles cannot
escape compression by moving sideways. The particle velocity, u, therefore equals the velocity
of the piston surface, vpiston . The power radiated by a piston is determined using Eq. 1.18 to
give the sound pressure,
p = ρ0 c0 vpiston
(2.199)
which is substituted in Eq. 1.19 to give the sound power radiated by the piston,
2
W = SI = Sρ0 c0 vpiston
t
(2.200)
When the radiation efficiency is unity, a plate radiates the same power as a baffled piston
(assuming the same ρ0 c0 ) which has the same area and mean-square velocity. We will shortly
see that plates can be much less or much more efficient in radiating sound than this baffled
piston.
This section starts by looking at sound radiation from bending waves on an infinite plate. This
provides a general insight into sound radiation and sets the scene for sound radiation from
individual bending modes on finite reverberant plates. We then look at the frequency-average
radiation efficiency as it is usually necessary to adopt a statistical approach to quantify sound
radiation from large numbers of bending modes.
2.9.1
Critical frequency
Sound radiation involves two different types of wave with different phase velocities. Bending
waves are dispersive, so for isotropic plates there will be a single frequency at which cB,p = c0 .
Example phase velocities for various plates are compared with c0 in Fig. 2.57. The frequency
at which cB,p = c0 is called the critical frequency, fc , and can be calculated from Eq. 2.66, giving
c2
fc = 0
2π
c2
ρs
= 0
Bp
π
√
c02 3
3ρs (1 − ν2 )
=
Eh3
πhcL
(2.201)
197
S o u n d
I n s u l a t i o n
1600
215 mm dense aggregate (94 Hz)
1400
140 mm concrete (122 Hz)
100 mm aircrete (341 Hz)
Phase velocity (m/s)
1200
12.5 mm plasterboard (3483 Hz)
1000
800
600
Air, c0 ⫽ 343 m/s
400
200
0
50
80
125
200
315
500
800
Frequency (Hz)
1250
2000
3150
5000
Figure 2.57
Comparison of the phase velocity for longitudinal waves in air with bending waves (thin or thick plate) on various plates. The
critical frequency is given in brackets for each plate in the legend. Note that only the 12.5 mm plasterboard acts as a thin plate
over the entire building acoustics frequency range, the others act as thin or thick plates depending upon the frequency.
For orthotropic plates, there are two critical frequencies, fc,x and fc,y , in the x- and ydirections respectively. To simplify some calculations an effective critical frequency, fc,eff , can
be calculated from the effective bending stiffness (Eq. 2.97), this is equivalent to
fc,eff = fc,x fc,y
(2.202)
Note that the critical frequency depends on the phase velocities of sound in air and bending
waves on the plate; hence it is dependent upon temperature. All examples relating to sound
radiation will assume c0 = 343 m/s and ρ0 = 1.21 kg/m3 unless stated otherwise.
2.9.2
Infinite plate theory
Before looking at finite plates, it is useful to consider sound radiation from a plane bending
wave propagating on an infinite plate without damping.
At the critical frequency, equal phase velocities corresponds to λB = λ as illustrated in Fig. 2.58.
The critical frequency is sometimes referred to as the lowest coincidence frequency. Coincidence, or trace-matching, occurs when the projection of the wavelength in air onto the surface
of the plate equals the bending wavelength (Cremer, 1942). The critical frequency describes
the lowest frequency at which coincidence occurs (i.e. at grazing incidence where the angle of
incidence θ = 90◦ ). At any frequency above the critical frequency there will always be an angle,
θ, that satisfies the coincidence condition (see Fig. 2.59) for which the following relationship
applies,
sin θ =
198
λ
λB
(2.203)
Chapter 2
λ
λB
Longitudinal wave in air
Solid homogeneous infinite plate
undergoing bending wave motion
Figure 2.58
Bending wave on an infinite plate and the radiated longitudinal wave in air at the critical frequency.
We will assume that the infinite plate lies in the xy plane and that the plate radiates into air in
the positive z-direction (see Fig. 2.60). The negative z-direction is assumed to be in vacuo. To
simplify matters, the problem can be restricted to one-dimension of the plate by only considering
a bending wave propagating in the positive x-direction. This bending wave has a lateral velocity
described by
v(x, t) = v̂ exp(−ikB x) exp(iωt)
(2.204)
The radiated sound pressure must have the same dependence on x as the bending wave (i.e.
kx = kB ), and propagate in the positive z-direction, hence
p(x, z, t) = p̂ exp(−ikB x) exp(−ikz z) exp(iωt)
(2.205)
which must satisfy the wave equation for longitudinal waves in air, so inserting Eq. 2.205 into
Eq. 1.14 gives
k 2 = kB2 + kz2
(2.206)
kz = ± k 2 − kB2
(2.207)
Equation 2.206 then gives kz as
At z = 0 the lateral plate velocity must equal the component of the particle velocity that is
perpendicular to the plate surface (i.e. the z-component), so
v(x, t) = uz (x, t)
at z = 0
(2.208)
199
S o u n d
I n s u l a t i o n
θ
λ
λB
θ
θ
λB
λ
Air
Solid homogeneous infinite plate
undergoing bending wave motion
Figure 2.59
Plane wave in air radiated by a bending wave on an infinite plate above the critical frequency.
The particle velocity is taken from Eq. 1.17, which gives
v̂ =
kz
p̂
ωρ0
(2.209)
Substituting Eqs 2.207 and 2.209 into Eq. 2.205 yields the radiated sound pressure,
ωρ0 v̂
exp(−ikB x) exp[−i(± k 2 − kB2 )z] exp(iωt)
p(x, z, t) =
(2.210)
± k 2 − kB2
where the choice of positive or negative sign depends on the ratio of k to kB and the physics
of the situation as discussed below.
Above the critical frequency, where k > kB (as well as cB,p > c0 and λB > λ), the value of kz is
real. So for sound to propagate away from the plate in the positive z-direction, the positive sign
200
Chapter 2
Radiated wave
θ
Vacuum
Air
x
z
Figure 2.60
Plane wave radiated by an infinite plate.
must be chosen in Eq. 2.210. As kz is real, the pressure at the plate surface is in phase with
the plate velocity. From our initial discussion on sound radiation above the critical frequency,
we already have a relationship between the wavelength in air and the bending wavelength for
the coincidence condition (Eq. 2.203). This can be rewritten in terms of the wavenumbers,
sin θ =
kB
k
(2.211)
Equation 2.210 can now be rewritten in terms of the propagation angle, θ, for the radiated
sound wave,
p(x, z, t) =
ρ0 c0 v̂
exp(−ikx sin θ) exp(−ikz cos θ) exp(iωt)
cos θ
(2.212)
To determine the radiation efficiency, the plate is temporarily assumed to be very large, rather
than infinite, so that S has a finite value and the power radiated from one side of the plate is
given by
1
Sρ0 c0 v̂ 2
W =
Re{pv ∗ }dS =
(2.213)
2
2 cos θ
S
The temporal and spatial average mean-square velocity for the plate is
v 2 t,s =
v̂ 2
2
(2.214)
201
S o u n d
I n s u l a t i o n
Therefore from Eq. 2.198, the radiation efficiency above the critical frequency is
σ=
1
k
=
cos θ
k 2 − kB2
(2.215)
At the critical frequency, k = kB (as well as cB,p = c0 and λB = λ). The sound wave radiated by
the infinite plate must therefore propagate parallel to the plate surface (θ = 90◦ ). Hence from
Eq. 2.215, the radiation efficiency at the critical frequency is infinite.
Below the critical frequency, where k < kB (as well as cB,p < c0 and λB < λ), the value of kz is
imaginary. Equation 2.210 must result in an exponentially decaying nearfield in the positive
z-direction, hence the negative sign needs to be chosen. The sound field represents a surface
wavefield (Cremer et al., 1973) where the sound pressure is given by
iωρ0 v̂
p(x, z, t) =
exp(−ikB x) exp(−z kB2 − k 2 ) exp(iωt)
kB2 − k 2
(2.216)
As kz is imaginary, the pressure at the plate surface is 90◦ out of phase with the plate velocity;
therefore the radiated power and the radiation efficiency are both zero. The nearfield decays
away rapidly with distance from the plate surface. The two components of the particle velocity
in this nearfield are given by Eqs 1.17 and 2.216,
ikB v̂
ux (x, z, t) =
exp(−ikB x) exp(−z kB2 − k 2 ) exp(iωt)
kB2 − k 2
uz (x, z, t) = v̂ exp(−ikB x) exp(−z kB2 − k 2 ) exp(iωt)
(2.217)
(2.218)
In this nearfield the particle velocity follows closed elliptical paths (Cremer et al., 1973). The
sound field is reactive as there is no net energy transport away from the plate surface. Conceptually it is simpler to imagine that the particles move parallel to the plate surface in a back
and forth motion. In this way the particles constantly manage to avoid any compression due to
the plate motion, and without compression, there can be no radiation of sound.
The above analysis of an infinite plate shows that we can consider sound radiation in three
distinct regions: below, at, and above the critical frequency. To illustrate this, the radiation
efficiency for an infinite plate is shown in Fig. 2.61. We will find that these three regions are
also important for finite plates, although the values for the radiation efficiency will differ. The
feature that finite and infinite plates share in common is that above the critical frequency, the
radiation efficiency tends towards unity. Below the critical frequency the radiation efficiency for
finite plates will often be low, but not zero; at the critical frequency it can be higher than unity,
but not infinite.
2.9.3
Finite plate theory: Radiation from individual bending modes
Sound radiation from individual bending modes can be calculated for a homogeneous, isotropic,
rectangular plate with simply supported boundaries. This plate lies within the plane of an
infinite rigid baffle to avoid interaction between the sound fields on opposite sides of the plate
202
Chapter 2
10
9
Radiation efficiency, σ(⫺)
8
7
6
5
4
3
2
1
0
0.1
1
k/kB ⫽ λB/λ ⫽
10
f
fc (⫺)
Figure 2.61
Radiation efficiency for an infinite plate.
z
r
θ
φ
Infinite
baffle
x
Ly
Lx
y
Figure 2.62
Coordinate system for sound radiation from a plate in an infinite baffle to a receiver point in the farfield.
(see Fig. 2.62). At any angular frequency, ω, the sound pressure at a point in the farfield due
to radiation from plate mode fp,q is (Wallace, 1972)
iωρ0
pp,q (r , θ, φ, t) =
exp(−ikr ) exp(iωt)
2πr
Ly Lx
0
0
αx
βy
v̂p,q ψp,q exp i
+i
dx dy
Lx
Ly
(2.219)
203
S o u n d
I n s u l a t i o n
where α = kLx sin θ cos φ, β = kLy sin θ sin φ, v̂p,q is a constant relating to the velocity in
the z-direction, and ψp,q is the local mode shape given by Eq. 2.156. The farfield condition
applies when (r − x sin θ cos φ − z sin θ sin φ) ≫ Lx and Ly (Fahy, 1985).
The sound intensity from mode fp,q in the farfield is given by Eqs 1.19 and 2.219 as
⎫2
⎧
⎪
⎪
β
α
⎪
⎪
⎪
⎪
⎪
2 ⎪
Ŵ
⎬
⎨
v̂p,q kLx Ly
2
2
&
'&
'
Ip,q (r , θ, φ) = 2ρ0 c0
2
2
⎪
⎪
r π3 pq
⎪
⎪
α
β
⎪
⎪
⎪
−1
−1 ⎪
⎭
⎩
pπ
qπ
(2.220)
where Ŵ is cos when p is an odd integer and sin when p is an even integer, is cos when q
is an odd integer and sin when q is an even integer.
The power radiated from one side of the plate is calculated from the surface integral over a
hemisphere encompassing the plate,
Wp,q
2π π/2
Ip,q (r , θ, φ)r 2 sin θ dθ dφ
=
0
(2.221)
0
To calculate the radiation efficiency we also need the temporal and spatial average meansquare velocity, which for mode, fp,q , on a plate with simply supported boundaries is
2
t,s
vp,q
1
=
Lx Ly
Ly Lx
0
2
v̂p,q
(v̂p,q Ψp,q )2
dx dy =
2
8
(2.222)
0
Hence, the radiation efficiency, σp,q , for mode fp,q is (Eq. 2.198) is
σp,q =
64k 2 Lx Ly
π 6 p2 q 2
⎫2
⎪
β
α
⎪
⎪
⎪
Ŵ
⎬
2
2
&
' &
' sin θ dθ dφ
⎪
⎪
⎪
⎪
α 2
β 2
⎪
⎪
⎪
−1
−1 ⎪
⎭
⎩
pπ
qπ
⎧
⎪
⎪
⎨
⎪
π/2π/2⎪
0
0
(2.223)
It is important to note that the radiation efficiency of a mode is a function of frequency. So far
we have only been considering modes at their specific resonance frequencies, such as when
counting the number of resonant modes in a frequency band, or looking at mode shapes. For
sound radiation it is important to note that whilst a mode has its own resonance frequency, we
can calculate the sound radiated by that mode at any frequency.
Equation 2.223 can be evaluated numerically, or the following approximations can be used
when kLx ≪ 1 and kLy ≪ 1 (Wallace, 1972)
σp,q ≈
+
*
Ly
k 2 Lx Ly
32k 2 Lx Ly
Lx
8
8
1
−
+
1
−
1
−
2
π 5 p2 q 2
12
( pπ)2 Ly
Lx
(qπ)
(2.224)
when p and q are both odd integers.
σp,q ≈
204
2k 6 L3x L3y
15π5 p2 q2
*
1−
5k 2 Lx Ly
64
1−
24
( pπ)2
+
Ly
Lx
24
+ 1−
2
Ly
Lx
(qπ)
(2.225)
Chapter 2
10
Radiation efficiency, σp,q (⫺)
1
0.1
0.01
0.001
+++++++
+++++++++
++
+++++++++++++++++++++++++++++++
++
+
+
+
+
+
++
++
Mode
+
++
++
+
f11
+
++
+
+
f21
++
++
+
++
f31
++++++
++
++
+
+
f22
+
+
++
++
+
+
f32
++
++
++
++
+
+
f54
++
++
0.0001
Lx ⫽ 4 m
Ly ⫽ 2.5 m
0.00001
0.1
1
k/kB ⫽ λB/λ ⫽
f
fc (⫺)
10
Figure 2.63
Radiation efficiency of individual modes.
when p and q are both even integers.
σp,q ≈
8k 4 Lx L3y
+
*
Ly
k 2 Lx Ly
24
8
Lx
1
−
+
1
−
1
−
2
2
5
2
2
3π p q
20
Ly
Lx
(pπ)
(qπ)
(2.226)
when p is an odd integer and q is an even integer (for the opposite combination, swap over p
and q, and, Lx and Ly ).
Example radiation efficiencies are shown in Fig. 2.63 for a number of modes on a rectangular
plate. These have been calculated using numerical integration of Eq. 2.223. They are plotted
as a function of k/kB , hence k/kB = 1 corresponds to radiation at the critical frequency. The
range 0.1 ≤ k/kB ≤ 10 covers the building acoustics frequency range for the majority of plates
typically used in buildings. Below the critical frequency, the fundamental mode, f11 , has the
highest radiation efficiency; although in the vicinity of the critical frequency, other modes have
slightly higher values. Just above the critical frequency, the radiation efficiency for all modes
reaches a peak. At this peak the radiation efficiency is slightly greater than unity. As the frequency increases further, the radiation efficiency asymptotes towards unity. There is significant
variation below the critical frequency. In this range, the general tendency is for the fundamental
mode to have the highest radiation efficiency, and modes with p and q as odd integers to have
high radiation efficiencies when compared against modes with p and q as even integers.
Below the critical frequency, the variation in radiation efficiency between individual modes is
due to the sinusoidal nature of the mode shapes. A physical interpretation can be found by
treating the plate as an array of point sources (Maidanik, 1962). Each point source is positioned
at the centre of an area defined by the nodal lines of the mode. The adjacent point sources are
π radians out-of-phase with each other. Figure 2.64 shows an example using the mode f54 .
205
S o u n d
I n s u l a t i o n
(a) Vibration field of the f54 bending mode on a rectangular plate in the plane of an infinite baffle
with simply supported boundaries.
Lx
Ly
(b) Representation of the f54 mode assuming that the areas between the nodal lines act as point sources
(rectangular piston radiators). The opposite phase of the point sources is indicated using black and white.
Figure 2.64
Radiation from bending modes on a rectangular plate.
We have already used the sound pressure in the farfield to quantify the radiated sound power
over the surface of a hemisphere that encloses the plate. If we look at the path length between
a position on this surface and two adjacent point sources of opposite phase on the plate we
will find positions where the phase difference due to the different path lengths is negligible
compared to the phase difference of π radians between the sources. At such positions there
will be almost complete cancellation of the sound pressure from the two point sources. The
path lengths depend upon the distances between adjacent point sources (Lx /p and Ly /q).
So depending upon frequency, mode numbers and plate dimensions there will be varying
degrees of cancellation over the surface of the hemisphere. To consider all the point sources
that represent the mode we now need to re-define the array of point sources. Hence we take
206
Chapter 2
(c) Corner radiator below the critical frequency.
Dashed lines within the transparent grey area indicate the areas containing dipole and quadrupole sources.
Ly /2q
Lx/2p
dipole
dipole
Ly /2q
Lx/2p
quadrupole
λB,y
λ
λB, x
λ
(d) Edge radiator below the critical frequency.
λ
λB, y
λB, x
λ
Figure 2.64
(Continued)
each area that is demarcated by the nodal lines and split it into four zones of equal area
that all have the same phase; each of these represents a point source. When the distance
between adjacent point sources is much smaller than a wavelength, the point sources are
considered in combination: a dipole is formed from two point sources (one of each phase),
207
S o u n d
I n s u l a t i o n
and a quadrupole is formed from four point sources (two of each phase). Below the critical
frequency, these dipoles and quadrupoles radiate significantly less power than a point source
(Morse and Ingard, 1968; Vér and Holmer, 1988).
Below the critical frequency, the wavelength in air is larger than the bending wavelength along
one or both of the plate dimensions. We start by considering the frequency at which λ > λB,x
and λ > λB,y for which the point sources combine to form dipoles or quadrupoles as shown in
Fig. 2.64c. The quadrupoles lie in the central region of the plate, with dipoles along the plate
edges. This leaves point sources in the corners that are responsible for most of the radiated
sound power; however, this will be affected by any interaction between the corner sources
when the plate dimensions are less than a wavelength. For the same mode we can look at
the frequency where λ < λB,x and λ > λB,y (see Fig. 2.64d). Now there are only two strips left
along the edges in the x-dimension that do not form dipoles or quadrupoles. As λ < λB,x there
are no dipoles formed along these edges, and these strips are responsible for most of the
radiated sound power. Below the critical frequency, plate modes can be classified as corner
radiators or edge radiators; the former usually radiate less power than the latter. This classification is frequency-dependent. The relatively weak radiation from corner modes and edge
modes results in radiation efficiencies that are less than unity, but not zero as with an infinite
plate.
As the fundamental mode is represented by a single point source, there are no interactions that
form dipoles or quadrupoles. It therefore tends to have the highest radiation efficiency below
the critical frequency. Compared to other modes it acts more like a simple piston. However,
when the plate dimensions are less than a wavelength in air, the air particles can partly escape
compression by moving sideways. This results in radiation efficiencies lower than unity below
the critical frequency. For modes with p and q as odd integers acting as corner radiators, the
corner sources are in phase. For this reason they tend to have higher radiation efficiencies
than modes with p and q as even integers for which there is a degree of cancellation due to
the opposite phase of the corner sources.
Below the critical frequency, radiation efficiencies for individual modes vary significantly with
different boundary conditions (Gomperts, 1977). When predicting the radiation efficiency for
plates in buildings, there is always uncertainty in describing the actual boundary conditions.
In practice it is often reasonable to assume that connected wall or floor boundaries are simply
supported. However we also need to consider the fact that many walls are built from components such as bricks and blocks; this introduces variation in the material properties over the
plate. Combined with variation in the boundary conditions, this results in mode shapes that
do not have a distinct, precise, sinusoidal pattern, and we cannot exactly predict the mode
frequencies. In addition, the mode shapes of flanking walls with doors or windows may be
considerably different to a homogeneous rectangular plate with simply supported boundaries
(refer back to Fig. 2.33). The classification of corner and edge radiators therefore becomes
rather blurred in practice. For frequency bands that are influenced by a large number of
modes, this is less of an issue because we can use a frequency-average radiation efficiency.
However, when there are low mode counts in bands below the critical frequency there may
be significant errors in using either the individual mode or the frequency-average radiation
efficiency.
Above the critical frequency, the entire surface of the plate radiates sound; at these frequencies
we can classify the mode as a surface radiator. In Section 4.3.1.5 we will see that this has important implications for modelling the sound insulation of finite plates using an infinite plate model.
208
Chapter 2
2.9.4
Finite plate theory: Frequency-average radiation efficiency
To determine a frequency-average radiation efficiency, it is assumed that the plate is homogeneous and rectangular with sinusoidal mode shapes. It is also assumed that the modal density
is sufficiently high that the radiation efficiency can be considered as a continuous function of
frequency. This leads to individual expressions for the radiation efficiency below, at, and above
the critical frequency. The radiation efficiency is usually required in one-third-octave or octavebands; for the following equations, the frequency, f , can be taken as the band centre frequency.
Below the critical frequency (Leppington, 1996; Leppington et al., 1982, 1984, 1987),
σ=
%
$
U
μ+1
2μ
ln
CBC COB − μ−8 CBC COB − 1
+ 2
2
μ
−
1
μ
−
1
2πμkS μ − 1
for f < fc
(2.227)
where U is the plate perimeter, S is the plate area, CBC is a constant for the plate boundary
conditions (CBC = 1 for simply supported boundaries, CBC = 2 for clamped boundaries), COB is
a constant for the orientation of the baffle that surrounds the edges of the plate (COB = 1 when
the plate lies within the plane of an infinite rigid baffle, COB = 2 when the rigid baffles along the
plate perimeter are perpendicular to the plate surface), and
μ=
fc
f
(2.228)
Above the critical frequency (Leppington et al., 1982; Maidanik, 1962),
1
σ=
1 − μ2
for f > fc
(2.229)
At the critical frequency (Leppington et al., 1982),
0.15L1
σ ≈ 0.5 −
L2
√
k L1
for f = fc
(2.230)
where L1 is the smaller and L2 is the larger of the rectangular plate dimensions, Lx and Ly (for
square plates, L1 = L2 = Lx = Ly ). The radiation efficiency for the frequency band that contains
the critical frequency can be calculated from Eq. 2.230 using k = 2πfc /c0 .
For orthotropic plates modelled as isotropic plates using the effective bending stiffness, the
effective critical frequency is used.
2.9.4.1 Method No. 1
As we will now introduce alternative calculations for the radiation efficiency, we will refer to the
calculations using Eqs 2.227, 2.229, and 2.230 as method no. 1.
Examples are shown in Fig. 2.65a; the infinite baffle chosen for this example would be
appropriate for plates used in the middle of a large wall that are flush with the adjacent wall
surface.
209
S o u n d
I n s u l a t i o n
(a) Plates set in the plane of an infinite baffle (method No.1)
10
6 mm glass (1.5 ⫻ 1.25 m), fc = 2079 Hz
Radiation efficiency, σ (⫺)
12.5 mm plasterboard (2.4 ⫻ 1.2 m), fc = 3483 Hz
1
0.1
0.01
0.001
50
80
125
200
315
500
800
1250
2000
3150
5000
3150
5000
One-third-octave-band centre frequency (Hz)
(b) Plate set in the plane of an infinite baffle (method No.2)
10
Radiation efficiency, σ (⫺)
6 mm glass (1.5 ⫻ 1.25 m), fc ⫽ 2079 Hz
1
0.1
0.01
0.001
50
80
125
200
315
500
800
1250 2000
One-third-octave-band centre frequency (Hz)
Figure 2.65
Radiation efficiencies for finite plates with simply supported boundaries.
2.9.4.2 Method No. 2
Accurate prediction of the radiation efficiency at and near the critical frequency is difficult to
achieve. For thin plates such as glass, plasterboard, and metal, it is quite common for the calculated radiation efficiency to overestimate the actual value in the vicinity of the critical frequency.
At and near the critical frequency, the radiation efficiency calculated using Eqs 2.227, 2.229,
and 2.230 can be considered as an upper limit. A corresponding lower limit can be determined
by using the calculated values and setting σ = 1 in the lowest frequency band for which σ > 1,
210
Chapter 2
(c) Masonry/concrete plate with baffles perpendicular to all plate edges (method No.3)
Radiation efficiency, σ (⫺)
10
1
0.1
0.01
140 mm concrete floor (4 ⫻ 3.5 m), fc ⫽ 122 Hz
100 mm aircrete wall (4 ⫻ 2.5 m), fc ⫽ 341 Hz
0.001
50
80
125
200
315
500
800
1250 2000
One-third-octave-band centre frequency (Hz)
3150
5000
(d) Masonry/concrete plate set in the plane of an infinite baffle. Comparison of method
No. 3 with No. 4.
10
Radiation efficiency, σ (⫺)
100 mm dense aggregate masonry wall (3.53 ⫻ 2.63 m)
f11 ⫽ 33 Hz, fc ⫽ 203 Hz
1
0.1
Method No.3
0.01
Method No.4
Individual modes
below fc with σp,q ≤ 1
0.001
10
100
1000
10 000
Frequency (Hz)
Figure 2.65
(Continued)
and then setting σ = 1 in all the higher frequency bands. In most cases the radiation efficiency
will lie between the lower and upper limits; for the example in Fig. 2.65b, this is shown by the
shaded area.
For plates commonly used in lightweight walls and floors (e.g. plasterboard, chipboard) that are
rigidly connected to a timber frame, the lower limit can be a better estimate than the upper limit.
211
S o u n d
I n s u l a t i o n
For laminated plates the radiation at critical frequency can be highly dependent on the number
and type of ply groups (Matsikoudi-Iliopoulou and Trochidis, 1992). For some laminates such
as plywood the radiation efficiency may not peak above unity at the critical frequency and the
lower limit can give a better estimate than the upper limit.
2.9.4.3 Method No. 3 (masonry/concrete plates)
For masonry/concrete walls and floors, there are some important issues relating to methods
no. 1 and no. 2. Firstly, these plates usually have a low modal density which goes against the
assumption of high modal density. The second is that whilst common thicknesses of glass,
metal, and plasterboard have a distinct peak in the radiation efficiency near the critical frequency, this is not usually the case with masonry/concrete plates that are rigidly connected to
other plates at the boundaries. An exception is when the plate boundaries are free, or isolated
from the surrounding structure by a resilient material and the coupling losses from the plate are
negligible. The third is that Eqs 2.227, 2.229, and 2.230 sometimes give irregular, rather than
smooth curves near the critical frequency (e.g. double peaks). For masonry/concrete plates
(which typically have fc < 500 Hz), semi-empirical adjustments that will be now be described
by methods no. 3 and no. 4 usually give reasonable estimates.
Method no. 3 requires the full calculation using Eqs 2.227, 2.229, and 2.230, then setting σ = 1
in the lowest frequency band for which σ > 1, and then setting σ = 1 in all higher frequency
bands. Examples are shown in Fig. 2.65c; these particular examples represent complete wall
or floor surfaces in a box-shaped room, so the chosen orientation for the baffle is perpendicular
to the plate surface along the perimeter of the plate.
2.9.4.4 Method No. 4 (masonry/concrete plates)
For homogeneous masonry/concrete plates with Ns < 3 in frequency bands below the critical
frequency, the radiation efficiency can be significantly underestimated by Eq. 2.227. Better
estimates than method no. 3 can be found in situations where the plate can be modelled as
being within the plane of an infinite rigid baffle and having simply supported boundaries. This
allows calculation of the radiation efficiency for individual modes below the critical frequency
using Eq. 2.223. For modes with σp,q ≤ 1, regression analysis is used to give a smooth radiation
efficiency curve as a function of frequency. This curve can then be used to determine the
radiation efficiency at the band centre frequencies below the critical frequency. As with method
no. 3, it is necessary to set σ = 1 in the lowest frequency band for which σ > 1, and set σ = 1
in all higher frequency bands. The example in Fig. 2.65d will be used in Section 4.3.1.3.4 to
predict the sound reduction index.
There are advantages in using regression rather than calculating a band-average radiation
efficiency from the modes that lie within each band. Assigning individual modes to particular
bands takes no account of uncertainty in calculating the mode frequencies. It also results
in bands that are not assigned any modes, even though there will be overlapping response
from modes in adjacent bands. These problems are avoided by using regression to give a
smooth curve. The result is generally easier to interpret, and easier to compare with individual
measurements which will usually have more irregular curves.
This method is well-suited to plates in the laboratory situation where the boundary conditions
around the perimeter are uniform, there is minimal interaction between the test element and
the laboratory structure, and where the quasi-longitudinal phase velocity or bending stiffness
212
Chapter 2
of the plate is known, or can be measured. It is less suitable for predictions in real buildings
because the boundary conditions are more variable, and the baffles are usually perpendicular
to the wall or floor. It is not suited to orthotropic plates with widely separated critical frequencies
because of large differences between the radiation efficiencies of the individual modes; in this
situation the band average is more appropriate (Craik, 1996).
2.9.4.5 Plates connected to a frame
Thin sheets or boards that are used to form lightweight walls and floors (e.g. plasterboard,
chipboard) are usually connected to a framework of studs, battens, or joists. With so many
different types of frames and connections, there is no single model for the effect of the frame
on sound radiation; in fact more than one model may be needed to cover a range of frequencies.
A light, flexible frame may have negligible effect on the radiation efficiency of the plate. If the
frame causes the plate to become slightly orthotropic (i.e. the two critical frequencies are not
wide apart) then the effective bending stiffness (Eq. 2.97) can be used to estimate a single
critical frequency. This can then be used to calculate the radiation efficiency for the plate.
Stiff, heavy frames can cause the plate to act as a number of sub-panels between the framing
members. When these sub-panels have similar spatial average vibration levels (but incoherent vibration fields), then the radiation efficiency of the plate is increased below the critical
frequency. The plate can be modelled as a single plate using method no. 1 or no. 2, but
Eq. 2.227 needs to be modified by replacing the plate perimeter, U, with U + Lframe , where
Lframe is twice the total length of the internal frame (Maidanik, 1962). Another possibility is that
sub-panels connected to different parts of the framework will have different boundary conditions
and different baffle orientations. The plate can then be modelled by a number of sub-panels
(assuming that each sub-panel acts as a reverberant plate). This approach has been used
for lightweight cavity walls (e.g. plasterboard walls on a timber frame) so that the boundary
conditions and baffle orientation can be dealt with individually at each timber stud (Craik and
Smith, 2000a).
Sheets or boards used to form lightweight walls and floors are usually screwed or nailed to
the framework. The boundary condition of the sub-panel is affected by whether the junction
between the plate and the frame acts as a line connection, or a number of point connections; this
affects the radiation efficiency. As a rule-of-thumb the transition between screws/nails acting as
a line connection (low frequency model) to individual point connections (high frequency model)
starts when the screw/nail spacing is approximately equal to half a bending wavelength (Craik
and Smith, 2000b). In practice there will be differences between screw/nail fixings on different
boards and different frames. This is partly due to the contact area between the board and the
frame being larger than the cross-section of the screw or nail (Bosmans and Nightingale, 1999).
2.9.5
Radiation into a porous material
For a plate undergoing bending wave motion and radiating into air, the radiation efficiency can
be used to predict the coupling loss factor between a plate, and a room or an empty cavity.
Many cavities in walls and floors are partly or fully filled with porous materials, and walls are
sometimes covered with a porous material. If the porous material is very close to the plate,
but does not actually touch it, then the porous material can be treated as an equivalent fluid to
calculate a radiation efficiency (Tomlinson et al., 2004). This may be important in calculating
213
S o u n d
I n s u l a t i o n
the total loss factor of the plate because the radiation efficiency into a porous material is
usually higher than into air. For plates with low internal loss factors and low coupling losses,
this becomes an important loss mechanism. Low internal losses are often associated with thin
metal plates hence this is relevant to metal roofing and metal cladding systems. For a thin metal
plate, the radiation damping can differ depending on whether the porous material is glued or
loose-laid on the surface (Trochidis, 1985). When it is not touching, the damping significantly
increases as the porous material is moved closer to the surface of the plate (Cummings et
al., 1999). Plasterboard has higher internal losses than most metal plates and so the effect
of radiation damping is less significant. However, sheets of plasterboard that form lightweight
walls and floors do tend to have higher total loss factors when there is a porous material
immediately next to them within the cavity.
To predict the direct sound transmission when a porous material is touching, or is fixed to
the plate, the individual aspects of damping, radiation, sound propagation through the porous
material, and vibration transmission into the frame of the porous material cannot be considered
in isolation from each other. For this kind of sandwich plate (i.e. plate–porous material–plate),
Biot theory can be used to model wave propagation in the porous material and predict the
sound reduction index (Lauriks et al., 1992).
2.9.6 Radiation into the soil
Concrete ground floor slabs undergoing bending wave motion are often highly damped due
to high radiation losses into the soil. For bending wave motion on a plate coupled on one
side to a semi-infinite elastic homogeneous medium representing the soil, both compressional
(longitudinal) and shear waves are radiated into the soil (Heckl, 1987). The radiation efficiency
depends on how the slab is excited. For excitation by a line force on a 600 mm concrete slab
with one side coupled to the soil, the coupling loss factor due to radiation into the soil is predicted
to be ≈10 dB higher in the low-frequency range than the total loss factor for the same slab that
is rigidly connected to other masonry/concrete plates along all edges and only radiates into air
(Villot and Chanut, 2000). Measurements on a 125 mm concrete slab using point excitation in
Fig. 2.66 indicate similarly high losses in the low- and mid-frequency range. Radiation losses
into the ground cause much higher values for the total loss factor than can be attributed to the
sum of the internal loss factor and other coupling losses. Note that these high radiation losses
do not always cause a significant decrease in vibration with distance across the floor.
Ground floor slabs form part of many flanking paths in a building and it is necessary to account
for the high damping of such slabs in a prediction model. This is not easily predicted and can be
quite variable which sometimes makes measurement necessary. In some masonry/concrete
buildings, errors in estimating the total loss factor of the floor slab will have negligible effect
on the prediction of airborne sound insulation between adjacent rooms when all other flanking
paths are included (Craik, 1996b).
2.9.7 Nearfield radiation from point excitation
When a mechanical form of point excitation excites bending waves (such as the ISO tapping
machine, electrodynamic shaker, force hammer) there is not only sound radiation from the
plate bending modes, but also from the nearfield generated in the vicinity of the excitation point.
214
Chapter 2
Loss factor (dB)
118
116
Measured (shown with
95% confidence intervals)
114
ILF, ηint ⫽ 0.005
112
Predicted TLF excluding
radiation into the ground
110
108
106
104
102
100
98
96
50
80
125
200
315
500
800
1250
2000
3150
5000
One-third-octave-band centre frequency (Hz)
Figure 2.66
Measured total loss factor of a 125 mm cast in situ concrete slab (ρ = 2400 kg/m3 , cL = 4200 m/s) on well-consolidated
hardcore blinded with sand. (NB Structural coupling losses to connected walls were relatively low for this slab because there
were only two connected brick walls, each forming an L-junction with the slab.)
Measured data from Hopkins are reproduced with permission from ODPM and BRE.
This can be important below the critical frequency where the radiation efficiency for bending
modes is well-below unity, although its contribution is often negligible.
The sound power radiated by the nearfield, Wn , can be calculated for point excitation on an
infinite homogeneous plate and is given by (Cremer et al., 1973)
Wn =
2
ρ0 c0 k 2 Frms
2πω2 ρs2
for f ≪ fc
(2.231)
2
is the mean-square force.
where Frms
Note that this nearfield power is independent of the plate stiffness and damping, and only
depends upon the mass per unit area.
For comparison, the power radiated by the bending wave field, Wb , due to point force excitation
on a finite homogeneous plate is
*
+
1
ρ0 c0 σ 2
Frms Re
(2.232)
Wb =
ωηρs
Zdp
Hence Wb can be compared directly with Wn using the ratio,
*
+
1
2πωρs σ
Wb
=
Re
for f ≪ fc
Wn
ηk 2
Zdp
(2.233)
and by substituting the driving-point impedance for an infinite homogeneous plate (Eq. 2.190)
this can be simplified to (Fahy, 1985)
πfc σ
Wb
=
Wn
4f η
for f ≪ fc
(2.234)
215
S o u n d
I n s u l a t i o n
18
140 mm concrete slab (fc = 122 Hz, η ⫽ 0.005 ⫹ 0.3f ⫺0.5)
22 mm chipboard (fc ⫽ 1340 Hz, η ⫽ 0.01)
10 lg (Wb/Wn,15 mm) (dB)
16
15 mm OSB (fc ⫽ 1683 Hz, η ⫽ 0.01)
14
12
10
8
6
50
80
125
200
315
500
800 1250 2000
One-third-octave-band centre frequency (Hz)
3150
5000
Figure 2.67
Ratio of the sound power radiated by plate bending modes to the sound power radiated by the nearfield due to point excitation
by a tapping machine hammer with radius, 15 mm. Plate dimensions are Lx = 4 m, Ly = 3.5 m and it is assumed that there
are baffles along the plate perimeter that are perpendicular to the plate surface. Curves valid for f ≪ fc .
Equation 2.234 is sufficient for most estimates such as for point excitation from a narrow
rod connected to a shaker. However, to take account of the point force acting over a finite
circular contact area with radius, r , the radiated power from the nearfield can be represented
by radiation from a piston with a radius, r + λB /4, where (Cremer et al., 1973)
Wn,r =
2
1
λB 4 Frms
ρ0 c0 k 2 π r +
2
2
4
Zdp
for f ≪ fc
(2.235)
and using the infinite plate impedance, the ratio of Wb to Wn,r , is
c04 σ
Wb
=
Wn,r
π5 ηfc f 3 r +
λB 4
4
for f ≪ fc
(2.236)
For point excitation from the ISO tapping machine it can be assumed that the hammer is
flat and circular with a radius of 15 mm. Figure 2.67 shows the ratio of Wb to Wn,15 mm for
three different plates calculated using Eq. 2.236. The loss factors used for these plates represent minimum total loss factors; for the chipboard and OSB these are based on the internal
loss factor. Therefore the ratio represents the maximum that could occur in practice. The
curves are only valid for f ≪ fc , but indicate that the power radiated by the nearfield on a
concrete floor can usually be ignored because the plate radiation from the bending waves
will dominate; this may not be the case with highly damped plates that form lightweight
floors.
One way of reducing sound radiation from a plate excited by a mechanical point source is
to increase the internal damping of the plate by applying a damping compound, or strips of
damping material. This would be relevant to impact sound from thin timber plates that form the
walking surface of a floor (sound is radiated into the room with the walker as well as into the
216
Chapter 2
cavity of a timber floor), or thin metal sheets used for roofing with rain excitation. In such cases
it is useful to estimate the nearfield power because there is a limit as to how much damping
material is worth applying; eventually the nearfield radiation will become more important than
radiation from the highly damped bending waves.
Similar expressions can be found for line excitation of a plate boundary (see Fahy, 1985).
2.10 Energy
For homogeneous beams or plates, the energy associated with each wave type is given by the
product of the mass of the beam or plate and the temporal and spatial average mean-square
velocity associated with that wave motion using
E = mv 2 t,s
(2.237)
For bending waves on plates, this calculation uses the wave motion perpendicular to the plane
of the plate, i.e. in the z-direction. However, beams of rectangular cross-section can support
a bending wave in both the y- and z-directions; hence energy can either be assigned to each
direction separately or the two energies can be combined.
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Chapter 3
Measurement
3.1 Introduction
M
easurements generally fall into three categories: laboratory measurements that provide information at the design stage, field measurements that demonstrate whether
the required sound insulation has been achieved in a building, and field measurements
that help an engineer solve sound insulation problems in existing buildings.
For many buildings the acoustic requirements are described in building regulations; hence
repeatability, reproducibility, and relevance (i.e. the link between the measured sound insulation and the satisfaction of the building occupants) are particularly important for airborne
and impact sound insulation. Laboratory measurements of the acoustic properties of materials
and building elements (e.g. walls, floors, windows, doors) are primarily used for comparing
products and calculating the sound insulation in situ. Measurements of material properties are
particularly useful in assessing whether one material in the construction could be substituted
for a different one, and for use in prediction models.
This chapter gives an overview of measurements that are relevant to sound insulation, outlining the basic principles alongside the underlying assumptions or limitations. For all procedural
aspects of measurements, requirements on laboratory facilities and test elements, and calculation of single-number quantities the reader is referred to the relevant National or International
Standards. In this chapter the latter are generally referred to as the ‘relevant Standard’ with a
reference to the Standard that was current at the time of writing.
3.2 Transducers
This section gives a brief overview of aspects relating to the transducers commonly used to
measure sound pressure and vibration; these are microphones and accelerometers.
3.2.1 Microphones
To measure sound pressure levels or reverberation times in reverberant fields we require an
omnidirectional microphone so that the response is independent of the direction of incident
sound. Over the building acoustics frequency range we would ideally like to measure sound
pressure without altering the sound field due to the presence of the microphone. However, the
finite dimensions of a microphone and its acoustic impedance mean that this is only achieved
when the wavelengths are large. For small wavelengths, diffraction effects from the microphone
become significant.
Microphones are usually designed to give a flat frequency response in a particular sound
field; typically a pressure field, free-field, or diffuse (random-incidence) field (Anon, 1996).
221
S o u n d
I n s u l a t i o n
In a pressure field, the sound pressure has the same magnitude and phase at all points, when
very close to (or flush with) a reflective surface, and within a very small closed cavity such as in
a sound level calibrator. In a free-field, waves propagate without any influence from reflecting
objects or surfaces, such as in an anechoic chamber. In a diffuse field, there is equal probability
of a wave arriving at the microphone from any direction. Although microphones are intended or
optimized for one specific field they can sometimes be used with negligible error in other fields.
A diffuse field microphone is specifically designed to give a flat response in a random-incidence
sound field. In practice, only close approximations to diffuse fields are encountered. These
microphones are suitable for sound pressure measurements in reverberant rooms with varying degrees of diffusivity. However, pressure field and free-field microphones may also be
considered for use in reverberant sound fields. Free-field microphones are designed to give a
flat frequency response at normal incidence (i.e. perpendicular to the microphone diaphragm).
The design aim is usually to achieve this over the majority of audio frequencies, which encompasses the building acoustics frequency range. However, the presence of the microphone in
a free-field alters the measured sound pressure depending upon the angle of incidence. For
angles of incidence other than normal incidence, the response is no longer flat at frequencies
in the high-frequency range. For a half-inch microphone, the deviation at specific angles in
the high-frequency range is typically within ±2 dB of the response at normal incidence. However, when the response of a free-field microphone is averaged over all angles of incidence to
calculate its response to random-incidence, the resulting response is often relatively flat over
the building acoustics frequency range. If it is not sufficiently flat, a special correcting device
can usually be attached to a free-field microphone to achieve a flat response. Similarly to freefield microphones, some half-inch pressure microphones (or smaller) also have a relatively flat
response in random-incidence sound fields over the building acoustics frequency range.
To measure façade sound insulation the external microphone is either fixed to the surface of the
test element (e.g. window) or positioned 2 m from the façade (ISO 140 Part 5). The incident
sound field may be from a loudspeaker at 45◦ , or from an environmental noise source at a
variety of different angles. For the latter it is often reasonable to assume random-incidence;
hence either pressure field, free-field or diffuse field microphones with a relatively flat response
in random-incidence sound fields can be used. Surface measurements require a half-inch
microphone or smaller to avoid interference minima in the building acoustics frequency range;
this is because of the distance between the centre of the microphone diaphragm and the test
element (refer back to Section 1.4.1). When the microphone diaphragm is almost parallel to
the direction of the direct sound, pressure field microphones may be used; otherwise free-field
microphones are appropriate if their directivity in the high-frequency range has negligible effect.
For measurements of façade sound insulation, a windshield should be used for outdoor
measurements of the sound pressure level. Windshields are not usually necessary for indoor
measurements unless there is significant airflow, but they are a useful way of protecting the
microphone whilst carrying it around a building.
3.2.2 Accelerometers
It is important to ensure that the accelerometer is capable of accurately quantifying the actual
vibration. In particular it is important to consider the fixing of the accelerometer, and how the
mass of the accelerometer can alter the vibration of lightweight structures. Details relating to
accelerometer mounting are contained in the relevant Standard (ISO 5348).
222
Chapter 3
Bending wave
measurement
Propagation direction
or
In-plane wave
measurement
Figure 3.1
Accelerometer alignment for measuring bending or in-plane wave vibration on a beam or plate. The arrow on the accelerometer
indicates the main axis of sensitivity.
The type of motion that we want to measure on beams and plates is usually bending or in-plane
wave vibration. An accelerometer is designed to be sensitive to motion along one main axis,
hence the accelerometer is aligned as shown in Fig. 3.1. Although there is only one main axis
of sensitivity, the accelerometer will also exhibit minor sensitivity to transverse motion. When
measuring bending wave motion the effect of this transverse sensitivity is usually negligible
when the accelerometer is mounted flush to a flat surface. However, it needs to be considered
when trying to measure in-plane wave motion in the presence of bending waves.
3.2.2.1 Mounting
Bees wax or petroleum wax is a very convenient method for mounting accelerometers on a
building element. However it is important to pay careful attention to the measured levels in the
high-frequency range because weak fixing can cause measurement errors. Strong fixing can
be obtained by smearing a thin layer of wax over the contact surface of the accelerometer. To
fix it to the surface, a fair degree of pressure needs to be applied to the top of the accelerometer
(usually with the thumb), and it sometimes helps to use a slight twisting motion. If too much wax
is applied, or the surface is fragile, crumbly, dusty, or powdery it can be difficult to achieve a
strong fixing. This usually results in low vibration levels being measured in the high-frequency
range. With simultaneous measurement of acceleration on the same surface using two or more
accelerometers, one accelerometer with weak fixing can usually be identified by significantly
lower levels at high frequencies, often with a sharp drop-off. If the surface texture prevents
a strong fixing with wax, then small metal washers can be glued/cemented to the surface to
allow wax fixing or fixing via a stud on the washer (see ISO 5348).
3.2.2.2 Mass loading
When an accelerometer is fixed to the surface of a structure, a mass has effectively been added
to the structure that can reduce the vibration level at the measurement point. This is referred
to as mass loading. Assuming a lump impedance model with a force source, the relationship
between the measured and the actual velocity is
vmeasured = vactual
Zdp
Zdp + iωmacc
(3.1)
where macc is the mass of the accelerometer and Zdp is the driving-point impedance of the
structure.
State-of-the-art accelerometers that are used to cover the building acoustics frequency range
often have a low mass. Hence because plates that form building elements tend to have
223
S o u n d
I n s u l a t i o n
relatively high surface densities, mass loading is rarely a problem. The effects of mass loading
can be avoided when the accelerometer impedance is much less than the plate impedance,
(3.2)
ωmacc ≪ Zdp
For a thin homogeneous isotropic plate this requirement can be calculated using the infinite
plate impedance (Eq. 2.190), which gives
macc ≪
0.37ρcL h2
f
(3.3)
Mass loading is rarely an issue with masonry/concrete elements but it needs to be checked in
the high-frequency range with building elements such as glass or plasterboard.
3.3 Signal processing
In any measurement, the signal processing has the potential to affect the measured response.
This section contains a brief overview of signal processing for aspects that are relevant to the
majority of sound and vibration measurements described in this chapter. It is restricted to filter
analysis and does not cover measurement techniques using the Fast Fourier Transform for
which many other texts are available (e.g. see Randall, 1987).
From the viewpoint of the analyser, the analogue input signal is a time-varying voltage; therefore the processing applies to both sound and vibration signals. The analogue input signal from
sound or vibration transducers is continuous in time, but in the analogue-to-digital conversion
it is sampled so that the subsequent analysis is carried out on discrete time data. For filter
analysis, the digitized input signal passes through the filters and then the detector as shown
in Fig. 3.2.
The effect of signal processing on the measurement of reverberation time is a specific issue
that is discussed in Section 3.8.3.
3.3.1 Signals
Signals can be considered in two distinct groups: stationary and non-stationary.
‘Stationary’ is used to indicate no change over time, or no change during the measurement
period. Stationary signals are commonly used in sound insulation measurements and fall into
two categories: random and deterministic signals. For stationary random signals, the probability
Input
signal
Filter
Squaring
Averaging
Detector
Figure 3.2
Basic components of signal processing using filtering to determine the frequency spectrum.
224
Output
spectrum
Frequency of occurrence
Frequency of occurrence
Chapter 3
⫺4
⫺2
0
x
2
4
⫺4
⫺2
0
x
Sample 1
4
2
4
Sample 2
3
2
x
1
...continuing to
t⫽∞
0
⫺1
⫺2
⫺3
⫺4
0
∞
Time, t (s)
0.4
Gaussian probability density function
p(x)
p(x) ⫽
0.2
(x ⫺ μ)2
1
exp ⫺
σ 2π
2σ2
μ⫽0
σ⫽1
0
⫺4
⫺3
⫺2
(μ⫺4σ) (μ⫺3σ) (μ⫺2σ)
⫺1
(μ⫺σ)
0
(μ)
x
1
(μ⫹σ)
4
2
3
(μ⫹2σ) (μ⫹3σ) (μ⫹4σ)
Figure 3.3
Gaussian white noise – finite and infinite length samples.
density function for the signal does not change with time. An example of a stationary random
signal is Gaussian white noise. Stationary deterministic signals can be predicted at any point in
time, such as a sinusoidal signal or a signal composed of a number of sinusoids. A specific type
of deterministic signal is a Maximum Length Sequence; this will be discussed in Section 3.9.
Non-stationary signals used in measurements are typically transients. For example, an impulse
generated by a gunshot to measure the reverberation time in a room, or an impulse generated
by a hammer hit to measure the velocity level difference between two walls.
White noise is a stationary random signal with constant power spectral density. Gaussian
white noise is a particularly useful type of white noise because its statistics are described by
the normal (Gaussian) probability distribution. Figure 3.3 shows a record of Gaussian white
225
S o u n d
I n s u l a t i o n
noise in terms of a positive or negative signal amplitude, x, varying with time. The white noise
shown here has a population mean of zero and a population standard deviation of unity. By
using a mean of zero, the standard deviation is equal to the rms value of the signal. For
an infinite length sample, the distribution of amplitudes is described by the smooth curve
of the probability density function, p(x), for a normal distribution. In practice, we use finite
length samples of white noise; Fig. 3.3 shows the distributions for two different samples using
histograms. The difference between these two histograms serves as a reminder that we will
soon need to consider the uncertainty associated with finite length samples when averaging
white noise over time in frequency bands. Fortunately, calculation of the statistics for temporal
averaging are simplified when the probability density function is known.
A continuous random signal, x(t), can be described using the expected value and the autocorrelation function. The expected value, E[x(t)], is the population mean from all realizations
of a random process; in keeping with the above discussion we will take this as zero. The
auto-correlation function equals the expected value of the product of the signal with itself at
another point in time, i.e. E[x(t − t1 )x(t − t2 )] where t1 and t2 , are two different points in time.
For stationary white noise there will be no statistical correlation between the signal at t1 and t2
when t1 = t2 ; this is due to the random nature of the signal. However, there will be correlation
at the same point in time, i.e. when t1 = t2 . E[x(t − t1 )x(t − t2 )] is therefore zero when t1 = t2
and non-zero when t1 = t2 . The auto-correlation function for white noise is therefore defined
using the Dirac delta function, δ(t2 − t1 ). A continuous random signal is defined as white if its
expected value and its auto-correlation function, Rxx (t2 − t1 ), satisfy
E[x(t)] = 0
(3.4)
Rxx (t2 − t1 ) = E[x(t − t1 )x(t − t2 )] = Nδ(t2 − t1 )
(3.5)
where N is the power spectral density.
The spectral density is determined by taking the Fourier transform of the auto-correlation function. The Fourier transform therefore gives equal power spectral density at all frequencies.
However, constant power per Hertz over an infinitely wide band of frequencies would result in
the total power being infinite. Hence although the theory for Gaussian white noise proves to be
useful in a number of derivations, in practice we only measure with band-limited white noise
signals that cover a broad-band of frequencies. For white noise, the one-third-octave-band or
octave-band level increases by 3 dB per doubling of the band centre frequency.
Pink noise is also used for airborne sound insulation measurements. The power spectral density
of pink noise is proportional to 1/f ; therefore the one-third-octave-band or octave-band level
decreases by 3 dB per doubling of the band centre frequency. Pink noise can be generated
from a white noise signal using cascaded filters to give a −3 dB/octave rolloff.
Measurements of airborne sound insulation generally use band-limited white or pink noise
signals; although quite often reference is simply made to broad-band noise. Whatever the input
signal, the spectrum of the measured sound pressure level in the source room will be affected
by the response of the room. For this reason the signal is often shaped using a graphic equalizer
as there is a requirement for the source room that the difference between the sound pressure
levels in adjacent one-third-octave-bands is no greater than 6 dB (ISO 140 Parts 3 & 4).
226
Chapter 3
⫺10
0
Attenuation (dB)
10
Bandwidth, B
70
∞
80
f
f1
Frequency (Hz)
f2
Figure 3.4
Ideal filter.
3.3.2 Filters
Filtering of the signal is almost always required for analysis and other calculations. It is important
to ensure that the filtering process does not significantly affect the measurement of a level or
reverberation time.
3.3.2.1 Bandwidth
In building acoustics, it is common to use filters that have a constant percentage bandwidth;
either one-third-octave or octave-bands. For calculations used to predict sound insulation, the
filter bandwidth, B, can be taken as a constant percentage of the band centre frequency, f ,
where B = 0.23f , for one-third-octave-bands, and B = 0.707f , for octave-bands. This corresponds to the concept of an ideal filter that has 0 dB attenuation at all frequencies within the
passband and infinite attenuation outside the passband (see Fig. 3.4). The bandedge frequencies that define the lower and upper frequencies of this ideal passband are f1 and f2 respectively.
The centre (or midband) frequency, f , of the filter is the geometric mean of f1 and f2 ,
f = f1 f2
(3.6)
and the bandwidth is
B = f2 − f1
(3.7)
Real-time filters that are used in measurement equipment cannot exactly recreate the properties of this ideal filter. There will not be uniform attenuation within the passband with a vertical
227
S o u n d
I n s u l a t i o n
rolloff slope at the edges of the passband. The performance requirements for filters and tolerances on the attenuation limits inside and outside the passband are given in the relevant
Standard (IEC 61260). In this chapter, a 6th order Infinite Impulse Response (IIR) Butterworth
filter (six poles) is used to create examples that are indicative of filters that can be used in
practice. Butterworth filters don’t have a particularly steep rolloff slope, but they can give relatively uniform attenuation in the passband. The attenuation for 1000 Hz one-third-octave and
octave-band filters is shown in Fig. 3.5 along with the minimum and maximum attenuation
limits for a Class 1 filter (IEC 61260). Both of the 6th order Butterworth filters are within the
attenuation limits of the Standard. Although the flanks of the filters cover the building acoustics frequency range, the attenuation is sufficiently high outside the passband that for relatively
smooth spectrum shapes (i.e. without discrete tones and steep slopes) this will not significantly
affect measurement accuracy.
Measured levels in one-third-octave-bands are sometimes combined to give the octave-band
level. Figure 3.6 shows the three one-third-octave-band filters that would be combined to form
the 1000 Hz octave-band alongside the 1000 Hz octave-band filter itself. This shows a marked
difference between the rolloff slopes for one-third-octave and octave-band filters. Hence there
can be differences between levels measured with octave-band filters, and octave-band levels
calculated from measurements with one-third-octave-band filters. The extent of these differences depends on the shape of the spectrum. To avoid dispute it is possible to require that
measurements be taken in one-third-octave-bands, and that these measured values are used
to calculate the octave-band values.
To quantify the effect of the rolloff slope and non-uniform attenuation in the passband it is necessary to introduce other descriptors for the filter bandwidth. These are the effective bandwidth
and the statistical bandwidth. For a stationary random signal, the effective bandwidth applies
to the power passed by the filter, whereas the statistical bandwidth applies to the statistics of
the power passed by the filter, i.e. the variance.
The definition of the effective bandwidth (or noise bandwidth) assumes that a stationary random
signal with constant power spectral density is sent as an input signal to the filter. For this
particular signal the effective bandwidth equals the equivalent bandwidth of an ideal filter that
would pass the same power (mean-square value) as the real filter. It is calculated by numerical
evaluation of the following integral (IEC 61260)
∞
Be =
10−A(f )/10 df
(3.8)
0
where A(f ) is the filter attenuation in decibels as a function of frequency.
The statistical bandwidth is relevant to the variance of random noise power that is passed by
a filter (Bendat and Piersol, 2000; Davy and Dunn, 1987). It is defined for a stationary random
signal that is band-limited. The statistical bandwidth equals the bandwidth of this signal when
the variance of its power (mean-square value) is equal to the variance of the power passed
by the filter. For an nth order Butterworth filter, the statistical bandwidth can be calculated from
the effective bandwidth using (Davy and Dunn, 1987)
Bs =
2n
Be
2n − 1
(3.9)
A 6th order Butterworth filter is the lowest order that would commonly be used in practice, for
which Eq. 3.9 gives Bs = 1.091Be . In the absence of information on specific types of filters it is
often necessary (and usually reasonable) to assume that Bs ≈ Be ≈ B.
228
Chapter 3
(a) 1000 Hz one-third-octave-band filter
⫺10
0
10
Attenuation (dB)
20
30
40
50
60
70
80
100
1000
Frequency (Hz)
10 000
1000
Frequency (Hz)
10 000
(b) 1000 Hz octave-band filter
⫺10
0
10
Attenuation (dB)
20
30
40
50
60
70
80
100
Figure 3.5
Filter attenuation for 6th order Butterworth filters (dotted lines) with a centre frequency of 1000 Hz along with the minimum
and maximum attenuation limits for a Class 1 filter according to IEC 61260 (solid lines).
229
S o u n d
I n s u l a t i o n
⫺10
0
Attenuation (dB)
10
20
30
40
50
60
70
80
100
1000
Frequency (Hz)
10 000
Figure 3.6
Filter attenuation for 6th order Butterworth filters. The 800, 1000, and 1250 Hz one-third-octave-bands are shown in solid lines
with the 1000 Hz octave-band in dotted lines. Note that 800, 1000, and 1250 Hz are the preferred centre frequencies (ISO
266) whereas the exact base-ten frequencies (IEC 61260) are used to define the one-third-octave-band filters.
3.3.2.2 Response time
When a signal is sent through a filter, there will be a short time delay before the filter responds
and gives an output that exactly corresponds to the amplitude of the input signal. This mainly
depends on the filter bandwidth, but also on the type of filter. As an input signal we take a single
sinusoid, sent through a one-third-octave or octave-band filter starting at t = 0. The frequency
of the sinusoid must be within the passband of the filter. The input signal and the filtered output
signal are shown as a function of normalized time in Fig. 3.7. Normalized time is the product of
the filter bandwidth, B, and time, t. It is dimensionless and is useful for looking at general trends.
A rough estimate of the time taken to respond to the input signal and to output a sinusoid with
exactly the same amplitude is usually taken as t = 1/B s, i.e. normalized time of unity (Randall,
1987). Figure 3.7 indicates that this is a slight underestimate for these filters, but it is usually
adequate for practical purposes.
3.3.3 Detector
The detector receives the time-varying AC signal from the filter, squares the signal, and then
carries out the required type of temporal averaging.
3.3.3.1 Temporal averaging
Linear averaging over time gives the equivalent continuous level, Leq . This corresponds to the
level of a continuous steady sound that contains the same energy as the actual time-varying
signal during the integration (averaging) time, Tint , hence
Tint 2
X (t)
1
dt
(3.10)
Leq = 10 lg
Tint 0
X02
230
Chapter 3
Signal (linear units)
(a) One-third-octave-band filter
0
1
2
3
Normalized time, Bt (⫺)
4
5
1
2
3
Normalized time, Bt (⫺)
4
5
Signal (linear units)
(b) Octave-band filter
0
Figure 3.7
Time delay due to filtering of a sinusoidal input signal (thin line) applied at t = 0, and the filtered output signal (thick line).
Dotted lines indicate the minimum and maximum amplitudes of the input signal.
where X is p, d, v, or a (sound pressure, displacement, velocity, or acceleration) and X0 is the
corresponding reference quantity.
The variation in level with time is assessed using the time-weighted level, L(t); this makes use
of an exponential weighting term, and is given by
t −u
1 t X 2 (u)
exp
−
du
(3.11)
L(t) = 10 lg
τ −∞ X02
τ
231
S o u n d
I n s u l a t i o n
where u is a dummy time variable and τ is the time constant for Fast (F) or Slow (S)
time-weighting (IEC 61672 Part 1). The Fast and Slow time constants are 125 ms and 1 s
respectively. There is also an Impulse (I) time-weighting where τ = 35 ms, although it is no
longer recommended for rating impulsive sounds (IEC 61672 Part 1).
The maximum time-weighted level, LFmax or LSmax , is the maximum value of L(t) within a defined
time interval (IEC 61672 Part 1).
The sound exposure level, LE , is used for single noise events such as an aircraft flyover or an
impulsive sound. It gives the steady sound level that when maintained over 1 s contains the
same energy as the actual time-varying signal for the noise event, hence
1 t2 p2 (t)
LE = 10 lg
(3.12)
dt
t0 t1 p02
where t0 = 1 s, with the noise event occurring between times t1 and t2 .
3.3.3.2
Statistical description of the temporal variation
For practical calculations relating to the temporal variation of sound pressure levels it is convenient to assume a Gaussian white noise signal. As previously seen in Fig. 3.3, there will
be variation between each short sample of noise used in the measurements. This uncertainty
can be described using the normalized standard deviation, ε (the ratio of the standard deviation to the mean), which is squared to give the normalized variance, ε2 . For Gaussian white
noise passed through a filter, the normalized variance of the signal due to temporal variation
is (Bendat and Piersol, 2000)
ε2 =
1
Bs Tint
(3.13)
where Bs is the statistical bandwidth of the filter and Tint is the integration time. The statistical
bandwidth Bs is not usually known; if this is the case, it is simplest to assume an ideal filter
with bandwidth, B.
Equation 3.13 shows that the normalized variance decreases with increasing averaging time
and increasing filter bandwidths. However, this equation only applies to the Gaussian white
noise itself. It doesn’t apply to the sound pressure measured at a stationary microphone position
in a reverberant room that is excited by Gaussian white noise. For sound insulation measurements, we need to quantify the normalized variance of the mean-square sound pressure at
stationary microphone positions in the source or receiving room.
For a source room excited by Gaussian white noise, the normalized variance of the meansquare sound pressure at a fixed point is twice the normalized variance of the noise itself; this
is due to the response of the room (Andres, 1965/1966).
ε2 (pS2 ) =
2
Bs Tint
(3.14)
where Bs Tint ≫ 1 ≫ Bs /f (f is the band centre frequency). This assumes that the spatial distribution of the complex Frequency Response Function in the room has a Gaussian distribution,
and that the room absorption is independent of frequency over the frequency band.
For a receiving room, Michelsen (1982) has shown experimentally that a reasonable estimate
of the normalized variance of the mean-square sound pressure at a fixed point is four times
232
Chapter 3
the normalized variance of the Gaussian white noise signal; i.e. the response of two rooms
has affected the signal,
4
(3.15)
ε2 (pR2 ) =
Bs Tint
The normalized standard deviation from Eq. 3.13, 3.14, or 3.15 can be used to give the standard
deviation in decibels using
σdB ≈ 4.34 ε2 (p2 )
(3.16)
For a sound pressure measurement at a stationary microphone position, the minimum integration time is typically Tint = 6 s (ISO 140) for which the standard deviation in decibels is shown
in Fig. 3.8. Ideally the minimum integration time would be assessed on a case-by-case basis
as it needs to be considered in the context of the uncertainty for the spatial variation of sound
pressure in a room (Section 1.2.7.9). Referring back to Fig. 1.42, the standard deviation is
shown for a 50 m3 room with reverberation times between 0.5 and 1.5 s. As a general observation from Fig. 1.42 and Fig. 3.8 we see that the standard deviation in decibels for the spatial
variation is significantly larger than for the temporal variation. In addition we note that the total
variance for the spatial and temporal variation is the sum of the two variances. Therefore the
minimum integration time that is needed for measurements can be determined by satisfying
the following inequality between the normalized variances for the temporal and spatial variation
(Lubman, 1971)
ε2temporal (p2 ) ≪ ε2spatial (p2 )
(3.17)
For frequencies between 0.2fS and 0.5fS , the normalized variance for the spatial variation
can be estimated from the number of modes in the frequency band (Eq. 1.147). A rule-ofthumb for appropriate integration times can now be found for one-third-octave-bands in this
frequency range with typical room volumes (20–200 m3 ). Because the Schroeder frequency
Standard deviation, σdB (dB)
1.0
0.9
Gaussian white noise signal
0.8
Source room
0.7
Receiving room
0.6
Integration time, Tint ⫽ 6 s
0.5
0.4
0.3
0.2
0.1
0.0
50
80
125
200
315
500
800
1250
2000
3150
5000
One-third-octave-band centre frequency (Hz)
Figure 3.8
Standard deviation for the temporal variation of Gaussian white noise and the temporal variation of the sound pressure level
in reverberant source and receiving rooms when the source room is excited with Gaussian white noise.
233
S o u n d
I n s u l a t i o n
is calculated from the room volume and the reverberation time, this rule-of-thumb is linked to
the reverberation times in the source and receiving rooms, which are denoted as TS and TR
respectively. If the inequality is interpreted as implying a factor of at least 10, then it will be
satisfied when Tint = 6 s in source rooms for TS < 2 s, and when Tint = 12 s in receiving rooms
for TR < 2 s.
For frequencies at and above 0.5fS , the normalized variance for the spatial variation is
dependent upon the filter bandwidth and the room reverberation time (Eq. 1.149); this allows
rules-of-thumb to be developed by substituting Eq. 1.149 along with Eq. 3.14 or 3.15 into
Eq. 3.17. Thus, the integration time for frequency bands above 0.5fS should be chosen such
that Tint > 3TS for the source room and Tint > 6TR for the receiving room.
3.4 Spatial averaging
Due to the spatial variation of sound pressure in a space, or vibration over a surface, we almost
always need to measure temporal and spatial average levels.
The temporal and spatial average sound pressure level in a room is usually referred to as the
average sound pressure level. Temporal and spatial average values for sound pressure and
velocity levels don’t tend to use different symbols; they simply use Lp and Lv respectively, and
rely on accompanying text to make it clear.
For N individual measurement positions, where Leq,n is measured at each position, the temporal
and spatial average level is
N
1 Leq,n /10
L = 10 lg
(3.18)
10
N
n=1
For sound pressure, spatial averaging can also be carried out with a continuously moving
microphone. In this case, Lp is equal to the measured Leq .
The sample standard deviation, sdB , can be calculated using
!
"
" N L2 − 1 N L
# n=1 eq,n
n=1 eq,n
N
sdB =
N−1
2
(3.19)
Calculation of confidence intervals requires knowledge of the probability distribution for the
variable of interest; this variable is usually the sound pressure level or the velocity level. Exact
probability distributions will vary and are rarely known (Lyon, 1969). For mean-square pressure, a gamma distribution applies to the source room and a log-normal distribution can be
assumed in the receiving room (Section 1.2.7.9). For mean-square velocity, it is also reasonable to assume a log-normal distribution (Section 2.7.6). Therefore, it can be assumed that
the spatial variation of sound pressure levels and vibration levels in decibels is described by
the normal (Gaussian) probability distribution (Craik, 1990; Lyon and DeJong, 1995; Weise,
2003). An estimate of the 95% confidence interval can therefore be calculated from the standard
statistical formula,
sdB
t√
(3.20)
N
where t is the t-value of the Student’s t-distribution for N − 1 degrees of freedom and a
probability of 0.05.
234
Chapter 3
3.4.1
Spatial sampling of sound fields
In a reverberant field, the temporal and spatial average sound pressure level is determined by
sampling the sound field at different positions in the central zone of the room; note that these
positions must be a certain distance away from the room boundaries and the sound source.
Spatial sampling is carried out by using a number of stationary microphone positions or a
continuously moving microphone (e.g. a microphone on a rotating boom). The aim is to use
sufficient samples to determine the average level with an acceptably low level of uncertainty.
3.4.1.1 Stationary microphone positions
Spatial averaging using a limited number of stationary microphone positions is only effective if
a sufficiently large number of positions are used, and these positions provide samples of the
mean-square pressure that are uncorrelated in space. Correlation between the sound pressure
at two microphone positions separated by a distance, d, in a three-dimensional diffuse sound
field can be assessed by using a spatial correlation coefficient; this coefficient will be unity
when the positions are the same, i.e. when d = 0.
A diffuse sound field can be considered as being comprised of plane waves of equal amplitude
and random phase. In this field there is equal probability of a plane wave arriving at the two
positions from any angle, θ. By initially considering a single plane propagating wave as shown
in Fig. 3.9, the sound pressures at the two positions can be defined as
p1 (t) = p̂ exp(iφR ) exp(iωt)
(3.21)
p2 (t) = p̂ exp(−ikd cos θ) exp(iφR ) exp(iωt)
(3.22)
and
where p̂ is an arbitrary constant for the peak amplitude and φR is a random phase for the
incident plane wave.
Propagation direction
θ
p1
p2
x
d
Figure 3.9
Measurement of sound pressure at two points in a propagating plane wave.
235
S o u n d
I n s u l a t i o n
The spatial correlation coefficient, R12 (kd), is
(
p1 p2 t
R12 (kd) =
p12 t p22 t
)
(3.23)
θ,φ
where θ,φ indicates the average over all possible angles in three-dimensional space.
The mean-square sound pressures are calculated by squaring the real parts of Eqs 3.21 and
3.22, then taking the time-average to give
p12 t = p22 t = 0.5 p̂2
(3.24)
The time-average of the product of the real parts of p1 and p2 gives
p1 p2 t = 0.5p̂2 cos (kd cos θ)
(3.25)
Hence the spatial correlation coefficient is (Cook et al., 1955)
2π π
1
sin (kd)
R12 (kd) = cos (kd cos θ)θ,φ =
cos (kd cos θ) sin θ dθ dφ =
4π 0
kd
0
(3.26)
The plot of R12 (kd ) in Fig. 3.10 shows that there is only weak correlation between two positions when kd ≥ π, i.e. when d ≥ λ/2. The correlation coefficient tends to zero when kd ≫ π.
This analysis only applies to single frequencies; however, it can be considered as a reasonable estimate for broad-band signals measured in one-third-octave-bands. Another important
caveat is that it only applies to a three-dimensional diffuse field; for typical rooms this will often
limit its application to the mid- and high-frequency ranges.
Uncorrelated samples can be taken by choosing microphone positions in the central zone of
a reverberant room that are all separated from each other by d ≥ λ/2. In typical rooms the
diffuse field assumption is only appropriate in the mid- and high-frequency ranges, i.e. at and
Spatial correlation coefficient, R12(kd) (⫺)
1
0.8
0.6
0.4
0.2
0
⫺0.2
⫺0.4
0
π
2π
kd
Figure 3.10
Spatial correlation coefficient in a three-dimensional diffuse field.
236
3π
Chapter 3
above 250 Hz. As it is common to use parallel filter measurements over the building acoustics
frequency range, the requirement for uncorrelated samples can be based on 250 Hz, which
means that d ≥ 0.7 m. This requirement is feasible in typical rooms, and is the minimum distance
quoted in sound insulation measurement Standards (ISO 140).
The advantage in using stationary positions is that the standard deviation (Eq. 3.19) and confidence intervals can be calculated as the measurement progresses. This allows the measurer
to make quick decisions as to when the uncertainty in the average sound pressure level is
sufficiently low.
Sometimes it is not possible to find sufficient positions that will give uncorrelated samples and
it is necessary to use d < λ/2. In this situation some positions will be correlated and others will
be uncorrelated. The equivalent number of uncorrelated samples, Neq,s , in the total number of
samples, N, can be calculated from (Lubman, 1971, 1974),
Neq,s =
N
1 N N
1+
[Rij (kdij )]2
N i=1 j=1
for i = j
(3.27)
where dij is the distance between positions i and j.
Inclusion of correlated samples where d < λ/2, means that Neq,s < N. In practice, Eq. 3.27 is of
limited use as it is only applicable to a three-dimensional diffuse field. If we use the Schroeder
cut-off frequency as an indicator of diffuse fields, we find that in room volumes less than 60 m3 ,
it will rarely be possible to use Eq. 3.27 to estimate Neq,s in the low-frequency range (refer back
to Fig. 1.25).
The aim of spatial averaging is to take sufficient samples to determine the mean value with a
low level of uncertainty. In a sound field that can be considered as diffuse, it is better to avoid
using correlated samples to increase the size of a data set because this risks introducing a
greater level of uncertainty in the mean value (Lubman et al., 1973).
3.4.1.2
Continuously moving microphones
One way of generating large numbers of samples is to use a continuously moving microphone.
The most common type of continuously moving microphone uses a rotating boom to trace out
a circular path of radius, r , at uniform speed. The boom is tilted at an angle to avoid averaging
in a plane that is parallel to a room surface. The effectiveness of this method in a diffuse field
depends upon the path radius, the speed, and the total averaging time. An approximation
for the equivalent number of uncorrelated samples along a circular path, Neq,c , is given by
(Lubman et al., 1973)
⎧
⎫
2U
⎪
⎨ 1
⎬
for
< 1⎪
λ
Neq,c =
(3.28)
⎪
⎩ 2U for 2U ≥ 1⎪
⎭
λ
λ
where U is the perimeter length of the circular path (2πr ).
Examples for different radii are shown in Fig. 3.11, including the minimum values referred to
in the Standards, namely r = 0.7 m (ISO 140 Part 4) and r = 1.0 m (ISO 140 Part 3). Note that
Eq. 3.28 is only applicable to three-dimensional diffuse fields, i.e. above the Schroeder cut-off
237
S o u n d
I n s u l a t i o n
frequency. We will assume that a rotating boom with a minimum radius of 0.7 m is typically
used as a substitute for five stationary microphone positions. From Fig. 3.11 it is clear that
the rotating boom becomes a very efficient way of generating large numbers of uncorrelated
samples in the mid- and high-frequency ranges. However, in the low-frequency range, the
required radius tends to becomes impractically long to achieve a minimum of five uncorrelated
positions (2.75 m at 50 Hz).
An alternative path for a continuously moving microphone is along a straight line. The equivalent
number of uncorrelated samples, Neq,l , along a line of length, L, falls within the range (Lubman
et al., 1973)
2L
2L
< Neq,l < 1 +
λ
λ
(3.29)
When commissioning a laboratory, there is time available to validate stationary and/or continuously moving microphones. In contrast, time is very limited in the field. The Schroeder
frequency can be estimated so the ideal approach might be considered as using stationary
positions below the Schroeder frequency to allow calculation of the standard deviation, and a
rotating boom above the Schroeder frequency. In practice, using both methods in the field is
far too time-consuming. For this reason, one method or the other is chosen; apart from practical considerations the choice may depend upon which part of the frequency range is deemed
to be more important or whether calculation of the standard deviation would be useful when
interpreting the results.
3.4.2
Measurement uncertainty
The sound insulation descriptors in this chapter are calculated from different temporal and
spatial average values. Hence we need to be able to calculate the uncertainty from combinations of measured temporal and spatial average values. Combining the variances is made
Equivalent number
of uncorrelated samples, Neq,c (⫺)
100
10
Radius, r (m)
0.4
5
0.7
1.0
1.3
1.6
1
50
80
125
200
315
500
800 1250 2000
One-third-octave-band centre frequency (Hz)
3150
5000
Figure 3.11
Equivalent number of uncorrelated samples for a continuously moving microphone on a circular path of radius, r.
238
Chapter 3
complicated by the different probability distributions (Bodlund, 1976), hence for practical
purposes it is necessary to make simplifying assumptions.
The majority of equations take the form,
Y (dB) = X1 (dB) ± X2 (dB) ± X3 (dB) ± . . .
(3.30)
For each individual component in Eq. 3.30, the sample standard deviation in decibels, sdB , can
be calculated using Eq. 3.19. If the variances for each individual component are uncorrelated,
then the combined variance can be estimated from the sum of the individual variances using
2
2
2
2
sdB
(X3 ) + . . .
(Y ) ≈ sdB
(X1 ) + sdB
(X2 ) + sdB
(3.31)
To estimate the uncertainty in the single-number quantities that are used to quantify sound
insulation (e.g. DnT ,w , Rw , Ln,w ) it is possible to use Monte-Carlo methods (Goydke et al.,
2003).
3.5 Airborne sound insulation
Laboratory airborne sound insulation measurements are primarily used to compare the sound
insulation provided by different test elements and to calculate the sound insulation in situ. The
role of field measurements is usually to check that a certain level of sound insulation has been
achieved.
3.5.1 Laboratory measurements
A transmission suite for airborne sound insulation measurements comprises two rooms; a
source room and a receiving room separated by a test element (see Fig. 3.12). At this stage
we simply assume that all sound is transmitted via the test element, and that the structure
Source
room
(1)
Test
element
Receiving
room
(2)
Figure 3.12
Outline sketch of a transmission suite for airborne sound insulation measurements.
239
S o u n d
I n s u l a t i o n
of the transmission suite itself plays no role other than defining the space for the source and
receiving rooms.
The transmission coefficient, τ, is defined as the ratio of the sound power transmitted by the
test element, W2 , to the sound power incident on the test element, W1 ,
τ=
W2
W1
The sound reduction index, or transmission loss, R, in decibels is defined as
W1
1
= 10 lg
R = 10 lg
τ
W2
(3.32)
(3.33)
By assuming that the sound fields in the source and receiving room are diffuse we can calculate
the incident and the transmitted sound power from sound pressure level measurements in each
room. To find these sound powers, the first step is to calculate the sound intensity incident
upon any surface in a diffuse field in terms of the average sound pressure level in the room.
This makes use of the mean free path, dmfp (Section 1.2.3.1). For any shape of room with
walls and floors that diffusely reflect an incident sound wave, the mean free path is given by
Eq. 1.47. Whilst this assumes diffuse reflection from the room surfaces, most walls and floors
in a transmission suite are flat and smooth with relatively uniform impedance. For this reason
one might assume that specular reflection is more likely to occur than diffuse reflection. In
practice, the reflections are neither specular nor diffuse. The procedures used to commission
transmission suites aim to provide a sound field in the central zone of the source or receiving
room that can be considered as reasonably diffuse. This is often achieved by hanging diffusers
in the room, hence, the reflections can be considered to be partially diffuse. On this basis we
move forward by assuming diffuse reflections. We can now use the mean free path to infer
that sound energy is reflected from the room surfaces c0 /dmfp times every second. The sound
energy in the room is equal to the product of the energy density, w, and the room volume, V .
Hence the intensity that is incident upon any surface in the room is
3 24
p t,s
wc0
c0 1
=
=
(3.34)
I = wV
dmfp ST
4
4ρ0 c0
3 4
where ST is the total surface area of the room and p2 t,s is the temporal and spatial average
mean-square sound pressure in the diffuse field.
To calculate the diffuse field intensity that is incident upon a surface, it is not necessary to
take account of interference patterns at the room boundaries (Vorländer, 1995). Therefore the
power that is incident upon the test element in the source room is
3 24
p1 t,s
W1 = I 1 S =
S
(3.35)
4ρ0 c0
where S is the area of the test element.
The power transmitted into the receiving room must equal the power absorbed in the receiving
room. Hence, for an absorption area, A, in the receiving room, Eq. 3.34 gives the transmitted
power as:
3 24
p2 t,s
w2 c0
W2 =
A=
A
(3.36)
4
4ρ0 c0
240
Chapter 3
We now need to consider the interference patterns at the room boundaries that result in nonuniform distribution of energy density in a reverberant room (Section 1.2.7.1). If we use the
average sound pressure in the central zone of the receiving room to calculate the transmitted
sound power, we will underestimate this sound power and overestimate the sound reduction
index. In theory, we could take account of the higher energy density near the boundaries by
including the Waterhouse correction in the calculation of the transmitted sound power using
3 24
ST λ
p2 t,s 1 +
8V
W2 =
A
(3.37)
4ρ0 c0
where ST is the total surface area of the receiving room and V is the volume of the
receiving room.
The Waterhouse correction is more important in the low-frequency range. For room volumes
between 50 and 150 m3 it is between 3 and 4 dB at 50 Hz, and less than 1 dB above 200 Hz.
The definition of the sound reduction index does not incorporate the Waterhouse correction. This is reasonable because the correction term is not sufficiently accurate to be used
with receiving rooms in transmission suites that have relatively small volumes; the minimum
receiving room volume is 50 m3 (ISO 140 Part 1). In addition, errors may arise from using the
correction term if the presence of large diffusers significantly changes the distribution of energy
density. Hence because no better estimate is available, and the Waterhouse correction is not
sufficiently accurate, a correction term is not incorporated that could cause incorrect ranking
of test results from different laboratories in the low-frequency range. However we will need to
look at this issue again when we want to compare measurements of the sound reduction index
made with sound intensity, to those made with sound pressure.
Returning to the derivation, Eqs 3.32, 3.35, and 3.36 give
3 24
3 24
p2 t,s
p1 t,s
S=
A
τ
4ρ0 c0
4ρ0 c0
3 24
hence,
p1 t,s S
1
= 3 24
τ
p2 t,s A
which is converted to decibels to give the sound reduction index,
3 24
p1 t,s
S
S
+ 10 lg
= Lp1 − Lp2 + 10 lg
R = 10 lg 3 2 4
A
A
p2 t,s
(3.38)
(3.39)
(3.40)
where Lp1 and Lp2 are the temporal and spatial average sound pressure levels in the source
and receiving rooms respectively.
Ideally, the test element would have the same dimensions as the element being assessed
for installation in a building. For practical reasons this is not always possible. Whether it is
absolutely necessary can only be established by measuring different size elements and/or with
a prediction model for the sound insulation. The relevant Standard (ISO 140 Part 3) requires
the following test element areas: ≈10 m2 for walls, and between 10 and 20 m2 for floors where
the smaller dimension is at least 2.3 m for both walls and floors. Hence these wall and floor
areas are representative of those in real buildings. The sound reduction index is an appropriate
descriptor where the test element area is well-defined and the test result is representative of
241
S o u n d
I n s u l a t i o n
identical elements with different dimensions. The latter is a reasonable assumption for many
walls and floors, and, to a lesser degree, for windows and doors. However, it is not appropriate
for many elements that are smaller than 1 m2 such as ventilation devices or cable ducts. For
these small building elements or devices the sound reduction index will rarely be representative
of one with different dimensions. It is therefore misleading to use this descriptor because it
gives the impression that we could use the test result for different sizes of the same element or
device. An alternative descriptor to the sound reduction index is therefore required which does
not involve the test element area, and which is attributable to a single element or device. It is not
appropriate to derive this descriptor using the sound power that is incident upon the element
or device because this uses the test element area; so the alternative descriptor is based on
the sound pressure level difference, D, between the source and receiving rooms where
D = Lp1 − Lp2
(3.41)
For a given sound power input into the receiving room, the mean-square sound pressure in the
room is inversely proportional to the absorption area, A (see Eq. 3.36). Hence the receiving
room sound pressure level needs to be ‘normalized’ to a reference absorption area using the
measured absorption area, A, in that room. For small building elements or devices, this gives
the element-normalized level difference, Dn,e ,
A
A
Dn,e = Lp1 − Lp2 + 10 lg
(3.42)
= D − 10 lg
A0
A0
where the reference absorption area, A0 , is 10 m2 (ISO 140 Part 10).
3.5.1.1 Sound intensity
The sound reduction index of a test element can also be measured using sound intensity. The
sound power incident upon the test element is calculated from sound pressure level measurements in the source room using Eq. 3.35. The transmitted power is calculated from the
average normal intensity, In , over the measurement surface, SM , that has been measured with
an intensity probe,
W2 = In SM
(3.43)
An intensity probe measures net intensity. The importance of this becomes apparent if we consider measuring a test element with an absorbent surface because the intensity measurement
gives an estimate of the net power, i.e. the transmitted power minus the absorbed power. To
avoid underestimating the transmitted power it is necessary to satisfy requirements on FpI and
δpI0 − FpI . These requirements are discussed in Section 3.10 and can be satisfied by introducing
sufficient absorption into the room where the intensity measurements are carried out.
For intensity measurements the sound reduction index in decibels is calculated by making use
of the relationship between ρ0 c0 and the reference quantities for sound pressure (20 × 10−6 Pa)
and sound intensity (10−12 W/m2 ), where 10 lg (ρ0 c0 ) = 10 lg ((20 × 10−6 )2 /10−12 ). Hence, the
intensity sound reduction index, RI , is given by
SM
RI = Lp1 − LIn − 10 lg
− 6 dB
(3.44)
S
where Lp1 is the average sound pressure level in the source room, LIn is the temporal and
spatial average normal sound intensity level over the measurement surface, and S is the area
of the test element.
242
Chapter 3
For small building elements, the intensity element-normalized level difference, DI,n,e , is
calculated using
SM
DI,n,e = Lp1 − LIn − 10 lg
(3.45)
− 6 dB
A0
where the reference absorption area, A0 , is 10 m2 (ISO 15186 Part 1).
In comparison with the sound pressure level method (ISO 140 Part 3), the sound intensity
method has advantages when there is flanking transmission from the laboratory structure, or
from the filler wall used with small test elements. This is because the intensity probe can be
used to measure only the sound intensity that is radiated by the test element. Therefore in the
presence of flanking transmission, RI tends to be higher than R. However, the definition of R
does not take account of the higher energy density near the receiving room boundaries when
calculating the transmitted sound power. This means that RI tends to be lower than R in the
low-frequency range when there is no significant flanking transmission.
3.5.1.1.1
Low-frequency range
In Sections 1.2.7.8 and 1.2.7.9 we saw that there can be significant spatial variation of the
sound pressure levels in the low-frequency range. In addition we know that when the sound
reduction index is determined using sound pressure level measurements in the receiving room,
no account is taken of the higher energy density near the room boundaries when calculating
the transmitted sound power. Hence, in the low-frequency range the transmitted sound power
that is determined from sound intensity measurements in the receiving room should be more
accurate than measuring sound pressure levels in the central zone of the receiving room. To
measure an even better estimate of the sound reduction index in the low-frequency range it is
necessary to make changes to the measurement method in the source room as well as to the
sound field in the receiving room. This is the measurement method proposed by Pedersen et al.
(2000) that is implemented in ISO 15186 Part 3. Inter-laboratory comparisons (Olesen, 2002;
Pedersen et al., 2000) indicate that this approach has better reproducibility than the traditional
method (ISO 140 Part 3) in one-third-octave-bands between 50 and 160 Hz. However, it is
primarily intended for one-third-octave-bands between 50 and 80 Hz (ISO 15186 Part 3).
The measurement accuracy and reproducibility in the low-frequency range are significantly
improved by effectively converting the receiving room into an anechoically terminated duct
(Pedersen et al., 2000; Roland, 1995). This is achieved by placing highly absorbent material on the back wall of the receiving room to try and simulate an anechoic termination (see
Fig. 3.13). In practice an anechoic termination is not possible, but high levels of absorption
can be achieved by using a very thick layer of mineral wool, often up to 1 m in thickness
(ISO 15186 Part 3). The test element therefore radiates sound into a duct-like receiving
room. We can compare common ducts used for HVAC purposes with this duct-like receiving room. For HVAC ducts, the wavelengths in the low-frequency range are usually large
compared to the dimensions of the duct cross-section, and at these frequencies only plane
waves will propagate along the duct. At higher frequencies, the duct cross-section can support modes that propagate along the duct, these are referred to as propagating cross-modes.
A receiving room in a transmission suite has much larger cross-sectional dimensions than a
typical HVAC duct. Hence, for typical receiving rooms there will be propagating cross-modes
in each one-third-octave-band from 50 Hz upwards. The measurement environment in this
duct-like room is well-suited to determining the transmitted sound power using sound intensity
243
S o u n d
I n s u l a t i o n
Transmission suite
aperture
Figure 3.13
Sketch of a receiving room with highly absorbent back wall to simulate an anechoically terminated duct. In practice, thick
layers of mineral wool are used rather than long anechoic wedges. The shaded area indicates the measurement surface for
intensity measurements with a test element that fills the aperture.
measurements; this is in contrast to normal rooms where it can be difficult to achieve acceptable
FpI values, particularly in one-third-octave-bands below 100 Hz.
In the source room a different measurement method is used to avoid the errors incurred by
calculating the incident sound power from sound pressure levels in the central zone of the
source room. This method uses sound pressure measurements near the test element surface.
For sound fields near room boundaries, we saw in Section 1.2.7.1 that very close to a perfectly
reflecting surface in a reverberant room the sound pressure level is 3 dB higher than the diffuse
field level. Hence it can be assumed that the diffuse field sound pressure level will be 3 dB less
than the spatial average sound pressure level over the surface of the test element, Lp1,s , when
measured at a distance less than 50 mm from the surface (ISO 15186 Part 3). To account for
the different measurement method in the source room, the intensity sound reduction index,
RI(LF) , is calculated using
RI(LF) = Lp1,s − LIn − 10 lg
SM
S
− 9 dB
(3.46)
and the intensity element-normalized level difference, DI(LF),n,e , is calculated using
DI(LF),n,e = Lp1,s − LIn − 10 lg
SM
A0
− 9 dB
(3.47)
Values of RI(LF) (or DI(LF),n,e ) in the low-frequency range can be combined with values of R
and/or RI (or Dn,e and/or DI,n,e ) in the mid- and high-frequency ranges to produce a continuous
spectrum between 50 and 5000 Hz.
244
Chapter 3
3.5.1.2
Improvement of airborne sound insulation due to wall linings,
floor coverings, and ceilings
3.5.1.2.1 Airborne excitation
For linings used on separating walls or floors, the improvement due to the lining for direct sound
transmission is determined by measuring the sound reduction index of a base wall or floor with
and without the lining. The sound reduction improvement index, R, is then given by
R = Rwith lining − Rwithout lining
(3.48)
Wall linings often improve the sound insulation in the high-frequency range at the expense of
reducing it in the low-frequency range. Hence whilst it is referred to as an improvement index,
it can take negative or positive values.
R depends upon the type of base wall or floor to which the lining is fixed; hence, the relevant
Standard (ISO 140 Part 16) describes different base walls or floors to allow a fairer comparison
between products.
3.5.1.2.2
Mechanical excitation
Linings used on flanking walls or floors are not always directly excited by a sound field but also
by structure-borne sound transmitted across one or more junctions. A wall or floor may also be
subject to impacts, or machinery/equipment that only acts as a structure-borne sound source
(i.e. without airborne excitation) may be fixed to one side. In these situations there is no nonresonant (mass law) transmission. Below the critical frequency, R includes both resonant
and non-resonant transmission; therefore, it is sometimes useful to measure the improvement
due to only resonant transmission. This is the resonant sound reduction improvement index,
RResonant and can be measured with mechanical rather than airborne excitation. (Note that at
the time of writing there was no published measurement Standard.)
In the laboratory it is convenient to apply mechanical excitation via a shaker to the side of the
wall without the lining. The test procedure requires measurement of the vibration level of the
base wall or floor (on the side of the wall without the lining) along with the resulting sound
pressure level in the receiving room. These measurements are carried out with and without
the lining. To avoid problems in measuring low-sound pressure levels it is convenient to take
dual-channel measurements using an MLS signal (Section 3.9) rather than broad-band noise.
Vibration and sound pressure can then be measured simultaneously; one in each channel.
M excitation positions are used with N different accelerometer positions for each excitation
position. The sound pressure is measured in the receiving room at R microphone positions
for each accelerometer position; stationary positions are required due to the use of MLS. In
a transmission suite, M = 3, N = 4, and R = 2 is usually sufficient. See Section 3.12.3.2.3 for
excitation and accelerometer positions.
For walls, the shaker is usually pushed up against the wall without any transducers to measure
the power input. This means that the same power input may not be applied at all excitation
positions. Therefore the vibration and sound pressure measurements need to be arithmetically
averaged to calculate a standardized pressure–vibration level, Lpv,T , using
5
'6
&
M
NR
N
T
1
1 Lv,n /10
1 Lp,r /10
Lpv,T =
− 10 lg
− 10 lg
10 lg
10
10
M
NR
N
T0
m=1
r =1
m
n=1
m
(3.49)
245
S o u n d
I n s u l a t i o n
where T is the reverberation time in the room where the sound pressure level is measured,
and T0 = 0.5 s. Lpv,T usually takes negative values. Note that for ceilings on floors it is possible
to use the ISO tapping machine to provide a constant power input.
The resonant sound reduction improvement index is then given by,
RResonant = Lpv,T (without lining) − Lpv,T (with lining)
(3.50)
In practice this measurement is more relevant to point excitation of a wall or floor (e.g. impacts,
attached machinery/equipment) than excitation due to structure-borne sound that occurs with
flanking transmission. The reason for this is that mechanical excitation at a point tends to
overemphasize the adverse effect of any mass–spring–mass resonance (see Section 4.3.8.2).
In addition, measurements above the thin plate limit are of limited applicability due to thick
plate effects and because they are specific to the contact area used for point excitation.
3.5.1.3 Transmission suites
The ideal test result from a transmission suite is one that is repeatable within the laboratory,
reproducible in a different laboratory, allows a fair comparison with different test elements, and
is in a form that can easily be used to estimate the sound insulation in situ. It is also important to
be able to validate transmission suite measurements against a theoretical model for relatively
simple test elements (e.g. non-porous, solid, homogeneous, isotropic plates).
There are many different sizes and types of test elements, so certain aspects of a transmission
suite design are often tailored to suit a particular element. For example, test elements such as
masonry walls need to dry out before they are tested, so it is convenient if they can be built
in a frame outside the transmission suite where they can be left to dry. When they are ready
to be tested they can then be ‘inserted’ between the source and receiving rooms. As sound
transmission across this type of element is affected by its total loss factor, the connections
that are made between the frame and the laboratory will affect the measured sound reduction
index. The relevance of total loss factor measurements will shortly be discussed in more detail.
Laboratory tests on separating walls or floors are often used to calculate the sound insulation
in situ. A separating wall or floor in a building usually forms one complete surface in a room,
so the connected flanking walls and/or floors effectively form baffles that are orientated perpendicular to its surface. Below the critical frequency of any plate that faces into a room, the
radiation efficiency is affected by the baffle orientation (Section 2.9.4). For this reason there
is logic in having laboratory walls and floors around the perimeter of the test element that are
perpendicular to the test element surface as well as having a test aperture that has dimensions similar to walls or floors in situ (Kihlman and Nilsson, 1972; ISO 140 Part 1). However
there are competing objectives in the low-frequency range; for measurement accuracy it can
be beneficial to create relatively large source and receiving rooms so that each room has a
high mode count in each one-third-octave-band. For a test aperture area that is ≈10 m2 , the
use of laboratory walls and floors that form perpendicular baffles means that room volumes are
limited to ≈50 m3 ; this is because an elongated room with the longest dimension perpendicular to the test element tends not to provide an incident sound field that can be considered as
diffuse. Another factor that requires consideration is that some test elements (such as curtain
walling used for facades) are much larger than ≈10 m2 and a laboratory structure that forms
perpendicular baffles may be less representative of in situ than baffles that lie in the same
plane as the test element.
246
Chapter 3
3.5.1.3.1
Suppressed flanking transmission
Transmission suites are referred to as having “suppressed flanking transmission’’ (ISO 140
Part 1). This acknowledges the fact that flanking is omnipresent but that it can be suppressed to
such a level that its effect can be considered as negligible; thus allowing useful measurements.
It is difficult to make general statements about the suppression of flanking because there are
many different designs of transmission suite. In addition, flanking transmission is not a fixed
property of the transmission suite itself; it depends on the type of test element, and how and
where the test element is connected to the laboratory structure that defines the source and
receiving rooms.
This is a useful point to introduce the commonly used classifications for transmission paths;
these are Dd, Ff, Fd, and Df, as shown in Fig. 3.14. Direct transmission via the test element
is denoted as Dd and the flanking paths are denoted as Ff, Fd, and Df. These classifications
can be used to consider the different permutations for the installation of test elements in transmission suites. For example, the test element could be structurally isolated from both rooms,
connected to certain walls or floors in one room, or connected around all its boundaries to
the structure of both rooms. These classifications simplify analysis and discussion of flanking
paths, although in Chapter 4 we will introduce SEA transmission paths to allow a more flexible
and detailed analysis of different transmission mechanisms.
The different flanking transmission paths can be described as follows:
(1) Transmission paths involving direct sound transmission between the source and
receiving rooms other than via the test element.
This transmission can occur via a filler wall that is used to temporarily change the
size of the test aperture. In some laboratories the test aperture does not cover one
complete room surface; hence, it can also occur via laboratory walls/floors that are
adjacent to, and in the same plane as the test aperture.
(2) Transmission paths via the laboratory structure (including any filler wall) that do
not enter or involve the test element (path type: Ff ).
(3) Transmission paths via the laboratory structure (including any filler wall) that involve
energy flow into the test element (path type: Fd).
(4) Transmission paths via the laboratory structure (including any filler wall) that involve
energy flow out of the test element (path type: Df).
Consideration of these flanking paths should not lead to the conclusion that a test element
needs to be completely isolated from all parts of the laboratory by using resilient connections
all around its border. In Section 3.5.1.3.2 we will discuss the effect of changing the connection
around the perimeter of heavyweight test elements because this can significantly change the
measured sound insulation.
Transmission suites may be heavyweight or lightweight structures, or a combination of the
two. For masonry/concrete test elements that are rigidly connected to a heavyweight laboratory structure, the coupling between the laboratory structure and the test element depends on
both their thicknesses and material properties. Laboratories with heavyweight structures are
usually built from masonry or concrete, this leads to a rule-of-thumb that the mass per unit area
of a solid masonry/concrete test element should be less than half the mass per unit area of each
connected laboratory wall or floor (Gerretsen, 1990). SEA predictions in a variety of different
laboratory designs also indicate that this is reasonable (Craik, 1992). However, energy flow
247
S o u n d
I n s u l a t i o n
Source
room
Test
element
Receiving
room
Direct path (Dd)
Flanking path (Ff )
Flanking path (Fd)
Flanking path (Df)
Figure 3.14
Transmission paths between two adjacent rooms. Direct transmission via the test element surface is indicated by D in the
source room, and d in the receiving room. Flanking surfaces are indicated by F in the source room and f in the receiving room.
In this illustration the ceiling has been chosen as the flanking element; the principle also applies to the side walls or the floor
in each room. Each flanking path Ff, Fd, or Df is only defined for vibration transmission across one junction (i.e. a physical
connection between the two grey plates); in practice there will also be flanking paths involving more than one junction.
248
Chapter 3
between a test element and a laboratory is sufficiently complex that it is not possible to define
a flanking limit for a laboratory in terms of one maximum achievable sound reduction index;
this varies depending on the test element (Craik, 1992). The relevant Standard (ISO 140 Part
1) adopts the pragmatic solution of measuring the maximum achievable sound reduction index
for a limited number of test elements that are representative of those that are most commonly
measured.
3.5.1.3.2 Total loss factor
To assess the role of the total loss factor (Section 2.6.5) in the measurement of the sound
reduction index we now consider a non-porous, solid, homogeneous plate as the test element.
We start by assuming that this plate is only excited by the sound field in the source room.
Vibrational energy then leaves the test element via: (a) structural coupling into the laboratory
to which it is connected, (b) radiation losses into the rooms, and (c) internal losses into heat.
As the laboratory structure will be excited to some degree, we need to assume that there is
no significant flow of vibrational energy from the laboratory structure into the test element, and
that the test element is the dominant radiating element in the receiving room.
Sound transmission across the plate is determined by both non-resonant and resonant transmission; these terms are discussed in Section 4.3.1. At frequencies where the sound reduction
index of the plate is partly or wholly determined by resonant transmission, it is very useful to
measure the structural reverberation time of the plate in order to calculate the total loss factor
(Section 3.11.3.3). The reason for this is that resonant transmission is directly affected by the
plate damping, so it is possible to change the measured sound reduction index of a plate by
changing its total loss factor. This is particularly important for masonry/concrete plates where
resonant transmission is the dominant transmission mechanism over the majority of the building acoustics frequency range (see examples in Section 4.3.1.3). The simplest way to achieve
a change in the total loss factor is to change the connections at the plate boundaries from rigid
to resilient connections; for example from rigid mortar to a resilient material such as mineral
wool (Gösele, 1961; Kihlman and Nilsson, 1972). The total loss factor for a rigidly connected
plate will generally be higher than for a resiliently connected plate. If there is no significant flow
of vibrational energy from the laboratory structure into the test element and the resonant transmission dominates, the sound reduction index will be higher for the rigidly connected plate. In
practice there is an added complication because significant changes to the plate boundaries
(which usually fall into an idealized category of being simply supported, clamped, or free) also
change the plate modes and the radiation efficiency below the critical frequency. For a wall, the
lower plate boundary usually needs a rigid connection in order to provide structural stability.
This introduces an additional complication with orthotropic walls if there is a mixture of rigid
and resilient boundary conditions (see Section 4.3.2.2).
The structure of the test aperture varies from one transmission suite to another; usually it is
formed from dense concrete, blocks or bricks of various thicknesses. Therefore when one
masonry/concrete wall or floor is tested in different transmission suites, the coupling losses
to the laboratory structure can vary significantly (Craik, 1992; Kihlman, 1970). The resulting
variation in the total loss factor means that it is possible to measure different sound reduction
indices for a single test element in different laboratories (Gerretsen, 1990; Kihlman and Nilsson,
1972; Meier et al., 1999). One possibility would be to build identical transmission suites all over
the world; opinions on the optimum design would no doubt vary, and such uniformity would not
always be helpful for testing non-standard elements.
249
S o u n d
I n s u l a t i o n
Above the critical frequency of a non-porous, solid, homogeneous plate, there is a direct
relationship between the sound reduction index and the total loss factor of the plate; note that
if this plate is orthotropic, then it is the higher of its two critical frequencies. This fundamental
relationship can be seen from the equation for an infinite plate with mass, stiffness, and damping
(Eq. 4.61), or a three-subsystem SEA model for resonant transmission across a finite plate
that separates two rooms (Eq. 4.23). We consider the situation where this plate has been
measured in laboratory A and has a sound reduction index, RA , with a total loss factor, ηA . The
same plate is then measured in laboratory B where the total loss factor is ηB . Assuming that all
measurement errors are negligible, we can convert RA to the sound reduction index, RB , that
would be measured in laboratory B using
ηB
RB = RA + 10 lg
(3.51)
ηA
An example is shown in Fig. 3.15 for a masonry wall with a high mass per unit area and a low
critical frequency. In the mid- and high-frequency ranges, the conversion adequately accounts
for the difference between the measured sound reduction indices. It is not appropriate in the
Ns ⫽ 0.5
MA ⫽ 0.1
MB ⫽ 0.2
0.7
0.1
0.2
0.8 1.0 1.3
0.1 0.2 0.2
0.2 0.3 0.3
1.6
0.3
0.5
2.1 2.6
0.4 0.3
0.8 0.8
3.3 4.2
0.3 0.4
1.2 1.4
5.2 6.5 8.4
0.4 0.5 0.5
1.2 1.8 2.1
10 13
0.7 0.9
1.8 2.9
16
1.1
2.1
80
60
f11(A) ⫽ 70 Hz f11(B) ⫽ 78 Hz
50
fc ⫽ 94 Hz
fB(thin) ⫽ 814 Hz
112
110
108
30
106
104
102
20
100
100 125 160 200 250 315 400 500 630 800 1000 1250 1600 2000 2500 3150
40
Total loss factor (dB)
Sound reduction index (dB)
70
RA (measured)
RB (measured)
RB (converted from RA using the total loss factors)
ηA
ηB
One-third-octave-band centre frequency (Hz)
Figure 3.15
Converting the sound reduction index measured in laboratory A to laboratory B, where the measured total loss factor of
the test element is different in each laboratory. Test element: 215 mm masonry wall (solid blocks) with a 13 mm lightweight
plaster finish (each side). Wall areas are 9.3 and 8.6 m2 in laboratories A and B respectively. Upper x-axis labels show the
predicted statistical mode count (average value of both walls is shown) and the modal overlap factor for each wall A and B
in each frequency band. Sound insulation measurements according to ISO 15186 Part 1. Measured data from Hopkins are
reproduced with permission from ODPM and BRE.
250
Chapter 3
low-frequency range where the plate has very low modal overlap and only two or three bending
modes in the entire low-frequency range. These bending modes vary between the two different
sized walls; hence, the conversion is not able to account for the measured difference in the
sound reduction index.
Measurements on solid concrete block walls with a much lower mass per unit area and a
relatively high critical frequency (ρs = 41 kg/m2 , fc = 570 Hz) have been tested in five different
laboratories (Kihlman and Nilsson, 1972). The results indicated that above the critical frequency
the conversion could adequately account for the variation between these laboratories (a range
of ≈3 dB). An intercomparison study between 12 laboratories on an orthotropic solid block
wall (ρs = 440 kg/m2 , fc,x = 180 Hz, fc,y = 108 Hz) showed that the conversion was satisfactory
above the highest critical frequency (Meier et al., 1999). In these two studies there were a few
laboratories that were outliers in the converted data set. It is reasonable to assume that the
assumptions in the opening paragraph of this section do not apply to some laboratory designs.
Ignoring the few laboratories that are outliers, these studies indicate that the possible range
for the sound reduction index in one-third-octave-bands above the critical frequency is ≈5 dB
for rigidly connected masonry/concrete plates in different laboratories. These studies indicate
that this variation can be reduced to less than 2 dB in individual frequency bands if RA from
each laboratory is converted to RB using a reference loss factor, ηB . This reference loss factor
may be arbitrarily chosen, or related to a typical value for a certain type of wall or floor. Note
that Eq. 3.51 is a conversion, not a correction. We can measure a range of values for the same
wall or floor in different laboratories but all of these measured values are ‘correct’ under the
assumptions described in the opening paragraph of this section.
The conversion is valid above the critical frequency for solid homogeneous plates that are nonporous. A bonded surface finish (e.g. plaster) usually removes the non-resonant transmission
path across porous blocks and/or mortar joints thus allowing the wall or floor to be treated as
a non-porous plate. For fair-faced solid masonry walls the porosity is highly variable and the
conversion will not be appropriate in all cases. Another exception occurs at frequencies where
sound transmission is primarily determined by thickness resonances; this occurs with some
thick hollow brick/block walls. For such plates it will be the internal loss factor relating to the
plate thickness that primarily determines the sound reduction index (see Section 4.3.1.4). Note
that structural reverberation time measurements will quantify the total loss factor and this may
be dominated by coupling losses that have negligible effect on sound transmission for this type
of plate.
Below the critical frequency, the theory for airborne sound insulation of an infinite plate with
mass, stiffness, and damping implies that there will only be non-resonant (mass law) transmission, and that the total loss factor plays no role in airborne sound transmission. For many
solid masonry/concrete plates this is not the case; resonant transmission may dominate below,
at, and above the critical frequency (see the examples in Section 4.3.1.3). In order to carry
out the conversion below the critical frequency using Eq. 3.51, it is necessary to estimate
the non-resonant transmission (Section 4.3.1.2). If this is negligible compared to the resonant
transmission the conversion can be made in the same way. If not, the non-resonant component
of the sound reduction index needs to be removed so that the conversion can be carried out
on the resonant component. The non-resonant component can then be re-introduced after the
conversion.
There are other factors that need to be considered in applying the conversion below the critical frequency. This concerns the low-frequency range where repeatability and reproducibility
251
S o u n d
I n s u l a t i o n
values are highest, i.e. least favourable (see ISO 140 Part 2). With some laboratories it may only
be possible to make reliable conversions when intensity measurements are used to determine
the sound reduction index (ISO 15186 Parts 1 & 3). In addition, laboratories often have different
arrangements of side walls/floors. These walls/floors act as baffles to the test element; hence,
they can change the radiation efficiency of the plate below the critical frequency. Another
important factor in the low-frequency range is that the bending mode counts are usually very
low in one-third-octave or octave-bands with ≈10 m2 masonry/concrete plates. For this reason
the uncertainty in the measured structural reverberation time needs to be considered in any
conversion.
For solid walls and floors it is clear that the total loss factor is useful in interpreting the measured
sound reduction index, and applying the result to the field situation. However, there are a
number of issues that limit its widespread application. Originally the problem concerned the
difficulty in measuring very short reverberation times (Utley and Pope, 1973). These difficulties
have been overcome through signal processing techniques (Section 3.8.3.2.2), although there
is still the issue of increased measurement uncertainty with non-diffuse vibration fields. We now
need to consider the different types of masonry/concrete plates that are measured in practice.
To simplify design decisions it would be ideal if all masonry/concrete plates were non-porous,
solid, homogenous, isotropic, and reverberant with high modal densities. All measured values
for the sound reduction index could then be supplied with the measured total loss factor; ideally
along with the measured internal loss factor. In fact, if all plates were like this we would probably
decide not to measure at all; prediction would be quite adequate. In practice, the walls and
floors that we need to measure do not satisfy these ideal criteria. Take an example where we
need to know whether a fair-faced, solid blockwork wall (slightly porous) is likely to achieve
the same sound insulation as a hollow block wall with a plaster finish. A robust assessment
can only be made by measuring the sound reduction index, total loss factor, internal loss
factor, longitudinal wavespeed or bending stiffness, and airflow resistivity. Even if we do not
have prediction models for these specific walls, we can still use models of non-porous, solid,
homogeneous plates (Section 4.3.1) to gain some insight into the transmission mechanisms.
This will allow us to make a decision as to whether these two walls are likely to provide similar
sound insulation in situ. On this basis the total loss factor often proves useful in making design
decisions on many different types of solid masonry/concrete plate.
Measurement of the sound reduction improvement index is also affected by the total loss factor of the base wall or floor. However with linings there is the added complexity of another
non-resonant mechanism which gives mass–spring–mass resonances that sometimes affect
the transmission over several one-third-octave-bands. In addition, any change in the total
loss factor of the base wall due to addition of the lining is often within the limits of measurement uncertainty. Hence there is no simple conversion that is valid across the entire building
acoustics frequency range for all types of base walls and floors and all types of lining.
For cavity walls with masonry/concrete leaves, the variation in the total loss factor of each leaf
also accounts for some of the variation in test results between different laboratories. However,
there is no simple conversion. The reasons for this will become apparent in Section 4.3.5.2.2
which looks at different sound transmission paths across cavity walls.
From the above discussion it is clear that the main issue concerning total loss factors is that it
is not possible to use the conversion with all types of walls and floors. For those types where it
can theoretically be used at all frequencies, it may not be possible to accurately measure the
structural reverberation time at all frequencies. For these reasons it is not particularly useful
252
Chapter 3
to convert a limited number of measurements to a reference total loss factor. However, any
assessment of a laboratory measurement for use in situ always requires consideration of the
in situ total loss factor and flanking transmission. Therefore the most helpful approach is to
simply provide the measured total loss factor alongside the measured sound reduction index
in the laboratory test report. At the very least this allows engineering judgement to be used in
converting laboratory measurements to the in situ situation; even if it is only possible over part
of the building acoustics frequency range.
For the impact sound insulation of solid homogeneous plates, there is only resonant transmission and the normalized impact sound pressure level, Ln(A) can be converted to Ln(B) , using
ηB
Ln(B) = Ln(A) − 10 lg
(3.52)
ηA
3.5.1.3.3
Niche effect
The transmission suite aperture, or the aperture of a filler wall which contains the test element,
is usually much thicker than the test element itself. This is usually necessary to suppress
flanking transmission via the laboratory structure or direct transmission across the filler wall.
The aperture therefore forms a niche on one or both sides of the test element, depending on
where the element is positioned within the aperture. Unless a specific niche arrangement is to
be tested, the Standards (ISO 140 Parts 1 & 3) describe an arrangement based on different
niche depths on each side of the element with a ratio of 2:1, and niche boundaries with an
absorption coefficient <0.1 over the building acoustics frequency range.
The reason for carefully specifying and reporting the niche arrangement used in the laboratory is
that the position of the element within the aperture can significantly change the sound reduction
index. This is often referred to as the ‘niche effect’. Although this name implies a single effect
that is linked purely to the niche geometry, it depends upon both the test element and the niche
on both sides; this results in a frequency-dependent effect that varies between test elements.
The following discussion is based around the niche effect in laboratory measurements, but the
effect is equally important in situ. We will assume that the niche has rigid boundaries and that
the test element is installed such that the plates or beams forming the niche do not introduce
a flanking path or radiate sound themselves.
As there are a number of variables that affect the measured sound reduction index of a test
element within a niche, it is quite possible to come across contradictory measurements concerning the niche effect. For this reason we will consider individual factors that contribute to the
niche effect whilst acknowledging that all factors need to be considered together. It is difficult
to gain a practical insight into the niche effect by considering only one theoretical model. This
is to our benefit as the following discussion will provide a useful backdrop to the application of
finite and infinite plate models that will be discussed in Section 4.3.1. To assess the niche effect
it is simplest to assume that the test element is a solid plate; we can therefore consider sound
transmission in terms of non-resonant and resonant transmission paths. The examples in Section 4.3.1.3 can be used to gain an impression of the importance of non-resonant transmission
relative to resonant transmission below the critical frequency of a plate.
The factors contributing to the niche effect can be considered as follows:
(1) Shielding of the test element surface from sound waves that impinge upon the
element at near-grazing angles of incidence.
253
S o u n d
I n s u l a t i o n
dn
Source
room
θmax
Lx
or
Ly
Receiving
room
Laboratory
structure
or
filler wall
Figure 3.16
Test element mounted in a niche (cross-section).
Tests in different laboratories generally show that the effect of shielding is most
significant at and below the critical frequency of the plate facing into the niche (Guy
and Sauer, 1984; Kihlman and Nilsson, 1972).
The angle of incidence for an incident plane wave is measured from a line perpendicular to the surface of the test element. When the test element is mounted in a
niche as shown in Fig. 3.16, the maximum angle of incidence, θmax , that can impinge
upon the centre point of the element is given by θmax = arctan(L/2dn ), where dn is
the niche depth and L is either Lx or Ly . If the test element is a full-size wall with
dimensions, 4 × 2.5 m and a niche depth of 25 mm facing into the source room,
the average θmax is 89◦ and the shielding effect is negligible. However, for a sheet
of glass with dimensions, 1.5 × 1.25 m and a niche depth of 150 mm, the average
θmax is 78◦ . At this point we can make a link to the prediction of non-resonant transmission based on an infinite plate acting as a limp mass. This is often calculated by
assuming that all angles of incidence are equally probable between 0◦ and 78◦ , and
is referred to as the field incidence mass law. For an infinite plate, non-resonant
transmission is highly dependent upon the angle of incidence with near-grazing
angles being very efficient at transmitting sound (refer forward to Fig. 4.6). The
infinite plate model indicates that shielding of near-grazing angles by the niche will
significantly change the sound reduction index due to non-resonant transmission
below the critical frequency.
This approach is too simplistic to fully explain the shielding effect. We have ignored
the two-dimensional reverberant sound field within each niche, as well as the
effects of diffraction and scattering at the niche edges. Despite this, it is still useful
254
Chapter 3
40
θmax
Plate dimensions
Sound reduction index (dB)
≈55°
≈69°
≈72°
0.75 ⫻ 0.75 m
1.6 ⫻ 1.6 m
2⫻2m
35
fc ⫽ 2856 Hz
30
225 mm 55 mm
25
θmax
20
Source room
Receiving room
45°
15
50
80
125
200
315
500
800
1250
2000
3150
5000
One-third-octave-band centre frequency (Hz)
Figure 3.17
Measured airborne sound insulation for different size sheets of 6.25 mm glass within a niche that gives different degrees of
shielding. Measurements according to ASTM E90-75. Measured data are reproduced with permission from Guy and Sauer
(1984).
because it makes a link to infinite plate models that are often used to predict
non-resonant (mass law) transmission; it also shows that there is little basis for
assuming that all angles of incidence are between 0◦ and 78◦ for any size of plate
with any depth of niche. For finite plates, non-resonant transmission is dependent upon the plate dimensions. The non-resonant sound reduction index tends
to increase with decreasing plate size (Section 4.3.1.2.2). Hence for relatively
small elements within a relatively deep niche, such as glazing or windows, it is
the finite size of the plate as well as the shielding of near-grazing angles that
affects non-resonant transmission below the critical frequency.
In the vicinity of the critical frequency, the absence of near-grazing angles in the
incident sound field due to shielding would be expected to have an effect on the
initial part of the coincidence dip. Using the infinite plate model for a plate with mass,
stiffness, and damping shows that the coincidence dip for field incidence is slightly
higher than for diffuse incidence (refer forward to Fig. 4.17). Measurements on
different sizes of glass in Fig. 3.17 also show that the coincidence dip tends to shift
upwards in frequency as θmax decreases (Guy and Sauer, 1984; Guy et al., 1985).
Shielding may partly account for the fact that using frequency-average radiation
efficiency formulae tends to predict deeper coincidence dips in the sound reduction
index than are measured in the laboratory (Leppington et al., 1987).
(2) Orientation of the baffle that surrounds the plate.
Plates that form building elements usually have a standard baffle orientation that
can be described as a baffle in the same plane as the plate, or a baffle around
the plate perimeter that is orientated perpendicular to the plate surface. Laboratory tests on elements such as glazing, windows, doors, skylights, or infill-panels
are usually carried out by fixing the element in a filler wall with a niche on one
or both sides. The baffle formed by the niche and the filler wall may or may not
255
S o u n d
(a)
I n s u l a t i o n
(b)
(c)
(d)
Figure 3.18
Examples of different types of niche used to mount test elements. (a) box-shaped niches in a solid wall, (b) tapered niches
in a solid wall, (c) niche with one step or stagger in a cavity wall, and (d) niche with more than one step or stagger in a
cavity wall.
be representative of the element when installed in a building. When flush with the
surface of the filler wall, these elements effectively lie within the plane of a baffle
formed by the filler wall. When positioned within a niche, the baffle orientation is no
longer simple to describe; it depends upon the niche geometry as well as sound
radiation from individual bending modes. The niche geometry may be more complex than a simple box-shaped cavity; it can have sloped, stepped, or staggered
boundaries (see Fig. 3.18).
Below the critical frequency, the frequency-average radiation efficiency for bending
modes on a finite plate depends upon the orientation of the baffle around the plate
perimeter (Section 2.9.4). Compared to a plate that lies within the plane of an infinite rigid baffle, the radiation efficiency below the critical frequency doubles when
rigid baffles are placed perpendicular to the plate perimeter (Leppington, 1996).
Most niche geometries could be described by one of these two idealized baffle
orientations, although part of the frequency range will inevitably fall somewhere
between the two. We now recall the discussion in Section 2.9.3 in which individual bending modes on a simply supported plate are classified as either corner or
edge radiators below the critical frequency; sound radiation from the edge radiators
tends to dominate. A crossover frequency between the two idealized baffle orientations can be estimated by describing the niche depth in terms of a fraction of a
wavelength. Hence we could assume that niche depths <λ/2 can be modelled as if
the plate lies in the plane of an infinite rigid baffle, and that with niche depths ≥λ/2,
the plate can be modelled as if there were rigid baffles perpendicular to the plate
perimeter.
In the laboratory, the depth of a box-shaped niche is typically in the range,
50 ≤ dn ≤ 200 mm. A 50 mm niche depth corresponds to λ/2 at 3430 Hz. Most
plates used in buildings (except thin metal sheets) have critical frequencies below
this frequency; hence resonant transmission below the critical frequency will be
similar whether the plate is flush with the filler wall or within a 50 mm deep niche.
256
Chapter 3
A 200 mm niche depth corresponds to λ/2 at 858 Hz; hence for plates such as
glass or plasterboard (which typically have critical frequencies above 2000 Hz) the
niche is likely to increase the radiation efficiency of modes just below the critical
frequency compared to when the plate is flush with the filler wall. This may or may
not be detectable in the sound reduction index because it depends upon the relative strengths of non-resonant transmission and resonant transmission at these
frequencies. This will be specific to the test element and its fixing at the boundaries
(e.g. simply supported, clamped) as well as the effect that the fixing has on the
coupling losses from the element.
(3) Two-dimensional modal sound field within the niche.
As the niche depth increases, the incident sound field on the test element is distinctly different to when it is flush with a room surface. Although a niche opens
out directly into the three-dimensional sound field of the room, we can find an
explanation for some aspects of the niche effect by considering the existence of
a two-dimensional modal sound field within the niche. When the niche on each
side of the test element has identical dimensions, then the niche modes and the
modes of the test element are likely to be strongly coupled. This links to the observation that airborne sound transmission in the low-frequency range tends to be
lower between rooms with identical dimensions, than between rooms with different dimensions (Heckl and Seifert, 1958). Deep niches with identical dimensions
on each side tend to reduce the measured sound reduction index at frequencies
above, at and below the critical frequency of the test element.
The effect of equal niche depths is most commonly observed below the critical
frequency (Cops et al., 1987; Guy and Sauer, 1984). An example is shown
in Fig. 3.19a where the sound reduction index decreases, as the niche depth
increases (Yoshimura, 2006). The effects of strong coupling between each niche
and the test element can be reduced or avoided by using different niche dimensions on both sides. This can be done by introducing a step and/or stagger into the
test aperture, or by using small diffusing elements around the niche boundaries to
alter the niche modes in one or both of the niches.
The adverse effects of a deep niche on only one side of the test element (source or
receiving room) are less pronounced than with equal niche depths. However, they
can still affect important features of the sound reduction index such as the depth of
the mass–spring–mass resonance dip as shown in Fig. 3.19b (Yoshimura, 2006).
(4) Absorption of sound energy by the niche boundaries.
Whether we consider transmitted waves at specific angles, or a two-dimensional
modal sound field in each niche, it is clear that absorptive surfaces within the niche
should increase the sound reduction index. Depending on the niche geometry
and the absorber characteristics, this can occur below, at and above the critical
frequency (Guy and Mulholland, 1979).
(5) Changes to the structural coupling losses from the test element to the laboratory
structure (or filler wall) that forms the test aperture.
This affects resonant transmission below, at and above the critical frequency and
can be significant when it is the structural coupling loss factors that primarily determine the total loss factor of the test element. Different positions of the test element
within the niche may change the structural coupling losses. Any change can be
assessed (and sometimes accounted for) by measuring the total loss factor of the
test element.
257
S o u n d
I n s u l a t i o n
(a) Equal niche depths
50
39 mm 39 mm
fc ⫽ 2495 Hz
fmsm ⫽ 219 Hz
Sound reduction index (dB)
45
40
100 mm 100 mm
100 mm 100 mm
35
Source room
Receiving room
30
100 mm
25
39 mm : 39 mm
139 mm : 139 mm
239 mm : 239 mm
20
15
50
80
125
200
315
500
800
1250
2000
3150
5000
One-third-octave-band centre frequency (Hz)
(b) Equal and different niche depths
50
39 mm 39 mm
fc ⫽ 2495 Hz
fmsm ⫽ 219 Hz
Sound reduction index (dB)
45
40
100 mm 100 mm
35
Source room
Receiving room
30
100 mm
25
39 mm : 39 mm
39 mm : 139 mm
39 mm : 239 mm
20
15
50
80
125
200
315
500
800
1250
2000
One-third-octave-band centre frequency (Hz)
3150
5000
Figure 3.19
Effect of a box-shaped niche on both sides of a 5-12-5 insulating glass unit (1.89 × 1.54 m) with equal and different niche
depths. The test aperture is formed by a 100 mm concrete plate, with deeper niches formed using sheets of 12 mm
plasterboard. Measurements according to ISO 140 Part 3. Measured data are reproduced with permission from Yoshimura
(2006).
3.5.2 Field measurements within buildings
For airborne sound insulation in the field, sound is not only transmitted by the separating
wall or floor, but also by flanking transmission. To calculate a sound reduction index for field
measurements it is therefore necessary to define a transmission coefficient, τ ′ , that includes
258
Chapter 3
the sound power, W3 , transmitted by flanking elements into the receiving room,
τ′ =
W2 + W 3
W1
The apparent sound reduction index, R ′ , is therefore defined as
1
S
W1
R ′ = 10 lg ′ = 10 lg
= Lp1 − Lp2 + 10 lg
τ
W2 + W3
A
(3.53)
(3.54)
where S is the area of the separating element. In the field there may be a step or a stagger
between the rooms, therefore S is the area that is common to both the source and receiving
rooms.
Note that whilst R ′ is primarily intended for field measurements, it is sometimes necessary to
quote laboratory test results using R ′ rather than R when laboratory measurements are known
to be significantly affected by flanking transmission.
For lightweight separating elements, the presence of poor workmanship and/or flanking transmission may be indicated by values of R ′ from field tests that are lower than R for the separating
element. However, this difference may also be due to structural coupling around the perimeter
of the wall or floor that is different in situ compared to the laboratory (see Section 4.3.5.4.4).
For heavyweight constructions, any comparison of R ′ and R that is used to try and identify poor
workmanship needs to take account of the total loss factor of the separating element (Section
3.5.1.3.2) as well as flanking transmission (Chapter 5).
For field measurements the airborne sound insulation can be described in terms of the sound
pressure level difference, D, between the source and receiving rooms. This can cause problems when setting sound insulation requirements for regulatory purposes, because adding or
removing sound absorptive material from the receiving room will change the measured sound
pressure level, and hence change the level difference. In some situations the reverberation
time in the receiving room may be fixed by other requirements and it may be appropriate just
to use the level difference. Otherwise it is necessary to measure the reverberation time in the
receiving room and to ‘standardize’ or ‘normalize’ the level difference. This provides a fairer
basis on which to set performance standards for sound insulation. The level difference, D, is
‘standardized’ using a reference value for the reverberation time and the level difference is
‘normalized’ using a reference value for the absorption area.
For a given sound power that is transmitted into the receiving room, the mean-square sound
pressure in this room is inversely proportional to the absorption area, A, of that room (Eq. 3.36).
On this basis the normalized level difference, Dn , is defined by using a reference absorption
area, A0 , of 10 m2 for the receiving room (ISO 140 Part 4)
A
A
Dn = Lp1 − Lp2 + 10 lg
(3.55)
= D − 10 lg
A0
A0
A similar approach is used to give the standardized level difference. For a given sound power
transmitted into the receiving room, the mean-square sound pressure in this room is proportional to the reverberation time, T , of that room. The standardized level difference, DnT , is
defined by using a reference reverberation time, T0 , for the receiving room, which for dwellings
is 0.5 s (ISO 140 Part 4)
T
T
DnT = Lp1 − Lp2 − 10 lg
(3.56)
= D + 10 lg
T0
T0
259
S o u n d
I n s u l a t i o n
Note that from Sabine’s equation (Eq. 1.97) the relationship between T and A only involves
the receiving room volume; hence Dn and DnT will be the same when the volume is 31 m3 .
Regulatory requirements for dwellings are often set using single-number quantities that are
calculated with the rating method in ISO 717 Part 1; other descriptors that are specific to
individual countries are not discussed here. The choice of single-number quantity depends on
what the regulation is aiming to achieve and how the regulation is implemented and enforced.
For this reason different countries may need to use different single-number quantities, although
the requirements are usually based around DnT or R ′ . Dn uses a reference absorption area
which is not usually appropriate because of the wide range of room dimensions, for which there
will be a wide range of absorption areas. R ′ uses the common area of the separating element;
this can be awkward if there is no common area between the rooms (e.g. diagonally offset
rooms). DnT is based on a reference reverberation time; however, typical reverberation times in
dwellings tend to vary between countries. They vary depending on the preferred type of surface
finish (e.g. carpet, ceramic tiles) and whether the room is furnished. For furnished rooms
(excluding kitchens and bathrooms) with volumes between 15 and 60 m3 , the reverberation
time over the building acoustics frequency range typically lies between 0.4 and 0.6 s (European
data from Vorländer, 1995). It is therefore reasonable to use T0 = 0.5 s for furnished dwellings
where the reverberation time tends to be independent of the room volume. This is because the
absorption area tends to be proportional to the volume; i.e. larger rooms usually have more
absorbent furnishings (van den Eijk, 1972). When testing between rooms of unequal volume
this means that the measured DnT will depend on the choice of source and receiving room.
If the receiving room is the larger room, Lp2 will be lower than if the smaller room had been
chosen; this is due to the higher absorption area in the larger room. Therefore when carrying
out testing in a single direction to assess compliance with regulations it is common to require
that the smaller room is used as the receiving room so that the lowest DnT values are measured.
When all sound is transmitted between two rooms via a separating element with a sound
reduction index, R, the relationship between DnT and R is
V
− 5 dB
(3.57)
DnT = R + 10 lg
S
which for box-shaped receiving rooms simplifies to
DnT = R + 10 lg d − 5 dB
(3.58)
where V is the volume of the receiving room and d is the receiving room dimension that is
perpendicular to the separating element. Note that DnT = R when d = 3.2 m.
Whilst DnT has some advantages in setting regulatory requirements, Eq. 3.58 indicates that
some caution is needed when a data set of field sound insulation tests is used to calculate
pass and failure statistics for a certain type of construction. The data set may be biased if all
the receiving rooms have very low or very high values of d.
3.5.2.1
Reverberation time
For measurements in diffuse fields it is appropriate to use T30 for the reverberation time. In typical rooms the building acoustics frequency range often covers sound fields that are non-diffuse
as well as those which can be considered to be diffuse. Previous discussions on reverberation
time (Section 1.2.6.3) show that non-diffuse fields can lead to curvature of the decay curve or
260
Chapter 3
distinct double slopes. To make the link between sound power and reverberant sound pressure
level in non-diffuse fields it is therefore necessary to use T10 , T15 , or T20 (Section 1.2.7.5.2).
In practice, a balance can usually be struck by using T20 for all of the building acoustics
frequency range.
3.5.2.2
Sound intensity
Sound intensity can also be used in the field to measure the sound insulation of a building
element in the presence of flanking transmission (ISO 15186 Part 2). Examples of relevant
building elements inside buildings include separating walls, flanking walls, doors, or small
building elements such as cable ducts.
For the sound power radiated by a separating element, the apparent intensity sound reduction
index, RI′ , is defined as
SM
RI′ = Lp1 − LIn − 10 lg
− 6 dB
(3.59)
S
where Lp1 is the average sound pressure level in the source room, LIn is the average sound
intensity level over the measurement surface for the separating element, SM is the total area
of the measurement surface, and S is the area of the separating element. When there is a step
or a stagger between the source and receiving rooms, S is the area that is common to both
rooms.
In the measurement of RI′ , the intensity probe is used to measure only the sound power radiated
by the separating element. Hence there is an important difference between RI′ and R ′ . With R ′
(ISO 140 Part 4) the sound power radiated by flanking elements is included, but with RI′ (ISO
15186 Part 2) it is not included.
′
For the sound power radiated by flanking element j, the intensity sound reduction index, RIF,
j,
is defined as
SM, j
′
RIF,
=
L
−
L
−
10
lg
− 6 dB
(3.60)
p1
In, j
j
S
where LIn is the average sound intensity level over the measurement surface for flanking element j, SM is the total area of the measurement surface for flanking element j, and S is the
area of the separating element. When there is a step or a stagger between the source and
receiving rooms, S is the area that is common to both rooms.
The intensity normalized level difference, DI,n , is intended for use when there is no common
surface area in the source and receiving rooms. This can occur with rooms that are sited
diagonally across from each other, or with rooms that are separated from each other by another
room. In this case one specific element is chosen for the intensity measurement.
SM
DI,n = Lp1 − LIn − 10 lg
(3.61)
− 6 dB
A0
where LIn is the average sound intensity level over the measurement surface for a chosen
element, and SM is the total area of the measurement surface.
For small building elements or devices, the intensity element-normalized level difference, DI,n,e ,
can also be used in the field situation and is calculated in the same way as in the laboratory.
SM
DI,n,e = Lp1 − LIn − 10 lg
(3.62)
− 6 dB
A0
261
S o u n d
I n s u l a t i o n
where LIn is the average sound intensity level over the measurement surface for the small
building element or device and SM is the total area of the measurement surface.
3.5.3 Field measurements of building façades
Airborne sound insulation measurement of building façades are categorized according to the
relevant Standard (ISO 140 Part 5). The measurement is either used to determine the sound
insulation of a single building element (e.g. window, door) or the entire façade.
3.5.3.1 Sound insulation of building elements
Measurements on a building element can either be carried out using a loudspeaker, or the
existing environmental noise source such as road, railway, or aircraft traffic.
3.5.3.1.1
Loudspeaker method
The apparent sound reduction index of an individual building element, such as a window, can be
measured from outside to inside using sound pressure level measurements with a loudspeaker
facing towards the façade. Note that we can only measure the apparent sound reduction index
because there will inevitably be some sound transmitted by the rest of the façade, which we
regard as flanking transmission. The transmission coefficient is defined by Eq. 3.53. However
the incident sound power, W1 , is calculated differently; this is because the incident sound field
on the element is different to that from within a room.
From Section 1.4.1 we recall that with a point source and sound pressure measurements on a
surface, pressure doubling can be assumed compared to the free-field value, and that interference dips can be avoided in the building acoustics frequency range. Hence, it is convenient to
measure the average surface sound pressure level, Lp1,s , on the building element under test,
and the average sound pressure level in the receiving room, Lp2 .
By placing the loudspeaker at a sufficient distance from the test element we can assume that
the incident sound field is comprised of plane waves. For measurements according to the
relevant Standard (ISO 140 Part 5) the loudspeaker is positioned at an angle, θ, of 45◦ ; θ is the
angle between the line normal to the centre of the test element, and the axis of the loudspeaker
that points towards the centre of the test element (see Fig. 3.20). The component of the plane
wave intensity that is incident upon the test element is I cos θ. The incident sound power, W1 ,
3 4
can be calculated from the average free-field mean-square sound pressure, p12 t,s , that would
exist at the test element position if the test element and the rest of the façade were not there
by using
3 24
p1 t,s
◦
W1 = SI1 cos 45 =
(3.63)
S cos 45◦
ρ0 c0
where the spatial average corresponds to the area that would be occupied by the test element
if it were present.
By assuming pressure doubling at the surface we can calculate W1 from the measurement of
3 2 4
the average surface mean-square sound pressure, p1,s
, using
t,s
3 2 4
1 p1,s t,s
W1 =
S cos 45◦
(3.64)
4 ρ0 c 0
262
Chapter 3
Test
element
h
θ
r
hs
d
S
Figure 3.20
Loudspeaker position (S) relative to the centre of the test element.
where the spatial average corresponds to measurements over the surface area of the test
element.
The power transmitted into the receiving room is calculated from Eq. 3.36, hence the apparent
′
sound reduction index, R45
◦ , is
′
R45
◦
= 10 lg
2
p1,s
t,s
p22 t,s
+10 lg
S
S
+10 lg( cos 45◦ ) = Lp1,s −Lp2 +10 lg
−1.5 dB (3.65)
A
A
where Lp1,s is the average sound pressure level measured on the surface of the façade and
Lp2 is the average sound pressure level in the receiving room.
Angle of incident sound: Airborne sound insulation is dependent upon the angle of the
incident sound (Section 4.3.1.5); hence differences are to be expected between the sound
reduction index measured with a diffuse incidence sound field and with sound incident at a
single angle. There is no general rule for conversion between the two types of incident sound;
this is obvious if one considers a single sheet of glass, a side-hung window opened at a narrow
angle, and a through-wall ventilator. However, for closed windows, θ = 45◦ often gives a reasonable estimate of the sound reduction index with a diffuse incidence sound field. For closed
windows, there is some evidence that the apparent sound reduction index measured using
θ = 60◦ instead of 45◦ , gives closer agreement with R measured in the laboratory (Jonasson
and Carlsson, 1986). However no single angle will give exact agreement over the building
acoustics frequency range and it is necessary to standardize a single angle. It is also difficult
to justify a change to 60◦ when for practical purposes it is simpler to position a loudspeaker
at an angle of 45◦ , the sound insulation of different façade elements will have different angle
263
S o u n d
I n s u l a t i o n
dependence, and there are other variables such as loudspeaker height that also affect the
measurement. The relevant measurement Standard therefore uses θ = 45◦ (ISO 140 Part 5).
Uniformity of incident sound field: Derivation of the incident sound power (Eq. 3.63) is based
around the free-field mean-square sound pressure at the test element position. Ideally there
would be a perfectly uniform incident sound field over the entire surface of the test element. In
practice this is not possible to achieve. This is only partly due to the directivity of the loudspeaker
because this aspect is controlled by the loudspeaker specifications in the Standard. The spatial
variation over the surface of the element is primarily determined by the combination of ground
impedance, element dimensions, and the source-receiver-façade geometry. The latter is constrained by the requirements of the Standard and the practical restrictions on site that limit the
available positions for a loudspeaker. The Standard requires that the loudspeaker height, hs ,
and distance from the façade, d, are chosen so that the spatial variation over the surface of the
element is minimized. To do this it requires that the loudspeaker should preferably be placed
on the ground or as high above the ground as is practical. For a loudspeaker placed on the
ground, the acoustic centre of the loudspeaker will typically be at a height, hs ≈ 0.5 m.
The effect of different loudspeaker heights has been investigated experimentally by Jonasson
and Carlsson (1986). Their measurements looked at the spatial variation of the free-field sound
pressure level over a 2 × 2 m area using nine microphones. Two different loudspeaker heights
were used, one on the ground and the other raised by 2.9 m. The distance from the loudspeaker
to the centre of the 2 × 2 m area was 5 m; this is the minimum distance required by the Standard
when measuring a façade element. For windows, typical mid-height values above the ground
are approximately 1.5 m for the ground floor, and 4 m for the first floor; in these measurements
the mid-point of the 2 × 2 m area was at a height of 1.6 or 4.2 m. Figure 3.21 shows the maximum
variation in the sound pressure level over the 2 × 2 m area. For a ground floor window, the
spatial variation in the low-frequency range is significantly lower with the loudspeaker on the
ground than when raised off the ground. The sound pressure at any point in the free field is
the combination of the direct path and the path involving a single reflection from the ground. For
the ground floor window, the first interference dip occurs in the mid-frequency range when the
loudspeaker is on the ground, but in the low-frequency range when the loudspeaker is raised
off the ground. The latter interference dip is deeper than the former, and significantly increases
the spatial variation over the 2 × 2 m area. In these experiments the ground cover was short
grass; however, the frequency and depth of the interference dips will vary depending on the
ground impedance.
3.5.3.1.2 Sound intensity
The apparent sound reduction index of an individual building element in the façade, such as a
window or door can also be measured using sound intensity. In this situation the measurements
are taken with the loudspeaker inside the room and the test element is scanned with the intensity
probe from outside the room to give the apparent intensity sound reduction index, RI′ ,
RI′
S
= Lp1 − LIn + 10 lg
SM
− 6 dB
(3.66)
Note that we refer to this sound reduction index as ‘apparent’ to indicate that it is a field
measurement, although the intensity probe is used to suppress measurement of sound that is
radiated by the flanking elements.
264
Chapter 3
16
(1) Ground floor window,
loudspeaker on ground
Maximum level difference between
any two positions in the grid (dB)
14
(2) Ground floor window,
loudspeaker raised off the ground
12
(2) First floor window,
loudspeaker raised off the ground
10
8
6
4
2
0
50
80
125
200
315
500
800
1250
2000
3150
5000
One-third-octave-band centre frequency (Hz)
2 ⫻ 2 m grid with
nine microphone
positions
45°
h
5m
hs
Ground: short grass
S
Figure 3.21
Maximum spatial variation of the free-field sound pressure level measured over a 2 × 2 m area. The key describes the
situation to which the measurements correspond, the heights are: (1) hs = 0.3 m, h = 1.6 m (2) hs = 2.9 m, h = 1.6 m and
(3) hs = 2.9 m, h = 4.2 m. Measured data are reproduced with permission from Jonasson and Carlsson (1985).
For small building elements or devices, such as a ventilator in the external wall, the intensity
element-normalized level difference, DI,n,e , is
SM
DI,n,e = Lp1 − LIn − 10 lg
(3.67)
− 6 dB
A0
where LIn is the average sound intensity level over the measurement surface for the small
building element or device, and SM is the total area of the measurement surface.
265
S o u n d
3.5.3.1.3
I n s u l a t i o n
Road traffic method
The apparent sound reduction index of an individual building element can also be measured
from outside to inside using road traffic noise as the sound source. Requirements on the
geometry of the road traffic source in relation to the façade and the number of passing vehicles
during the measurement period are described in the Standard. The noise level and spectrum will
depend on the geometry of the situation as well as traffic flow, traffic speed, traffic composition
in terms of the type of vehicles, road gradient, and road surface. Compared to loudspeaker
measurements it is difficult to estimate the background noise level in the receiving room from
measurements unless the traffic noise can be stopped. This means that background noise
corrections are rarely possible.
For road traffic along streets that are lined with buildings on both sides, the external sound field
begins to resemble a reverberant field. However, sound is not usually incident from above the
building element unless there are features, such as balconies, which reflect or scatter sound
down onto the element from above. In general, wherever a road runs approximately parallel
to the façade it is reasonable to assume that sound will be incident upon the building element
from a wide range of angles.
As with loudspeaker measurements, the external sound pressure level is measured on the
surface of the element. In Section 1.2.7.1 we looked at sound fields near room boundaries
and the interference patterns that exist near surfaces in rooms. For road traffic, the theory for
sound incident upon a surface from all directions is applied, hence sound pressure levels at
the surface are assumed to be 3 dB higher than at a distance far away from the surface. It is
therefore necessary to subtract 3 dB from the surface sound pressure level, Lp1,s , to modify
′
Eq. 3.54; this gives the apparent sound reduction index for a road traffic noise source, Rtr,s
, as:
S
′
Rtr,s
= Lp1,s − Lp2 + 10 lg
− 3 dB
(3.68)
A
3.5.3.1.4
Aircraft and railway noise
The apparent sound reduction index can also be measured using aircraft or railway traffic
′
′
as the sound source, giving Rat,s
or Rrt,s
respectively. In the same way as with loudspeaker
measurements, the external sound pressure level is measured on the surface of the element.
In the context of sound insulation, aircraft noise events are particularly complex. During an
aircraft flyover, the sound insulation provided by the test element varies with time. This is due
to the time-varying incident sound field, as well as the sound insulation of the element varying
with the angle of incidence. As an aircraft travels along a certain flight path, the source–
receiver–façade geometry changes, which therefore changes the incident sound field and the
angle of incidence. Sound insulation is often measured in buildings quite close to airports,
and there is usually more than one possible flight path for take-off and landing. In addition,
the sound emitted from an aircraft can be highly directional and its level and spectrum will
vary depending on the type of aircraft as well as whether it is taking-off, cruising, or landing.
A further complication is the fact that we can only measure the apparent sound reduction index.
It is difficult to confirm that the dominant sound transmission path throughout the duration of
the noise event occurs via the element under test, such as a window. As the aircraft changes
its orientation to the test element and the rest of the building, flanking paths such as those
involving the roof may become dominant for some or all of the duration of the noise event.
For the above reasons, comparison of R from the laboratory with a single measurement of
′
Rat,s
in the field is rarely meaningful.
266
Chapter 3
In contrast to loudspeaker and road traffic noise measurements, both aircraft and railway noise
consist of individual noise events. Therefore, the sound pressure inside and outside the building
is measured in terms of the sound exposure level, LE . During the noise event it is assumed that
the sound will be incident upon the test element from the majority of directions. Therefore the
apparent sound reduction index for railway or aircraft noise is calculated using the same equation as for road traffic noise (Eq. 3.68), but by using LE for the sound pressure levels. For the
mean of many measurements, the assumption of energy doubling for the surface sound pressure measurement is a reasonable estimate in the mid-frequency range, but for aircraft there
can be significant differences in the low- and high-frequency ranges (Bradley and Chu, 2002).
3.5.3.2 Sound insulation of façades
We now look at measuring the sound insulation of an entire façade, rather than a single
building element. The apparent sound reduction index of a façade can be measured from
outside to inside, either by using a loudspeaker, or by using the existing environmental noise
source. The preferred method is to use the existing environmental noise source because the
sound insulation is usually being measured to assess the ability of the façade to reduce the
environmental noise level inside the building. In addition, the external sound field associated
with environmental noise may be too complex to adequately reproduce with a loudspeaker.
For example, the sound incident upon one part of the façade can be significantly different
to another part, such as with aircraft flyovers where the sound that is incident on a window
in a side wall can be significantly different to that which is incident on the roof. However, if
the environmental noise source is not sufficiently loud, or the sound insulation is particularly
high, then it may not be possible to measure levels in the receiving room that are above the
background noise level, and the appropriate solution is to use a loudspeaker.
For the entire façade it is not practical to measure the spatial average surface sound pressure
level over every single façade element. In addition, the spatial average becomes less meaningful when there are large variations in level across the façade, particularly when façades contain
elements with significantly different levels of sound insulation (e.g. a thick masonry wall and
a single pane of glass). The approach that is adopted in the relevant measurement Standard
(ISO 140 Part 5) is to measure the external sound pressure level at the middle of the façade at
a height of 1.5 m above the floor of the receiving room, and at a distance of 2 m perpendicular to
the plane of the façade. For this scenario we have seen in Section 1.4.1 that there will be interference between the different sound propagation paths from the source to the receiver in the
low-frequency range. The interference patterns are affected by the orientation of the source,
receiver, and the façade, as well as by the ground and façade impedances. These patterns will
therefore vary from site to site. In addition we have seen that in the mid- and high-frequency
ranges, the assumption of energy doubling at a distance of 2 m from the façade is reasonable for
the mean of many measurements; but not for any individual measurement. For these reasons
the incident sound power cannot be accurately calculated from sound pressure level measurements made at a distance of 2 m from the façade. The external sound pressure level can
therefore be treated purely as a reference level that is needed to calculate the level difference.
For loudspeaker or road traffic noise measurements, the level difference, D2m , is
D2m = Lp1,2m − Lp2
(3.69)
where Lp1,2m is the sound pressure level measured 2 m in front of the façade.
267
S o u n d
I n s u l a t i o n
For railway or aircraft traffic measurements, the level difference, DE2m , is
DE2m = LE1,2m − LE2
(3.70)
where LE1,2m is the sound exposure level measured 2 m in front of the façade.
As with other field sound insulation measurements the reverberation time in the receiving room
is measured in order to replace the actual room damping by a reference value so that the level
difference can be ‘normalized’ or ‘standardized’.
For loudspeaker or road traffic noise measurements the normalized level difference, D2m,n , is
defined by using a reference absorption area, A0 , of 10 m2 for the receiving room (ISO 140
Part 5).
A
D2m,n = D2m − 10 lg
(3.71)
A0
For railway or aircraft traffic measurements, the normalized sound exposure level difference,
DE2m,n , is
A
DE2m,n = DE2m − 10 lg
(3.72)
A0
For loudspeaker or road traffic noise measurements the standardized level difference, D2m,nT , is
defined by using a reference reverberation time, T0 , for the receiving room, which for dwellings
is 0.5 s (ISO 140 Part 5).
T
D2m,nT = D2m + 10 lg
(3.73)
T0
For railway or aircraft traffic measurements, the standardized sound exposure level difference,
DE2m,nT , is
T
(3.74)
DE2m,nT = DE2m + 10 lg
T0
When measurements of D2m,n or D2m,nT are taken according to the relevant Standard (ISO 140
Part 5) there will not usually be significant differences between results taken using road traffic
noise or a loudspeaker as the sound source (EN 12354 Part 3).
3.5.4
Other measurement issues
In this section we look at the effect of background noise on sound pressure measurements and
conversion of measurements from one-third-octave to octave-bands. As it is possible to determine the sound reduction index using either sound pressure or sound intensity measurements
we see how the two results can be compared. In addition a few practical issues are discussed
such as the drying out of test elements and identifying sound leaks.
3.5.4.1
Background noise correction
When measuring constructions that have high levels of airborne or impact sound insulation, the
sound pressure levels in the receiving room may not be high above the background noise level.
This quite often occurs in the field because occupied buildings or semi-completed buildings
268
Chapter 3
7
6
Lp ⫺ Lp,c (dB)
5
4
3
2
1
0
1
2
3
4
5
6
7
8
9
10 11 12 13 14 15
Lp ⫺ Lp,b (dB)
Figure 3.22
Background noise correction. This shows the correction, in terms of Lp − Lp,c , that is made to the measured sound pressure
level, Lp to remove the effect of background noise level, Lp,b .
with ongoing building work have higher background noise levels than a laboratory. The effect of
measuring sound pressure levels that are close to the background noise is to give a higher level
than would be measured in the absence of the background noise. The sound pressure level
measured in the receiving room can be corrected to remove the effect of steady background
noise. The corrected sound pressure level, Lp,c , is given by:
Lp,c = 10 lg (10Lp /10 − 10Lp,b /10 )
(3.75)
where Lp is the measured sound pressure level in the presence of background noise and Lp,b
is the background noise level.
The effect of background noise on the measured sound pressure level is shown in Fig. 3.22
using Eq. 3.75.
For airborne and impact sound insulation in the relevant Standards, this correction is only prescribed for specific ranges of Lp − Lp,b . For laboratory measurements (ISO 140) this correction
is applied when 6 dB < Lp − Lp,b < 15 dB. For field measurements (ISO 140) this correction is
applied when 6 dB < Lp − Lp,b < 10 dB.
Under controlled conditions in a laboratory the background noise level is usually steady and
can be quantified with greater accuracy than in the field. When Lp − Lp,b ≥ 15 dB, we find that
Lp − Lp,c ≤ 0.1 dB. As we measure sound pressure levels to the nearest 0.1 dB the correction
is not used when Lp − Lp,b ≥ 15 dB. In the field there is less control over the measurement
environment, and estimates of the background noise level during the measurement are more
prone to error.
When Lp − Lp,b ≤ 6 dB there is insufficient signal to ensure that the large corrections (see
Fig. 3.22) are appropriate. The approach taken in the relevant Standards is to limit the
269
S o u n d
I n s u l a t i o n
correction to 1.3 dB rather than to continue to use Eq. 3.75 (ISO 140). This value corresponds
to the correction that is calculated when Lp − Lp,b = 6 dB.
The background noise in the receiving room is often measured once before, and once after the
measurement with the loudspeaker, to check for any significant change. We therefore have to
assume that these background noise levels are representative of the background noise during the measurement. This is reasonable for background noise that can be considered as a
stationary signal; one example could be environmental noise from a steady flow of road traffic
that is transmitted through the building façade. However, with field measurements in occupied buildings or semi-completed buildings with ongoing building work, there is an increased
likelihood of unsteady background noise. Hence, it is useful to monitor the maximum sound
pressure level during the measurement period to check for transient noises, such as doors
slamming in the building.
3.5.4.2 Converting to octave-bands
Airborne sound insulation descriptors in one-third-octave-bands can be converted to octavebands using the three one-third-octave-bands from which the octave-band is formed,
XOB
3
1 −XTOB,n /10
= −10 lg
10
3
n=1
(3.76)
′
where X represents R, R45
◦ , Dn , DnT , etc.
3.5.4.3
Comparing the airborne sound insulation measured using
sound pressure and sound intensity
When choosing suitable constructions at the design stage it may be necessary to compare
the sound reduction index, R, for one product, with the intensity sound reduction index, RI , for
another product. To compare the performance of the two products we need to account for the
different methods by which R and RI are calculated (Jonasson, 1993). In the derivation of R, we
noted that the transmitted sound power is underestimated in the low-frequency range because
no account is taken of the higher energy density near the room boundaries. The reason that
no correction term is used in the definition of the sound reduction index is partly because the
Waterhouse correction is not considered sufficiently accurate for the smaller room volumes
that can be used in transmission suites (i.e. 50 m3 ). However, if we want to compare R with
RI then the Waterhouse correction can still be used as an estimate. The modified intensity
sound reduction index, RI,M , is introduced to represent RI after it has been modified with the
Waterhouse correction (ISO 15186 Part 1),
RI,M = RI + CW
(3.77)
where CW is the Waterhouse correction in decibels (Eq. 1.157) for the sound field in the receiving room. Note that calculation of CW requires knowledge of the volume and total surface area
of the receiving room.
RI,M can now be compared directly with R. Note that RI in the low-frequency range is the better
estimate of the ‘true’ sound reduction index as defined by Eq. 3.33; it is only for the purpose of
this comparison that it is necessary to introduce some additional uncertainty.
270
Chapter 3
3.5.4.4
Variation in the sound insulation of an element due to moisture
content and drying time
For some elements, such as masonry walls or screed floating floors, the sound insulation will
vary depending on the degree to which they have dried out. These changes are relevant in the
field as well as in the laboratory.
In the laboratory, individual components (such as bricks, blocks or sheet material) can be
stored in dry conditions before building the test element. However, the moisture content usually
increases once they are bonded together with mortar or a water-based adhesive. Surface
finishes, such as wet plaster also increase the moisture content of the wall or floor.
Potential changes to the sound insulation can be considered in terms of the mass, stiffness,
coupling losses, and internal losses of the element. In most cases we only have information
on the increase in the mass per unit area due to moisture when compared to standard conditions. At frequencies where the sound reduction index is only determined by non-resonant
(mass law) transmission, estimates for the increase in sound insulation can be calculated (see
Section 4.3.1.2). When there is only resonant transmission the effect of moisture on the plate
stiffness and internal damping needs to be considered, although information is rarely available
to quantify the effect. This is partly due to other confounding factors. As the plate dries it tends
to shrink away from the test aperture, which can reduce the plate coupling losses, and hence
the total loss factor. This potentially introduces air paths at cracks or gaps around the plate
perimeter (Schmitz et al., 1999). Resonant transmission dominates over most of the frequency
range for masonry/concrete elements and it is possible for the effect of moisture to increase
or decrease the sound reduction index at different frequencies.
3.5.4.5 Identifying sound leaks and airpaths
With most constructions care needs to be taken to avoid gaps or cracks around the perimeter
of the element. However, when the airborne sound insulation curve of a wall or floor decreases
or forms a plateau in the high-frequency range it does not necessarily mean that there is
a sound leak or air path. If problems due to high background noise have been ruled out,
this feature of the curve could be due to a number of factors including: a plateau due to the
thick plate effect, a porous plate material, a plate with a high critical frequency that forms the
surface of the wall or floor, or flanking transmission. These aspects are discussed in Chapters 4
and 5.
High-frequency leaks can often be detected with ones ears, but they can also be detected
using a sound intensity probe (Section 3.10.1). The axis of a p–p sound intensity probe can
be slowly scanned parallel to the surface that is suspected of containing the leak. This is
different to how the probe is normally used to determine the sound power, for which the
probe axis is perpendicular to the measurement surface. As the probe moves past a leak,
the intensity reading will flip from positive to negative or vice versa. Leaks tend to transmit
significant levels of sound in the mid- or high-frequency range, so a 12 mm spacer is usually
needed.
With windows and doors there are many possible airpaths around the frame or locking devices.
An assessment of their effect on the airborne sound insulation is most simply made by testing
before and after sealing all the joints with tape. Airtightness testing may also be used to detect
potential problems with window units before they are installed.
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S o u n d
I n s u l a t i o n
3.6 Impact sound insulation (floors and stairs)
As an introduction to impact sound insulation it is useful to briefly re-consider airborne sound
insulation. To measure airborne sound insulation we use sound pressure level measurements
in the source and receiving rooms to relate the transmitted sound to the incident sound. More
specifically, for the sound reduction index we determine the ratio of the sound power that is
incident on the test element to the sound power transmitted by that same element. Hence when
power is ‘injected’ into the source room via a loudspeaker there is no need to know its sound
power output. As long as the sound transmission process from one room to the other is a linear
process (which it normally is), we can take measurements using loudspeakers with different
sound power outputs and still calculate the same value for the airborne sound insulation. This
assumes that there is a relatively flat spectrum in the source room and that we can measure a
signal in the receiving room that is well-above the background level.
To measure the impact sound insulation we need to ‘inject’ power into a floor using a structureborne sound source. Taking the approach used for airborne sound insulation, the starting point
is to consider how we could determine the structure-borne sound power input into the floor.
Although there are methods for measuring this power input, it would be difficult and very timeconsuming to take accurate measurements on all types of floors. So to avoid measuring the
power input, the logical step is to standardize the excitation source. This is the approach that
is used for impact sound insulation measurements in the Standards. However, we will soon
see that this does not fully take into account the complexities involved with power input into a
structure.
There are many different structure-borne sound sources on floors including footsteps, impacts
from children playing, dropped objects, chairs or other furniture being dragged across the floor,
and vibrations from machinery such as washing machines or mechanical services. Some of
these impact sources generate impulses due to a falling mass, such as a dropped object or
footsteps; these only apply a force and are often described as force sources. Other sources
may apply forces and/or moments to the structure and generate an impulse or a continuous
signal. No single artificial source can accurately represent all of these real sources. Even if
we restricted ourselves to representing impacts from footsteps, there is such a wide range of
body weights, walking styles, footwear, and impact velocities that we would need to identify
which type of footstep the structure-borne sound source should represent. This is illustrated
by the wide range of different spectra for different walkers in Fig. 3.23 (Vian and Drouin, 1977).
The inherent difficulty in defining a measurement and rating system for impacts on floors has
occupied researchers ever since the first experimental work in the late 1920s. For a thorough
historical review the reader is referred to papers by Schultz (1981) and Cremer (1976/1977);
the following papers can then be used as starting points to continue beyond 1980: Bodlund
(1985), Rindel and Rasmussen (1996), Warnock (1998), Tachibana et al. (1998), and Scholl
(2001).
For the purpose of standardization it is clear that a well-defined impact source is necessary to
establish minimum regulatory requirements and to allow the ranking of impact sound insulation
provided by different floors. To achieve this it is not only necessary to standardize the source,
but also the rating procedure that is used to produce a single-number quantity. The primary
source is the ISO tapping machine which is described in International Standards (ISO 140
Parts 6 & 7). The test method (ISO 140 Parts 6 & 7) is defined along with a rating procedure
(ISO 717 Part 2) to allow comparison of the impact sound insulation of different floors.
272
Chapter 3
Average maximum time-weighted sound level (dB)
60
50
40
30
20
10
63
125
250
500
1000
2000
Octave-band centre frequency (Hz)
Figure 3.23
Sound pressure level spectra (40 curves) in a receiving room below a 140 mm concrete floor slab from 40 different walkers
(male and female) in a variety of footwear walking in a natural manner. For each of the 40 walkers the average value for the
maximum time-weighted sound pressure level (time constant, τ = 35 ms) was calculated from the loudest 25% (i.e. upper
quartile) of levels measured during a time period containing approximately 50 footsteps. Measured data are reproduced with
permission from Vian and Drouin (1977).
3.6.1 Laboratory measurements
Laboratory measurement of impact sound insulation requires measurement of the temporal and
spatial average sound pressure level in a room, Lp , when the floor above is excited by the ISO
tapping machine (see Fig. 3.24). For a given sound power transmitted into the receiving room,
the mean-square sound pressure in this room is inversely proportional to the absorption area,
A, of that room (Eq. 3.36). The normalized impact sound pressure level, Ln , is therefore defined
by using a reference absorption area, A0 , of 10 m2 for the receiving room (ISO 140 Part 6).
Ln = Lp + 10 lg
A
A0
(3.78)
Note that there is no need to normalize the sound pressure level to the surface area of the test
element as with the sound reduction index. This can be seen from the derivation of Ln for a
solid homogeneous isotropic plate in Section 4.4.1.
3.6.1.1
Improvement of impact sound insulation due to floor coverings
There are many different types of floor covering ranging from soft floor coverings (e.g. carpet,
vinyl) to rigid walking surfaces (e.g. floating floors, parquet, ceramic tiles). The improvement of
impact sound insulation due to a floor covering depends upon the base floor. For this reason the
base floor constructions are prescribed in the laboratory measurement Standards; these are
categorized as either heavyweight (ISO 140 Part 8) or lightweight (ISO 140 Part 11) base floors.
273
S o u n d
I n s u l a t i o n
ISO tapping
machine
Receiving
room
Test
element
Figure 3.24
Outline sketch of a transmission suite for impact sound insulation measurements (upper room is displaced upwards).
3.6.1.1.1
Heavyweight base floor (ISO)
The heavyweight base floor can generally be referred to as a 140 mm reinforced concrete
slab; the exact details are given in the relevant Standard (ISO 140 Part 8). The improvement
of impact sound insulation, L, is measured using the ISO tapping machine and is defined as:
L = Ln0 − Ln
(3.79)
where Ln0 is the normalized impact sound pressure level for the ISO heavyweight base floor
without the floor covering and Ln is the normalized impact sound pressure level with the floor
covering.
3.6.1.1.2
Lightweight base floors (ISO)
Compared with heavyweight base floors, it is more awkward to choose a single lightweight base
floor that adequately represents the range of lightweight floors that are built around the world.
For this reason, three different timber floor bases are described in the relevant Standard (ISO
140 Part 11); this makes it important to state which base floor has been used. For lightweight
base floors, the improvement of impact sound insulation is usually measured with the ISO
tapping machine, but to assess heavy impacts, measurements can also be made using the
ISO rubber ball (ISO 140 Part 11).
274
Chapter 3
ISO tapping machine: The improvement of impact sound insulation, Lt , measured with the
ISO tapping machine is
Lt = Ln,t,0 − Ln,t
(3.80)
where Ln,t,0 is the normalized impact sound pressure level for the ISO lightweight base floor
without the floor covering and Ln,t is the normalized impact sound pressure level with the floor
covering.
Equation 3.80 also applies to the modified ISO tapping machine that is described in the relevant
Standard (ISO 140 Part 11); this modification is discussed in Section 3.6.3.4.
ISO rubber ball: For each impact from the rubber ball, the maximum time-weighted (Fast)
sound level, LFmax , is measured and averaged for different excitation and microphone positions
to give the impact sound pressure level, Li,Fmax . The improvement of impact sound insulation
is then calculated using:
Lr = Li,Fmax,0 − Li,Fmax
(3.81)
where Li,Fmax,0 is the impact sound pressure level for the ISO lightweight base floor without the
floor covering and Li,Fmax is the impact sound pressure level with the floor covering.
3.6.2
Field measurements
Calculation of the normalized impact sound pressure level for the field situation, L′n , is identical
to that in the laboratory (ISO 140 Part 7) where:
A
L′n = Lp + 10 lg
(3.82)
A0
To find the standardized impact sound pressure level, L′nT , we know that for a given sound
power transmitted into the receiving room, the average mean-square sound pressure in this
room is proportional to the reverberation time, T , of that room. Hence L′nT is defined using a
reference reverberation time, T0 , for the receiving room, which for dwellings is 0.5 s (ISO 140
Part 7).
T
′
(3.83)
LnT = Lp − 10 lg
T0
3.6.3
ISO tapping machine
The ISO tapping machine (see Fig. 3.25) has a line of five equally spaced hammers that are
driven in such a way that there are 10 impacts upon the floor every second, with 100 ms between
successive impacts (ISO 140 Parts 6 & 7). The reason for using a train of impacts instead of
a single impact was to allow accurate measurements at a time when instrumentation was
better suited to measuring continuous signals; nowadays it is possible to accurately measure
a single impulse. The requirement for each tapping machine hammer is that the momentum
of each hammer impact should represent a free-falling mass of 0.5 kg with a drop-height of
0.04 m. In practice the mass will not be free-falling, and there will be some friction losses
in the guidance device used to minimize lateral motion of the hammer. The relevant ISO
Standard gives tolerances on the mass of the hammer, and the velocity at impact, hence,
275
S o u n d
I n s u l a t i o n
Figure 3.25
ISO tapping machine: an example of a commercially available machine. Photo provided by Norsonic.
where necessary, slightly greater drop-heights than 0.04 m can be used to compensate for
friction losses.
The ISO Standards contain a thorough specification for the ISO tapping machine and its hammers because relatively small changes in the design have been found to affect its power input
(e.g. see Bodlund and Jonasson, 1983; Gösele, 1956; Goydke and Fischer, 1983). Critical
aspects of the design include: the distance between the hammers, the distance between the
hammers and the tapping machine supports, vibration isolation of the supports, hammer mass,
hammer dimensions, curvature and material of the hammer impact surface, hammer velocity
at impact, time between impacts, and the time between impact and the lifting of the hammer.
3.6.3.1 Force
The ISO tapping machine applies a force due to the hammers that impact upon the surface of
the floor at regular time intervals. For a free-falling mass, m, which falls from a height, h, under
acceleration due to gravity, g, the kinetic energy equals the potential energy, therefore:
mv 2
= mgh
2
(3.84)
from which the velocity, v0 , of the mass at impact is
v0 =
2gh
and the hammer velocity at impact for the ISO tapping machine will be 0.886 m/s.
276
(3.85)
Chapter 3
Force, F(t)
Ti
t⫽0
Time, t
Figure 3.26
Time history of impact forces from the ISO tapping machine.
From Vér (1971) and Cremer et al. (1973) the force can now be estimated for impacts upon a
surface that occur with a frequency, fi , and a time period, Ti , between impacts, where Ti = 1/fi .
For the tapping machine, fi is 10 Hz and Ti is 0.1 s. The impact forces are as indicated in
Fig. 3.26 where the first impact is positioned at time t = 0. In the time domain the force, F(t),
is represented by the Fourier series:
F(t) =
∞
Fn cos(2πfi nt)
(3.86)
n=1
where n = 1, 2, 3, etc. for the sequence of Fourier frequency components, Fn . These
components are calculated using:
2 Ti
F(t) cos(2πfi nt)dt
(3.87)
Fn =
Ti 0
For short duration impacts it can be assumed that cos (2πfi nt) ≈ 1, hence,
2 Ti
F(t)dt
Fn ≈
Ti 0
(3.88)
Over the building acoustics frequency range it is reasonable to assume that short duration
impacts occur on homogeneous bare concrete floors with a thickness of at least 100 mm. We
will soon look at this assumption in more detail. Returning to Eq. 3.88, the integral is solved
with Newton’s second law of motion, F = ma. This can be written as F = m dv
and re-arranged
dt
to give Fdt = mdv, from which the integral over the time period, Ti , is equal to the change in
momentum, such that,
Ti
F(t)dt = mvTi − mv0
(3.89)
0
At time, Ti , after a hammer impact, the velocity of that hammer should be zero. Therefore the
magnitude of the peak force from the tapping machine, |Fn |, can be calculated from:
|Fn | ≈
2
mv0 = 2fi m 2gh
Ti
(3.90)
The force spectrum is a line spectrum as shown in Fig. 3.27 where the frequency lines occur
at 10 Hz intervals. For the practical purposes of measurement and prediction, we need to use
277
S o u n d
I n s u l a t i o n
Force, Fn
fi
Frequency, f
Figure 3.27
Fourier frequency components from the ISO tapping machine.
v0
Tapping machine hammer
m
jK
K
Floor
jZ
Zdp
F1(t)
Figure 3.28
Lumped element model representing the hammer impact upon the floor using a mass–spring–dashpot system.
this line spectrum to calculate the mean-square force spectrum in one-third-octave or octavebands. For a frequency band with a bandwidth, B, there will be B/fi frequency lines in the band,
so the mean-square force in a frequency band is
2
=
Frms
|Fn |2 B
2fi
(3.91)
where B = 0.23f for one-third-octave-bands, B = 0.707f for octave-bands, and f is the band
centre frequency.
Hence, when it can be assumed that the impacts are of short duration, the mean-square force
from the tapping machine is
2
Frms
= 3.9B
(3.92)
We will now look at the assumption of short duration impacts. The time-history of a single force
pulse from the ISO tapping machine hammer can be calculated using a lumped element model
(Brunskog and Hammer, 2003; Lindblad, 1968). This model accounts for the effect of both
the hammer impedance and floor impedance on the force pulse using a mass–spring–dashpot
system (see Fig. 3.28). The hammer is represented as a lump mass. The floor is represented
by a spring for the contact stiffness, K , in series with a dashpot damper for the floor impedance,
Zdp . It is assumed that both K and Zdp are frequency-independent. When we assumed a short
duration impact, the force was only dependent upon the hammer mass. Now we can use this
278
Chapter 3
Over-critical oscillation
Under-critical oscillation – initial force pulse
Force, F1(t)
Under-critical oscillation – successive oscillations
Time, t
Figure 3.29
Force pulse from a single tapping machine hammer – examples of over-critical and under-critical oscillations.
more complete model to assess the effect of the contact stiffness and the floor impedance on
the force applied by the hammer.
The equations of motion for the mass–spring–dashpot system in terms of the displacement,
ξ, are:
d2 ξK
= K (ξK − ξZ )
dt 2
dξZ
K (ξZ − ξK ) = Zdp
dt
m
(3.93)
(3.94)
From Lindblad (1968), these equations can be solved to give the force pulse for a single
hammer impact, F1 (t). The resulting equations for the force pulse depend on whether the
2
mass–spring–dashpot system gives rise to an over-critical oscillation (Km ≥ 4Zdp
), or an under2
critical oscillation (Km < 4Zdp ). Examples of the different force pulses due to over-critical and
under-critical oscillations are shown in Fig. 3.29. For over-critical oscillations, the force pulse
decays to zero and takes only positive values. After the initial force pulse of an under-critical
oscillation, F1 (t) oscillates with alternate positive and negative values about the zero force line.
2
When Km ≥ 4Zdp
(over-critical oscillation),
⎞
⎛
2
K
K
−Kt
v0 K exp
− ⎠
sinh ⎝t
2Zdp
2Zdp
m
F1 (t) =
2
K
K
−
2Zdp
m
(3.95)
279
S o u n d
I n s u l a t i o n
2
When Km < 4Zdp
(under-critical oscillation),
−Kt
v0 K exp
2Zdp
F1 (t) =
⎛
sin ⎝t
K
−
m
K
−
m
K
2Zdp
2
K
2Zdp
2
⎞
⎠
(3.96)
The magnitude of the peak force from the tapping machine, |Fn |, can now be calculated by
taking the Fourier transform of the force pulse, F1 (t), for this single hammer impact, and
accounting for the impact repetition rate from the tapping machine, fi . Equation 3.91 can then
be used to calculate the mean-square force from the tapping machine in a frequency band. For
the under-critical oscillations only the initial force pulse that has zero or positive force values is
used to determine the force spectrum, with all subsequent values of F1 (t) due to the oscillations
set to zero before taking the Fourier transform.
There are two types of contact stiffness that can be used in the calculation of the force pulse.
One type is for the contact stiffness of a plate material, such as concrete or chipboard, in the
contact area of the hammer. The other type is for a soft floor covering, such as carpet on
a heavyweight floor, where it can be assumed that the contact stiffness is only determined
by the covering and is not affected by the contact stiffness of the plate underneath the soft
covering. Effectively we are treating the soft floor covering or the plate material that deforms
in the contact area as a linear spring.
The contact stiffness of the plate material, K , can be estimated using (Timoshenko and Goodier,
1970):
2rE
K=
(3.97)
1 − ν2
where r is the radius of the circular contact area of the hammer with the plate, E is the Young’s
modulus of the plate, and ν is the Poisson’s ratio of the plate.
The contact stiffness for a soft covering, K , is
K=
Eπr 2
d
(3.98)
where d is the thickness of the soft covering and E is the Young’s modulus of the soft covering.
Calculation of the contact stiffness assumes a circular contact area of radius, r , for the tapping
machine hammer. The hammer radius, 15 mm, can be used as an estimate because the
actual contact area is not well-defined due to the spherical impact surface of the hammer, and
deformation of the plate or soft floor covering upon impact.
For the floor we will assume an infinite homogeneous plate, for which the driving-point
impedance, Zdp , is frequency-independent and given by Eq. 2.190. It is now possible to look
at the force pulse on a range of flooring materials using material properties that are based on
concrete, sand-cement screed, chipboard, and oriented strand board (OSB). For a 140 mm
concrete slab or a 65 mm sand-cement screed, the plate impedance and the contact stiffness
are significantly higher than with 22 mm chipboard or 15 mm OSB. For the concrete slab and
screed, the hammer rebounds from the plate with an under-critical oscillation and the force
pulse has a short duration (see Fig. 3.30). However, for the chipboard and OSB plates there
is no distinct rebound as with a concrete slab, and the pulse has a much longer duration due
280
Chapter 3
20 000
140 mm concrete slab (308 kg/m2)
18 000
16 000
65 mm sand-cement screed (130 kg/m2)
Force, F1 (t) (N)
14 000
12 000
10 000
8000
6000
4000
2000
0
0
0.2
0.4
0.6
0.8
1
1.2
1.4
1.6
1.8
2
Time (ms)
Figure 3.30
Force pulse from the tapping machine hammer on concrete and sand-cement screed plates. (NB: Only the initial part of the
under-critical force pulse is shown which has zero or positive force values.)
2000
22 mm chipboard (17 kg/m2)
1800
1600
15 mm OSB (9 kg/m2)
Force, F1 (t) (N)
1400
1200
1000
800
600
400
200
0
0
0.2
0.4
0.6
0.8
1
1.2
Time (ms)
1.4
1.6
1.8
2
Figure 3.31
Force pulse from the tapping machine hammer on chipboard and OSB plates.
to the over-critical oscillation (see Fig. 3.31). So for common flooring materials the range of
values for both the impedance and the contact stiffness has a significant effect on the shape
of the force pulse, which, in turn, determines the force spectrum.
Figure 3.32 shows the force spectrum in terms of the magnitude of the peak force, |Fn |, calculated from the Fourier transform of the force pulses for the four different plates. We can assess
the trends from the viewpoint of both increasing and decreasing frequency. With decreasing
281
S o u n d
I n s u l a t i o n
Magnitude of the peak force, |Fn| (N)
10
|Fn|upper
|Fn|lower
1
140 mm concrete slab (308 kg/m2)
65 mm sand-cement screed (130 kg/m2)
22 mm chipboard (17 kg/m2)
15 mm OSB (9 kg/m2)
0.1
10
100
1000
10 000
Frequency (Hz)
Figure 3.32
Force spectrum for the tapping machine on four different plates calculated using force pulses from the mass–spring–dashpot
model.
frequency, |Fn |, asymptotes between two limits (Brunskog and Hammer, 2003):
|Fn |lower =
1
mv 0
Ti
(3.99)
|Fn |upper =
2
mv 0
Ti
(3.100)
The upper limit, |Fn |upper , corresponds to the situation where the hammer rebounds from the
floor with the velocity, v0 , and is the same as Eq. 3.90. The lower limit, |Fn |lower , occurs when the
hammer momentum is dissipated during the impact such that the hammer does not rebound.
In terms of the mean-square force, the difference between the lower and upper limit is 6 dB.
The short duration pulse on the 140 mm concrete slab gives a relatively flat force spectrum
across the building acoustics frequency range. Below 4000 Hz the mean-square force values
are within 1 dB of the upper limit, |Fn |upper . For concrete floor slabs of at least 100 mm thickness
it is reasonable to estimate the mean-square force using Eq. 3.92 which assumes short duration
impacts.
For the 22 mm chipboard and 15 mm OSB plates, the longer duration pulse means that the force
spectrum is not flat and decreases above 100 Hz. In addition, the force is significantly lower
than with the concrete slab, with |Fn | tending towards the lower limit, |Fn |lower , at frequencies
below 100 Hz. This implies that use of Eq. 3.92 will overestimate the mean-square force for
the 22 mm chipboard and 15 mm OSB plates.
With increasing frequency, there is a cut-off frequency, fco , above which there is no longer a
relatively flat force spectrum and the force is significantly reduced. This cut-off frequency can
282
Chapter 3
be calculated from (Brunskog and Hammer, 2003):
⎡
⎤
2
1 ⎣ K
K⎦
K
fco =
−
−
2π 2Zdp
2Zdp
m
(3.101)
2
when Km ≥ 4Zdp
(over-critical oscillation), and
fco
1
=
2π
K
m
(3.102)
2
when Km < 4Zdp
(under-critical oscillation).
The lumped element model gives an insight into the differences between the force spectra
from the tapping machine with different plate materials. This assumes that the walking surface behaves as an infinite plate with frequency-independent driving-point impedance and
that the estimate of the contact stiffness is reasonable. Examples of measured driving-point
impedances can be found in Section 2.8 although they are presented in terms of the drivingpoint mobility; these indicate when the assumption of an infinite plate is reasonable. Lightweight
floors do not always have impedances that are independent of frequency because they often
need joists or battens to support the plate; the impedance can vary over the surface depending on whether the hammer makes contact with the plate on top of the joists or battens,
or in-between them. For lightweight floors with frequency-dependent impedance, numerical
methods can be used to calculate the force spectrum (Brunskog and Hammer, 2003).
3.6.3.2 Power input
In the previous section the mean-square force was calculated using two different approaches.
The first approach was purely based around the hammer momentum. This assumed a short
duration impact and made no further consideration of the interaction between the hammer and
the floor. This resulted in the simple expression in Eq. 3.92. The second approach took account
of the interaction between the hammer and the floor by using a mass–spring–dashpot model
to calculate the force pulse from a single hammer impact.
2
When Frms
is calculated using the mass–spring–dashpot model the power input is given by:
+
*
1
2
(3.103)
Win = Frms
Re
Zdp
where Zdp is the driving-point impedance of the floor.
2
When Frms
is calculated from Eq. 3.92, the power input needs to take account of the hammer
impedance, Zh , which is in series with the driving-point impedance of the floor during the impact
(Cremer et al., 1973). Hence, the power input is
+
*
1
2
(3.104)
Win = Frms
Re
Zdp + Zh
where the impedance of the hammer is that of a lump mass, m, hence Zh = iωm, where
m = 0.5 kg.
Using the infinite plate impedance, the power input is therefore:
2
Win = Frms
2.3ρcL h2
(2.3ρcL h2 )2 + (ωm)2
(3.105)
283
S o u n d
I n s u l a t i o n
125
140 mm concrete slab (308 kg/m2)
Power input (dB)
120
65 mm sand–cement screed (130 kg/m2)
115
22 mm chipboard (17 kg/m2)
110
15 mm OSB (9 kg/m2)
105
100
95
90
85
80
75
50
80
125 200 315 500 800 1250 2000 3150 5000
One-third-octave-band centre frequency (Hz)
Figure 3.33
Power input for the tapping machine on four different plates calculated using force pulses from the mass–spring–dashpot
model.
Figure 3.33 shows the power input calculated using force pulses from the more accurate mass–
spring–dashpot model for the four different plates. The power inputs are quite different for the
four plates. However, this single fact does not imply a favourable or unfavourable bias towards
one type of floor by using the tapping machine; any falling mass that impacts upon a floor will
have a different power input on different floors, whether it is the hammer of a tapping machine
or a child jumping. When the magnitude of the hammer impedance is negligible in comparison with the driving-point impedance of the floor, the hammer impedance does not affect the
power input. Hence, the power input in one-third-octave or octave-bands increases by 3 dB
per doubling of the band centre frequency. This occurs for the 140 mm concrete slab and the
65 mm sand-cement screed below 1000 Hz. Above 1000 Hz, the hammer impedance starts to
become significant and slightly reduces the power input in the high-frequency range. For the
22 mm chipboard and 15 mm OSB, the magnitude of the hammer impedance is significant in
comparison with the plate impedance across the majority of the building acoustics frequency
range. Hence, the curves do not increase by 3 dB per doubling of the band centre frequency.
Above a limiting frequency, flimit , the power input starts to decrease with increasing frequency. This is defined by the frequency at which the plate impedance equals the magnitude of
the hammer impedance (Cremer et al., 1973),
flimit =
1 Zdp
2π m
(3.106)
3.6.3.3 Issues arising from the effect of the ISO tapping machine hammers
The effect of the hammer impedance on the power input has implications for the comparison
of impact sound insulation on different floors. For a floor with a driving-point impedance, Zdp ,
we can use Eq. 3.104 to calculate the ratio of the power input from the ISO tapping machine
hammer, to the power input from any real impact due to a falling mass (Schultz, 1975). If the
284
Chapter 3
ratio of the power inputs for any type of floor is constant for each frequency band, then we
can simply use the ISO tapping machine to assess the impact sound insulation against the
real impact by comparing the normalized (or standardized) impact sound pressure levels for
different floors. To calculate this ratio we will assume the same impact repetition rate, fi , for
the real impact as for the ISO tapping machine. This gives the ratio of the power inputs as:
Win(h)
=
Win(r)
2
Frms(h)
2
Frms(r)
+
1
Zdp + Zh
*
+
1
Re
Zdp + Zr
Re
*
(3.107)
where the subscript, h, is used for the ISO tapping machine hammer and the subscript, r, is
used for the real impact.
It is reasonable to assume that footsteps are the most common source of impacts on floors;
hence, we will use this source as our real impact. However, without evidence we cannot
simply assume that footsteps are the most common source of annoyance when people give a
subjective evaluation of the impact sound insulation. The ratio of the power inputs is the product
2
2
of two components, the force ratio, Frms(h)
/Frms(r)
, and the impedance ratio, Re{1/(Zdp + Zh )}/
Re{1/(Zdp + Zr )}. To appreciate the issues involved with footsteps and other real impacts it
is useful if we look at these two components individually, before considering them together.
For the sake of simplicity it would be convenient if both the force and impedance ratios were
independent of frequency, and invariant with the type of floor surface. This would allow a simple
ranking of the impact sound insulation using the normalized (or standardized) impact sound
pressure levels. Unfortunately, we will soon see that there is no simple outcome.
2
2
We first look at the force ratio, 10 lg (Frms(h)
/Frms(r)
) where the mean-square force for the real
impact corresponds to the force applied by the heel of an adult female walking in a pair of
high-heeled shoes as measured by Watters (1965). It has previously been noted that the force
depends on the contact stiffness of the surface upon which the impact is made. The calculated
force is shown in Fig. 3.34 for two different walking surfaces: hardwood floor and thin carpet on
a concrete slab. The first point to note is that the force from the ISO tapping machine is higher
than with footsteps. A high input force from the ISO tapping machine is used to ensure that
the impact sound pressure levels are well-above the background noise for both laboratory and
field measurements. The next point is that the force ratio varies with frequency. For thin carpet
on a concrete slab, the force ratio rises with a particularly steep gradient above 125 Hz which
indicates that the force spectrum for the ISO tapping machine hammer is significantly different
to this particular footstep. This leads on to a particular issue with some soft floor coverings;
non-linearity with high input forces.
As a consequence of the ISO tapping machine applying a high input force, the response of
some soft floor coverings is non-linear (Lindblad, 1968, 1983). This occurs where the contact
stiffness of the soft floor covering depends upon the force. Therefore, when the ISO tapping
machine is used to measure the impact sound insulation of floors with soft floor coverings, the
results are only representative of impacts that are the same as those from the ISO tapping
machine hammer. Lindblad’s (1968) investigation into non-linearity used a modified ISO tapping machine to give three different drop-heights, 40, 12.6, and 4 mm; the lower the drop-height,
the lower the force. For linear systems, each successive decrease in drop-height would correspond to a 5 dB reduction in the normalized impact sound pressure level for each frequency
band. This indicates whether the contact stiffness of a soft floor covering can be considered as
285
S o u n d
I n s u l a t i o n
55
50
10 lg (F2rms(h) /F2rms(r)) (dB)
45
40
35
30
25
20
15
Hardwood floor
10
Thin carpet on concrete slab
5
0
31.5
63
125
250
500
1000
2000
Octave-band centre frequency (Hz)
4000
Figure 3.34
Mean-square force ratio for the ISO tapping machine relative to the force applied by the heel of an adult female walking in a
pair of high-heeled shoes. Measured data are reproduced with permission from Watters (1965).
a linear or a non-linear spring. The results from two different soft floor coverings (a) and (b) on
a concrete slab are shown in Fig. 3.35. Covering (a) shows a general trend of successive 5 dB
reductions in the low-frequency range; the differences become larger at higher frequencies
but the curve shapes are generally similar across the frequency range. Covering (b) shows
distinct non-linearity. The different curve shapes indicate that the contact stiffness is effectively
increasing with increasing force and behaves as a non-linear (hardening) spring.
Non-linearity complicates matters because the non-linear response of a soft floor covering may
be relevant to some real impacts, but irrelevant to others. This makes it difficult to optimize
and rank-order the performance of soft coverings to provide impact sound insulation against
a range of impact sources (Lindblad, 1983). Another issue with soft floor coverings is that the
measured improvement of impact sound insulation also depends upon the mass and contact
area of the ISO tapping machine hammer (Vér, 1971); these are not necessarily representative
of the mass and contact area of footsteps.
We will now look at the impedance ratio, Re{1/(Zdp + Zh )}/Re{1/(Zdp + Zr )}. This ratio will only
be constant if the hammer impedance, Zh , is equal to the impedance of the real impact source,
Zr for footsteps. The impedance of the human walker results from the combination of footwear
and the mechanics of the human body. For footsteps we are interested in the impedance of
the heel of the foot or shoe as it lands upon the floor. Watters (1965) and Scholl (2001) carried
out heel-impedance measurements as part of their investigations into the tapping machine.
Watters (1965) measured the heel-impedance of an adult female (weight of at least 45 kg) in a
pair of high-heeled shoes. Scholl (2001) measured the heel-impedance of two adult males when
barefoot and in flat-heeled shoes. These measurements are shown in Fig. 3.36 when the male
or female subjects are standing. The measurements indicate a wide range of heel-impedances
for different walkers and different footwear, particularly in the low- and mid-frequency ranges.
286
Chapter 3
(a)
Normalized impact sound pressure level, Ln (dB)
70
60
50
40
30
20
Drop-height (mm)
40
12.6
10
0
100
4
160
250
400
630
1000 1600
One-third-octave-band centre frequency (Hz)
2500
160
250
400
630
1000 1600
One-third-octave-band centre frequency (Hz)
2500
(b)
Normalized impact sound pressure level, Ln (dB)
70
60
50
40
30
20
10
0
100
Figure 3.35
Normalized impact sound pressure levels measured using different drop-heights (40, 12.6, and 4 mm) on a modified ISO
tapping machine. (a) Floor: 0.8 mm vinyl glued to a resilient backing of 2 mm vinyl foam on 160 mm concrete floor slab.
(b) Floor: 1.1 mm vinyl glued to a resilient backing of 2 mm felt on 160 mm concrete floor slab. Measured data are reproduced
with permission from Lindblad (1968).
In the same way that the tapping machine hammer can be represented as a lump mass,
we can represent the heel-impedance using lump elements of mass, Z = iωm, and stiffness, Z = k/iω (Watters, 1965). Figure 3.37 clearly shows that the impedance of a 0.5 kg
mass used for the ISO tapping machine hammer does not represent the range of measured
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80
High-heeled shoes – Female (a)
Barefoot – Male No.1 (b)
Barefoot – Male No.2 (b)
Impedance – magnitude (dB)
70
Shoes – Male No.1 (b)
Shoes – Male No.2 (b)
60
50
40
30
10
100
1000
10 000
Frequency (Hz)
Figure 3.36
Measured heel-impedances for male and female walkers with different footwear. Measured data are reproduced with
permission from (a) Watters (1965) and (b) Scholl (2001).
80
Measured heel-impedance:
minimum & maximum values
Mass: 0.5 kg (tapping machine
hammer)
70
Impedance – magnitude (dB)
Mass: 0.08 kg
Stiffness: 70 000 N/m
60
50
40
30
10
100
1000
10 000
Frequency (Hz)
Figure 3.37
Range of measured heel-impedances from Fig. 3.36 compared with mass and stiffness impedances. Measured data are
reproduced with permission from Watters (1965) and Scholl (2001).
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Chapter 3
m2
m1
F
v
F
v
Figure 3.38
Illustration of the mass–spring–mass system used to represent the heel-impedance ‘seen’ by the floor.
iωm1
v
m2
k
R
k
iω
iωm2
F
R
v
m1
F
Figure 3.39
Equivalent electrical circuit used to calculate the input impedance of a mass–spring–mass system representing the heelimpedance of a walker in shoes.
heel-impedances. Therefore for footsteps we can say that Zh = Zr . However, the heelimpedance can be approximately represented by a stiffness of 70 000 N/m in the low-frequency
range, and a mass of 0.08 kg in the mid-frequency range; these values are purely indicative
because of the wide range in measured heel-impedances. These lump elements illustrate the
general trend that is apparent in the individual measurements (Fig. 3.36) where the impedance
in terms of 20 lg|Z| initially decreases by 6 dB per octave to a minimum value before increasing
by 6 dB per octave.
These lump elements of mass and stiffness can be used to create a mass–spring–mass system
that is representative of the heel-impedance ‘seen’ by the floor as shown in Fig. 3.38 (Scholl,
2001; Warnock, 1983). The input impedance at the heel is calculated using an equivalent
electrical circuit (see Fig. 3.39) by applying a force to the mass, m1 , that impacts upon the floor,
k
iωm2
+R
F
iω
= iωm1 +
(3.108)
Z=
k
v
+R
iωm2 +
iω
A minimum value in the input impedance occurs at the resonance frequency of this mass–
spring–mass system; we need this resonance to correspond to the minimum value that occurs
in the heel-impedance spectrum. However, the spring needs to be highly damped (R term) with
an equivalent loss factor of 0.8 to avoid very low input impedance at this resonance frequency.
The input impedance below the resonance frequency is primarily determined by the spring
stiffness although there will also be a peak due to an anti-resonance of the mass–spring–mass
system. Above the resonance frequency the input impedance is primarily determined by the
mass, m1 , that impacts upon the floor, hence we assign a value of 0.08 kg to m1 . Mass m2
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80
Mass: 0.08 kg
Stiffness: 70 000 N/m
Impedance – magnitude (dB)
70
m1 ⫺
k
⫺ m2
(0.08 kg) (70 000 N/m) (3 kg)
60
50
40
30
10
100
1000
10 000
Frequency (Hz)
Figure 3.40
Impedance of a mass–spring–mass system used to represent the heel-impedance.
is then chosen so that the mass–spring–mass resonance frequency occurs close to the point
where the impedance of the 0.08 kg mass equals the impedance of the 70 000 N/m stiffness.
By using a value of 3 kg for m2 , the mass–spring–mass resonance frequency is 151 Hz and
the peak due to the anti-resonance is shifted well-below 50 Hz to a frequency of 24 Hz. The
impedance for this mass–spring–mass system is shown in Fig. 3.40. It is clear that if the mass
of the tapping machine hammer were to be replaced by a hammer that was effectively a mass–
spring–mass system, then an attempt could be made to simulate an idealized heel-impedance.
Warnock (1983, 2000) effectively implemented a hammer based on a mass–spring–mass system in a machine that was used to simulate the force pulse generated by footsteps. As with
real footsteps, the force levels from this machine were low, which meant that background noise
often affected measurement of the impact sound pressure level.
We now have a mass–spring–mass system for a heel-impedance that can be used to look at
the impedance ratio. This ratio is shown in Fig. 3.41 for four different plates commonly used
for the upper surface of a floor or floating floor. The concrete and screed plates have relatively
high impedances, hence the ratio is approximately 0 dB across the building acoustics frequency
range because Zh and Zr have negligible effect. The chipboard and OSB plates have much
lower impedances, so the frequency-dependent values of Zh and Zr have a significant effect,
and cause the ratio to vary with frequency. For many floor surfaces, Zdp will also vary with
frequency in the low- and mid-frequency ranges where the assumption of an infinite plate is
not appropriate (Section 2.8.3). This will cause the impedance ratio to become a more complex
function of frequency that is specific to particular types of floors with specific dimensions.
To summarize, we have seen that both the force ratio and the impedance ratio can be quite
complex functions of frequency with large variations that depend upon the type of floor surface.
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Chapter 3
2
10 lg (Re{1/(Zdp⫹Zh)}/Re{1/(Zdp⫹Zr)}) (dB)
0
⫺2
⫺4
⫺6
⫺8
⫺10
140 mm concrete slab (308 kg/m2)
⫺12
65 mm sand–cement screed (130 kg/m2)
⫺14
22 mm chipboard (17 kg/m2)
15 mm OSB (9 kg/m2)
⫺16
50
80
125
200
315
500
800 1250 2000
One-third-octave-band centre frequency (Hz)
3150
5000
Figure 3.41
Impedance ratio for four different plates.
Hence, the ratio of the power inputs will also be a relatively complex function of frequency.
This allows a conclusion to be drawn relating to one specific impact source, footsteps. The
implication is that the normalized (or standardized) impact sound pressure level in individual
frequency bands will not always correctly rank order all types of floor (with or without floor
coverings) in terms of their impact sound insulation against footsteps. However, by calculating
a single-number quantity from the individual frequency bands we will soon see that good
correlation can be achieved with subjective assessment of impact sound insulation in dwellings.
Similarly, the improvement of impact sound insulation in individual frequency bands will not
always correctly rank order all types of floor covering in terms of their impact sound insulation
against footsteps.
The advantages and disadvantages of the ISO tapping machine were known before it was
adopted in the Standards. Subsequent studies to investigate potential improvements have
generally produced one of the following options: (1) change the ISO tapping machine to simulate one particular type of impact, such as footsteps or a heavy impact from a child jumping, or,
(2) keep the ISO tapping machine but use a rating system that combines the frequency band
levels to produce a single-number quantity that correlates well with the subjective evaluation
of impact sound insulation.
3.6.3.4 Modifying the ISO tapping machine
Proposals to change the ISO tapping machine have generally been to make modifications to
the hammer impedance so that it is more representative of footsteps, or to change the impact
source completely. Regarding the former, it has been suggested that a resilient material (acting
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as a spring) could be inserted between the face of the metal hammer and the floor (e.g. see
Gerretsen, 1976; Scholl, 2001; Schultz, 1975). This modification to the hammer impedance has
been incorporated into the relevant Standard for laboratory measurement of the improvement
of impact sound insulation due to floor coverings on lightweight floors (ISO 140 Part 11).
Springs can either be attached to the ISO tapping machine hammers or a resilient material can
be placed on the floor underneath the hammers. This approach is based on representing the
heel-impedance by a mass–spring–mass system. However, it assumes that the floor covering
is a rigid plate with a sufficiently high mass that the lower mass in the mass–spring–mass
system can be omitted, i.e. m1 in Fig. 3.39 (Scholl, 2001). The resilient material significantly
reduces the impact sound pressure level in the mid- and high-frequency ranges, therefore low
background noise levels are important to ensure accurate measurements. A practical issue in
using a resilient material is that its dynamic properties must not vary with time (either during
the test or after repeated use) and temperature. An advantage of the steel hammers of the
ISO tapping machine hammers is that they are generally hard wearing, although their diameter
and curvature must be checked periodically. Note that the proposals described above to use
a resilient material are different to the rubber coating that was previously in old versions of
the measurement Standard (now superseded) which was to prevent the hammers damaging
fragile floor coverings (ISO 140 Part 6:1978).
Other investigations have looked at completely different impact sources to the ISO tapping
machine. Based on the work by Watters (1965), a mechanical machine was built to simulate
the impact made by the heel of a shoe with interchangeable shoe types, although unwanted
mechanical noise and vibration limited its application (Josse, 1970). To represent heavy, soft
impacts on floors, research in Japan has investigated the use of rubber balls and tyres; these
are discussed in Section 3.6.4.
3.6.3.5 Rating systems for impact sound insulation
The alternative to changing the ISO tapping machine has been to make alterations to the
rating system that combines levels from individual frequency bands to produce a single-number
quantity. The aim being that it should correlate well with subjective evaluation of impact sound
insulation. This approach should not be viewed as ‘correcting’ features associated with the
tapping machine such as the effects of hammer impedance and non-linearity with some soft
floor coverings. However, rating systems that place emphasis on the low-frequency range may
fortuitously avoid some of the problems with non-linearity that tend to be more apparent in the
mid- and high-frequency ranges.
A rating system can form a relationship between objective measurements with a standardized
impact source, and the subjective evaluation of impact sound insulation. We expect this to
be a complex relationship, partly because of the variety of structure-borne sound sources
in dwellings that range from footsteps to washing machines, but also due to the complexity
in the subjective evaluation of impact sound. It is reasonable to assume that there will be a
relationship between acceptable impact sound pressure levels and background noise due to
a masking effect. However, there are also indications that impact noise in dwellings tends to
disturb the occupant by startling them, which becomes important for sleep disturbance (Raw
and Oseland, 1991). Although footsteps on floors and stairs are often listed as an impact source
that is heard (Grimwood, 1997), there is uncertainty as to whether annoyance from footsteps
is linked to the loudness, or to the fact that they are detectable (Gerretsen, 1976). It is also
reasonable to assume that average subjective evaluations will vary between different countries
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Chapter 3
and cultures. We are therefore expecting rather a lot from a rating system that essentially relates
impact sound pressure levels using the ISO tapping machine to a subjective rating via a ‘black
box’ of complex relationships.
There have been several proposals for rating systems (e.g. see Choudhury and Bhandari,
1972; Fasold, 1965; Gerretsen, 1976). Some countries have implemented their own rating
systems, but here we will only discuss the rating method described in ISO 717 Part 2. This
Standard describes calculation of the single-number quantities, Ln,w , L′n,w , and L′nT ,w along
with spectrum adaptation terms, CI , CI,50–2500 for one-third-octave-bands, and CI,63–2000 for
octave-bands.
We start by looking at the link between objective and subjective evaluation of impact sound
insulation. Field measurements by Bodlund (1985) in Sweden were used to compare different
rating methods with subjective evaluations of the impact sound insulation in dwellings. This
data set comprised 14 different groups of housing with 22 different construction types (concrete
and timber joist floors with a variety of floor coverings).
The subjective rating used a seven-grade scale, where one was “Quite unsatisfactory’’ and
seven was “Quite satisfactory’’. For each single block of multi-storey apartment houses, or
residential block with a particular construction, the mean subjective score and the mean singlenumber quantity for the impact sound insulation was used rather than attempting to relate each
individual score from an interviewee to a specific sound insulation measurement. We have to
accept that it is difficult to relate objective and subjective ratings on an individual basis, but that
relationships can be found by grouping the objective and subjective ratings to calculate mean
values.
The correlation coefficient, r , between the mean weighted normalized impact sound pressure
level, <L′n,w >, and the mean subjective score was found to be 75% with the following straightline relationship,
L′n,w = 80.6 − 5.48 X
(3.109)
where X is the mean subjective score (1–7).
Hence L′n,w (and also L′nT ,w ) can be considered as adequate descriptors when we need to
define performance standards for impact sound insulation in the field. However, to improve the
correlation, Bodlund proposed an alternative to the rating curve in ISO 717 Part 2; this consisted
of a straight line between the 50 and 1000 Hz one-third-octave-bands with a positive gradient
of 1 dB per one-third-octave-band. Bodlund’s rating curve is significantly different to the ISO
rating curve that is used to calculate Ln,w , L′n,w , or L′nT ,w as can be seen from Fig. 3.42. The ISO
rating curve tends to emphasize the insulation in the mid- and high-frequency ranges. Bodlund
placed greater emphasis on the low-frequency range, but used the procedure described in ISO
717 Part 2 to shift the reference curve to determine a new single-number quantity, denoted
here as L′B,w . This emphasis on low frequencies is necessary because of the low-frequency
content of common impacts such as footsteps.
Footsteps on floors that have a soft floor covering, and/or a floating floor tend to generate the
highest sound pressure levels in the low-frequency range (Bodlund, 1985; Warnock, 2000).
Although there are wide variations between different walkers with different footwear (refer back
to Fig. 3.23), the general trend from a male walker in shoes with hard-heels can be seen in
measurements from Warnock (2000). These data were collected in the laboratory with one
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I n s u l a t i o n
ISO 717-2:1996
Bodlund
(dB)
5 dB
50
80
125 200 315 500 800 1250 2000 3150 5000
One-third-octave-band centre frequency (Hz)
Figure 3.42
Rating curves used to calculate single-number quantities for impact sound insulation. Permission to reproduce extract from
ISO 717-2 is granted by BSI on behalf of ISO.
male walker in one pair of shoes walking on different floor constructions. The maximum timeweighted sound levels generated by the walker in the receiving room below are shown in
Fig. 3.43. Without a soft floor covering or a floating floor, the timber floors gave significantly
higher levels than the concrete slab below 100 Hz. For both timber and concrete floors with a
soft floor covering or a floating floor, the timber floors had significantly higher levels than the
concrete slab below 50 Hz. Above 50 Hz there was a wide range of sound pressure levels from
the different floor constructions and it is difficult to identify a trend.
Returning to Bodlund’s work, the mean single-number quantity, L′B,w , calculated using
Bodlund’s rating curve which emphasized the low-frequency range showed stronger correlation with the mean subjective score (r = 87%) than L′n,w . The results are shown in Fig. 3.44
and give the following straight-line relationship,
L′B,w = 86.3 − 5.53 X
(3.110)
Bodlund found that when the mean subjective score was less than 4.4, at least 20% of the
interviewees rated the performance with a score less than 3. A mean subjective score of 4.4
corresponded to 51% of interviewees rating the performance with a score higher than 4. Bodlund proposed that mean subjective scores less than 4.4 should be deemed as unsatisfactory.
Hence the straight-line relationships described above can be used to help identify suitable
performance standards for impact sound insulation; although consideration should be given as
to whether subjective scores from one country (in this case, Sweden) can be applied to other
countries.
ISO 717 Part 2 does not directly implement Bodlund’s rating system; however, it can be used to
assess the single-number quantities that are defined in this Standard. ISO 717 Part 2 defines
spectrum adaptation terms, CI , or CI,50–2500 such that when they are added to Ln,w , L′n,w , or L′nT ,w ,
the resulting single-number quantity is equal to the energetic sum of the one-third-octave-band
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Chapter 3
(a) Without a soft floor covering or floating floor
Maximum time-weighted sound level (dB)
100
150 mm concrete slab (N ⫽ 1)
90
Timber joist or truss floors (N ⫽ 8)
80
70
60
50
40
30
20
10
0
25
40
63
100 160 250 400 630 1000
One-third-octave-band centre frequency (Hz)
(b) With a soft floor covering or floating floor
Maximum time-weighted sound level (dB)
100
90
150 mm concrete slab with a soft floor covering
or floating floor (N ⫽ 5)
80
Timber joist or truss floors with a soft floor
covering or floating floor (N ⫽ 25)
70
60
50
40
30
20
10
0
25
40
63
100 160 250 400 630 1000
One-third-octave-band centre frequency (Hz)
Figure 3.43
Maximum time-weighted sound levels (time constant, τ = 35 ms) in the receiving room from a male (approximate weight of
83 kg) walking in leather soled shoes with a hard rubber tip on the heels. Measured data are shown for floors with and without
a soft floor covering or floating floor. Each graph shows the average of N measurements. When N > 1, the minimum and
maximum levels from the set of measurements are shown using the same line style (solid or dotted lines) as the average
values but without symbols. Measured data are reproduced with permission from Warnock (2000) and the National Research
Council of Canada.
values (Ln , L′n , or L′nT ) minus 15 dB. Hagberg (1996) has assessed this use of the spectrum
adaptation term by using Bodlund’s rating system; this was done on the basis that if an alternative single-number quantity, such as L′n,w + CI,50–2500 , has a strong correlation with L′B,w , then
it is reasonable to assume that it will also have a strong correlation with subjective evaluation.
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80
r ⫽ 87%
Mean L⬘B,w (dB)
70
W
W W
W
W
C CCC
C
W
C
60
C
W
W
W W
50
C
M
M
C
W
40
1
2
3
4
5
Mean subjective score
6
7
Figure 3.44
80
80
70
70
L⬘B,w (dB)
L⬘B,w (dB)
Correlation between objective and subjective assessment of the impact sound insulation, where the subjective score and L′B,w
are both mean values for a single block of multi-storey apartment houses or a residential block. The subjective rating used
a seven-grade scale where 1 = Quite unsatisfactory and 7 = Quite satisfactory. Data are reproduced with permission from
Bodlund (1985).
Legend: , separating floors in multi-storey apartment houses; •, horizontal impact sound insulation for floors in attached
houses and multi-storey apartment houses; W, timber joist floor; C, concrete floor; M, mixed floor structures (concrete ground
floor and lightweight second floor in two-storey attached dwellings).
60
r ⫽ 76%
60
r ⫽ 96%
50
50
40
40
30
40
50
60
L⬘n,w (dB)
70
80
30
40
50
60
70
L⬘n,w ⫹ CI,50–2500 (dB)
80
Figure 3.45
Correlation between L′n,w and L′n,w + CI,50-2500 with Bodlund’s single-number quantity, L′B,w . Data are reproduced with permission
from Hagberg (1996).
Hagberg used 146 measurements to investigate the correlation between L′n,w and L′B,w , as
well as L′n,w + CI,50–2500 and L′B,w ; the results are shown in Fig. 3.45. For the straight-line
relationship between L′n,w and L′B,w the correlation coefficient, r , is 76%. There is significantly higher correlation (r = 96%) between L′n,w + CI,50–2500 and L′B,w due to emphasis on the
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Chapter 3
low-frequency range. The correlation between L′n,w + CI and L′B,w is slightly lower (r = 90%)
because L′n,w + CI does not use one-third-octave-bands below 100 Hz.
The average value of CI,50–2500 is −3 dB for concrete floors with a typical range of −11 to +1 dB,
compared to an average of +2.5 dB for timber floors with a typical range of −2 to +13 dB
(Rindel and Rasmussen, 1996). Rindel and Rasmussen note that there are potential problems
in using high negative values of the spectrum adaptation term. These typically occur with bare
concrete floors, or concrete floors with non-resilient floor coverings for which L′n,w + CI,50–2500
will not adequately account for the relatively high impact sound pressure levels that occur at
high frequencies. However L′n,w + CI,50–2500 can still be used to make a stronger link between
the objective and subjective rating of any floor that has a resilient floor covering or floating floor.
3.6.3.6 Concluding discussion
Despite all the criticisms that can be (and have been) made about the ISO tapping machine
over the years, its use in building regulations has certainly improved the impact sound insulation of the building stock and identified constructions with poor impact sound insulation. From
a regulatory point of view, there are pragmatic reasons to maintain the use of the ISO tapping
machine and rely on a rating system to make the link between the subjective and objective
ranking of impact sound insulation in dwellings. This approach ensures that any historical
database of measurements does not become redundant because it allows measurement data
to be re-processed with any new rating procedure. The ISO tapping machine also allows accurate measurements in the field by producing sound pressure levels that are usually well-above
background noise levels. This is essential for regulations that require field tests to demonstrate compliance with a performance standard. For regulations on impact sound insulation
in dwellings the choice of single-number quantity depends on what the regulation is aiming to
achieve and how the regulation is implemented and enforced. As with airborne sound insulation
this means that different countries may need to use different single-number quantities.
For engineers involved in floor design and the measurement of impact sound insulation (as
well as manufacturers designing flooring elements) it is necessary to be aware of what can,
and what cannot be inferred from measurements using the ISO tapping machine. It is therefore
useful to have a general understanding of the tapping machine and the background behind the
rating system.
Regulations often specify performance standards to be achieved in field measurements using
the ISO tapping machine. In this case the situation is relatively straightforward. Laboratory
measurements using the ISO tapping machine can be used to aid design decisions and can
be incorporated into models that include flanking transmission for impact sound insulation (see
Section 5.4.2). However, the situation can be different with bespoke designs; for example when
establishing a suitable level of impact sound insulation below a dance studio or a sports hall
in a multi-storey building. The existence of the ISO tapping machine and the ISO 717 Part 2
rating system make it convenient to use this source to specify the required level of impact
sound insulation and to demonstrate compliance with measurements in the finished building.
This may not always be appropriate. When assessing the ability of different floors to provide
insulation against specific impact sources by using the ISO tapping machine, it is necessary
to be aware of its limitations. The two main issues are the effect of the hammer impedance on
the power input in comparison with the impedance of the real impact source, and the existence
of non-linearity with some soft floor coverings. Over the years, some floating floors and floor
coverings have inevitably been designed to take advantage of the physics of the ISO tapping
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machine, rather than to attenuate real impact sources. In some cases it can be beneficial
to take laboratory measurements of the impact sound insulation with more than one impact
source, for example with the ISO tapping machine and a heavy impact source such as the ISO
rubber ball.
3.6.4 Heavy impact sources
Impacts from people on floors can generally be categorized as either light, hard impacts (e.g.
footsteps in hard-heeled shoes), or heavy, soft impacts (e.g. children running and jumping,
adults exercising, footsteps in bare feet). We have previously discussed the fact that the ISO
tapping machine does not simulate any specific impact source, but that the ISO rating system provides a link between the subjective and objective rating of impact sound insulation in
dwellings. In previous surveys of impact sound insulation in dwellings (e.g. Bodlund, 1985) it is
reasonable to assume that both light and heavy impacts occurred on a daily basis, so the subjective assessments are likely to have considered both types of impact. Hence, by using the ISO
tapping machine and a rating system that gives a single-number quantity that correlates well
with the subjective assessment, then to some unknown extent, both light and heavy impacts
are taken into account. However, an issue arises purely because of this unknown extent, and
the fact that the ISO tapping machine is not well-suited to ranking the impact sound insulation of floors against specific heavy impacts. In countries such as Japan, hard-heeled shoes
are not worn in dwellings and the majority of dwellings are of lightweight construction. Heavy,
soft impacts on lightweight timber or steel frame floors tend to give rise to higher impact sound
pressure levels than heavyweight concrete floors; although there are inevitably exceptions due
to the many different types of lightweight floor construction. To allow measurement of heavy,
soft impacts, Japan has developed two impact sources, these are the rubber ball and the bang
machine (see Fig. 3.46).
The bang machine consists of a tyre dropped from a height of 0.9 m. The test method and the
bang machine specification are described in Japanese Industrial Standard JIS A 1418-2. It is
well-suited to laboratory measurements, but less suitable for field tests as it produces heavy
blows that could cause minor damage to decorated dwellings. In addition, it is less convenient
to transport than the ISO tapping machine.
During the development of the rubber ball, different versions were produced to modify the
characteristic of the impact force (Tachibana et al., 1998). The final specification for the ISO
rubber ball is described in the relevant Standard (ISO 140 Part 11). The ISO rubber ball is a
hollow sphere of 30 mm thick silicone rubber with an outer diameter of 180 mm. It has a weight
of approximately 2.5 kg and is dropped from a height of 1 m above the floor, measured from
the lower surface of the ball.
During the development of heavy impact sources, there was a wide range of equipment that
could accurately measure peak, impulse, or maximum time-weighted sound levels. This is in
contrast to the early development period for the tapping machine when it was not possible
to accurately measure impulses (Schultz, 1981). This led to the design of a tapping machine
that produced a continuous signal from a train of hammer impacts. However, when people
provide a subjective assessment of footsteps and assess the ability of background noise to
mask the noise from footsteps, the A-weighted peak, impulse (τ = 35 ms), or maximum timeweighted (Fast, τ = 125 ms) sound level generally gives better correlations than the equivalent
continuous sound level (Ford and Warnock, 1974; Hamme, 1965; Olynyk and Northwood,
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Figure 3.46
Heavy impact sources. Left: Rubber ball – manually dropped from a height of 1 m. Right: Bang machine. Photo provided by
Dr J. Yoshimura at the Kobayasi Institute of Physical Research, Japan.
1965, 1968; Warnock, 1983). Nowadays the maximum time-weighted sound level, peak sound
level, and impulse sound levels are all clearly defined as different descriptors (IEC 61672
Part 1); however, there is sometimes ambiguity about their usage in the past.
When a multi-modal system (e.g. a room) is excited by an impulse, the first peak (or trough) in
the instantaneous measured response will not necessarily have the largest magnitude (Brüel,
1987). For this reason the maximum time-weighted sound level is determined during a time
interval starting from just before the impact, to a time when the response has decayed to a
negligible level. For each impulse that is produced from a single drop of the ISO rubber ball, the
maximum time-weighted (Fast) sound level, LFmax , is measured in the receiving room. This is
averaged for a number of excitation positions and stationary microphone positions to give the
impact sound pressure level, Li,Fmax (ISO 140 Part 11). Some proposals for short test methods
to measure the impact sound pressure level from heavy impact sources have used a single
stationary microphone position, often near the centre of the floor at mid-height in the room;
however, measured data indicates significant spatial variation in the sound field in the low- and
mid-frequency ranges (Broch, 1983; Warnock, 2000). It is therefore necessary to measure at
several different microphone positions to calculate the spatial average level.
Unlike the impact sound pressure level measured with the ISO tapping machine, the level
measured with the rubber ball is not normalized to the absorption area, or standardized to the
299
S o u n d
I n s u l a t i o n
100
ISO tapping machine
Impact sound pressure level (dB)
90
Bang machine
Rubber ball
80
Child jumping off a chair
70
60
50
40
30
20
31.5
63
125
250
500
1000 2000
Octave-band centre frequency (Hz)
4000
Figure 3.47
Impact sound pressure levels in a receiving room underneath a timber floor due to the ISO tapping machine, bang
machine, rubber ball, and a child jumping off a chair. Measured data are reproduced with permission from Tachibana et al.
(1998).
reverberation time of the receiving room (ISO 140 Part 11). Normalization or standardization
is essential for the comparison of tests made with the ISO tapping machine, and to allow
calculation of the impact sound pressure level in different rooms from laboratory test data.
However, peak sound levels from an individual impact tend not to be significantly affected
by the absorption area or reverberation time of the receiving room (Hamme, 1965, Ford and
Warnock, 1974, Schultz, 1975). The same tendency has been assumed for maximum timeweighted (Fast) levels. Hence, normalization or standardization is not carried out. Another
factor is that the reverberation times in laboratory receiving rooms are not excessively long
or short; typically being between 1 and 2 s (ISO 140 Part 1) which also minimizes variation
between laboratories.
Figure 3.47 allows comparison of the impact sound pressure levels from a timber floor with the
ISO tapping machine, bang machine, rubber ball, and a child jumping off a chair (Tachibana
et al., 1998). In a similar way to the ISO tapping machine, the bang machine and the rubber
ball do not simulate one specific type of heavy impact, although the rubber ball can give similar
impact sound pressure levels to a child jumping off a chair. With heavy impacts on lightweight
floors, the impact sound pressure levels are usually highest in frequency bands below 100 Hz.
For this particular floor, the spectrum shape for the ISO tapping machine is distinctly different to
the spectrum for heavy impacts. When trying to compare different floor constructions in terms
of their impact sound insulation specifically against heavy, soft impacts it will be more reliable
to use a heavy impact source.
Heavy impact sources are useful measurement tools for lightweight floors. However, because
of the complexity in determining their input force, and the use of a maximum time-weighted
sound level, there is no simple method for the prediction of the impact sound pressure level as
there is with the ISO tapping machine.
300
Chapter 3
3.6.5
Other measurement issues
Although most issues relating to the ISO tapping machine have been covered, there are some
specific practical issues, such as the effect of time dependency on the impact sound pressure
level and the effect of dust and dirt under the tapping machine hammers. These are discussed
here along with more general aspects such as the size of the test specimen and static load.
3.6.5.1 Background noise correction
See discussion in Section 3.5.4.1.
3.6.5.2 Converting to octave-bands
Impact sound insulation descriptors in one-third-octave-bands can be converted to octavebands using the three one-third-octave-bands that form the octave-band,
3
(3.111)
10XTOB,n /10
XOB = 10 lg
n=1
where X represents Ln , LnT , etc.
The improvement of impact sound insulation is converted using:
3
1 −LTOB,n /10
LOB = −10 lg
10
3
(3.112)
n=1
3.6.5.3 Time dependency
After the tapping machine is switched on, the physical properties of some floor surfaces and
floor coverings change with time. It is usually a matter of minutes before the properties stabilize,
although in some cases, a steady-state is never reached and the sound pressure levels continue to vary with time. The relevant Standards (ISO 140 Parts 6 & 7) require the sound pressure
level to have reached a steady-state before measurements are taken; and if a steady-state is
not reached, then an appropriate measurement period should be established.
Time dependency occurs with some soft floor coverings. During each impact the tapping
machine hammers can increase or decrease the contact stiffness and damping properties
of the soft floor covering directly underneath the hammers. Hence the power input from the
tapping machine changes with time due to variation in the physical properties of the floor covering. This can occur with materials such as carpet (Michelsen, 1982) or felt-backed PVC
(Bodlund and Jonasson, 1983). An example is shown in Fig. 3.48 for carpet on a concrete
slab where the sound pressure level was measured using successive 5 s samples (Michelsen,
1982). After the tapping machine was switched on, the sound pressure levels initially increase
with time in the low- and mid-frequency range before stabilizing after ≈4 min.
Time dependency also occurs when the tapping machine causes damage to the surface of the
floor during the test. This can occur with plate materials where the hammers cause indentations,
or where a fragile surface finish breaks up due to the force of the impacts. Fournier and Val
(1963) quote an example of a concrete floor with a fragile crust on the surface. The tapping
machine hammers damaged the crust of the floor causing it to crack and break over a period
301
S o u n d
I n s u l a t i o n
1 dB
Sound pressure level, Lp(t) in 5 s samples (dB)
125 Hz
250 Hz
500 Hz
1000 Hz
5
35
65
95
125 155 185
Time, t (s)
215
245
275
Figure 3.48
Variation of the one-third-octave-band sound pressure level in the receiving room with an ISO tapping machine on top of a
carpet on a concrete floor slab. Consecutive samples were measured with an integration time of 5 s, starting at the point that
the tapping machine was switched on. Measured data are reproduced with permission from Michelsen (1982).
of a few minutes. Significant damage from the tapping machine is less common when building
materials have been designed to pass durability tests. However, the floor surface should be
checked before and after the measurement for signs of damage that could have affected the
results. In addition the sound pressure level shown on the analyser should be monitored during
the test for any signs of a change.
Another factor that can vary the impact sound insulation over longer periods of time is the
floor temperature; this is relevant when floors have an underfloor heating system. A significant
effect can occur with soft floor coverings comprising foams and plastic coverings (Bodlund and
Jonasson, 1983) where an increase in the floor temperature changes the stiffness and damping
properties of the covering. This results in a change to the impact sound pressure level.
302
Chapter 3
3.6.5.4
Dust, dirt, and drying time
Building work is rarely a clean process, and impact sound insulation tests in the field are not
always carried out before the floors have been swept of wood shavings, sand, etc. This is
relevant to measurements with the ISO tapping machine because excessive amounts of dust
and dirt underneath the hammers can affect the results. The preceding discussion on the
tapping machine indicated that the contact area between the hammer and the floor, as well as
the contact stiffness affects the force pulse delivered by the hammer. Hence, dust and dirt can
change the power input from the tapping machine into the floor. Once the tapping machine is
switched on, any effect will initially change with time as the floor vibration causes the dust and
dirt to move around. An indication of the effect of dust and dirt on two different floors is shown
in Fig. 3.49. At the point that the tapping machine was switched on, the layer of dust and dirt
was approximately 2 mm deep. Its effect is to reduce the sound pressure level in the mid- and
high-frequency range. The effect on the single-number quantity can be significant with some
floors, such as the concrete slab, but negligible for others.
As with airborne sound insulation, the impact sound insulation can vary depending on how
much the test element has dried out. For a walking surface such as a floating floor screed,
the contact stiffness changes as it dries out; this affects the power input from the ISO tapping
machine. It is also worth noting that the stiffness and damping properties of adhesive used to
fix floor tiles or other coverings can also change as it dries.
3.6.5.5 Size of test specimen
With soft floor coverings, such as carpet or vinyl flooring, it is appropriate to take measurements
using small samples that are large enough to support the ISO tapping machine (ISO 140 Parts
8 & 11). However, the results can differ depending on whether the covering is loose-laid or if
a fixing adhesive is used (Bodlund and Jonasson, 1983).
Floating floors need to cover the complete surface of the reference floor (ISO 140 Parts 8 & 11).
The intention is that the measurement should only consider structure-borne sound transmission
from the tapping machine into the reference floor via the floating floor. If a small area of floating
floor is used, unwanted flanking transmission can occur via the uncovered area of the reference
floor (lightweight or heavyweight). The floating floor that is excited by the tapping machine not
only transmits structure-borne sound via the resilient material into the reference floor but also
radiates sound into the source room (i.e. the room containing the tapping machine) which, in
turn, excites the uncovered area of the reference floor. Using only a small area of floating floor
tends to reduce the measured improvement in the sound pressure level in the mid- and highfrequency ranges. A full-size floating floor is also necessary to ensure that the driving-point
mobility (which determines the power input from the tapping machine) is representative of the
full-size floor. This may not occur with small areas of flooring that have a much lower modal
density than the full-size floating floor.
3.6.5.6 Static load
Floating floors that provide impact sound insulation usually incorporate a resilient material.
Therefore the static load that occurs in situ (such as from furniture or office equipment) can significantly alter the performance of the floating floor. This is usually more relevant to lightweight
floating floors than heavier ones such as screeds.
303
S o u n d
I n s u l a t i o n
Normalized impact sound pressure level, Ln (dB)
(a) 140 mm concrete floor slab
80
70
60
50
40
30
20
Clean 79 (⫺13)
10
Initial layer of 2 mm dust and dirt 74 (⫺10)
0
50
80
125 200 315 500 800 1250 2000 3150 5000
One-third-octave-band centre frequency (Hz)
Normalized impact sound pressure level, Ln (dB)
(b) Timber floating floor on a 140 mm concrete floor slab
80
Clean 49 (1)
70
Initial layer of 2 mm dust and dirt 50 (1)
60
50
40
30
20
10
0
50
80
125 200 315 500 800 1250 2000 3150 5000
One-third-octave-band centre frequency (Hz)
Figure 3.49
Effect of dust and dirt on the impact sound insulation measured with the ISO tapping machine. The rating for the singlenumber-quantity (ISO 717 Part 2) is shown as Ln,w (CI ) on the legend. Measured data from Hopkins are reproduced with
permission from BRE Trust.
In current measurement Standards, simulation of normal furnishing is defined by a uniformly
distributed load of 20 to 25 kg/m2 (ISO 140 Parts 8 & 11). Note that in a previous version
of the Standard (now superseded), a loading of 100 kg/m2 was recommended (ISO 140 Part
8:1978). On lightweight floating floors, 20 to 25 kg/m2 typically reduces Lw by up to 2 dB,
whereas 100 kg/m2 can reduce Lw by up to 6 dB. It is therefore important to carry out impact
sound insulation tests with the static load relevant to the end use.
304
Chapter 3
A static load is most commonly provided by using an array of concrete blocks. These lump
masses need to be placed directly on the floor. If they are placed on small strips of resilient
material they effectively become an array of mass–spring resonators that can alter the response
of the floating floor; this interferes with the required assessment of static load.
3.6.5.7 Excitation positions
The power input from the ISO tapping machine partly depends on the driving-point mobility
of the floor. This varies across the surface of the floor so it is necessary to average several
different excitation positions. The driving-point mobility on ribbed or spatially periodic plates with
beams can differ between excitation positions on top of the ribs or beams compared to those
in-between them. When the positions of the ribs or beams are not known and a sufficient number
of random positions are used, this is partly overcome by positioning the tapping machine at
an angle of 45◦ to the ribs/beams (ISO 140 Parts 6 & 7). This is also beneficial because this
type of plate can have a strong modal response between the ribs or beams; hence, aligning
the tapping machine with the beams could bias the average.
When measuring impact sound insulation in the field it is also useful to be aware of floors that
show a significant decrease in vibration with distance across the floor (Section 2.7.7), because
flanking transmission via connected walls may depend on the position of the tapping machine.
3.7 Rain noise
In some countries consideration of rain noise inside buildings is only necessary for moderate
rainfall, whilst in others it is relevant to long rainy seasons with torrential downpours. As with
impact sound insulation on floors, standardizing an impact source for rainfall, requires a degree
of pragmatism. An artificial rainfall source in the laboratory needs to be linked to natural rainfall,
but it also needs to generate sufficiently high levels to allow measurements that are unaffected
by background noise.
3.7.1 Power input
To gain an overview of relevant parameters we start by looking at a raindrop as a structureborne sound source. The impact of a drop of water upon a surface can be considered in two
phases (Petersson, 1995). In the initial impact phase the mass of the drop remains unchanged
and there is rapid deceleration. This is followed by a flow phase in which the drop ‘breaks
open’ and the mass of the drop decreases. A falling drop is assumed to be initially spherical
but as it travels through the air its shape becomes distorted, although its volume is assumed to
remain constant. The force applied upon impact will depend on its distorted shape; however,
as its exact shape is uncertain it is necessary to adopt idealized drop shapes. For a drop shape
described by a paraboloid, the force pulse can be described by (Jagenäs and Petersson, 1986;
Suga and Tachibana, 1994):
⎧
⎫
8r ⎬
⎨ρ πr 2 v 2 1 − 3v0 t
for
0
≤
t
≤
w
0
F(t) =
(3.113)
8r
3v0
⎩
⎭
0
for all other t
where ρw is the density of water, r is the radius of the initially spherical drop, and v0 is the drop
velocity in the flow phase.
305
S o u n d
I n s u l a t i o n
(a) Paraboloid drop shape
8r/3
2r
(b) Force pulse from a single drop
Force, F(t)
ρw πr2(v0)2
0
Time, t
0
8r/3v0
Figure 3.50
Idealized raindrop.
The paraboloid drop and its associated triangular shape force pulse from Eq. 3.113 are shown
in Fig. 3.50. The Fourier transform of the force pulse (Eq. 3.113) is used to give the energy
spectrum, |F(f )|2 , from which the FFT line spectrum can be combined to give one-third-octave
or octave bands.
It is assumed that each raindrop falls upon a dry surface. In practice we are not really interested
in the first few raindrops on a dry roof, but from the sound radiated during steady rainfall. This
means there will be a layer of water over its surface and as most roof elements are sloped and
non-porous there will be some water flowing down the surface. The kinetic energy of a falling
drop can be transferred to an existing layer of water on the surface; this broadens the force
pulse with an increase in the force applied at low frequencies and a reduction at high frequencies
(Petersson, 1995). The effect of an existing layer of water is inherent in a measurement.
However, it is not included in the above calculation of the force pulse; we simply note that
306
Chapter 3
10
9
Terminal velocity (m/s)
8
7
6
5
4
3
2
1
0
0
1
2
3
4
5
Raindrop diameter (mm)
6
Figure 3.51
Terminal velocities for different raindrop diameters.
this introduces some additional uncertainty. A comparison of measurement and prediction in
Section 4.5 indicates that the effect can sometimes be considered as negligible.
For a falling drop, equilibrium between the external forces acting upon it results in a terminal
velocity, vT . For simplicity these external forces can be taken as gravity and aerodynamic
drag. Raindrops falling vertically in a calm atmosphere can be assumed to have reached their
terminal velocity before hitting the ground. To calculate the force pulse (Eq. 3.113) it can be
assumed that v0 equals vT for which the terminal velocity of a raindrop can be calculated from
the empirical equation (Best, 1950):
'6
5
&
D 1.147
(3.114)
vT = 9.58 1 − exp −
1.77
where the raindrop is assumed to be spherical with diameter, D in mm.
Raindrops with diameters larger than 1 mm can be classed as rain whereas smaller drops are
generally referred to as drizzle. Raindrop diameter depends upon the rainfall rate, temperature,
and humidity. In temperate climates there is rarely any need to consider drop diameters larger
than 5 to 6 mm; larger drops than this will break up into smaller drops on their way down. Terminal velocities are shown in Fig. 3.51 and generally increase with increasing raindrop diameter.
The terminal velocity tends towards a plateau for large drops because the aerodynamic drag
force increases due to the distorted drop shape.
For 2 and 5 mm drop diameters the minimum fall heights needed to achieve terminal velocity, when starting from an initial velocity of zero, are approximately 7 and 15 m respectively
(e.g. see McLoughlin et al., 1994). Height limitations in the laboratory mean that arranging
fall heights for artificial raindrops to reach terminal velocity is not always practical. Therefore
the average fall velocity is usually lower and is determined by measurement or calculation.
Drops generated from pressurized nozzles have an initial velocity, so a shorter fall height may
achieve terminal velocity.
307
S o u n d
I n s u l a t i o n
Water supply
Tank with
perforated base
Figure 3.52
Measurement set-up using a water tank to provide a median drop diameter within specified tolerances.
In the laboratory, artificial rain can be generated with a single drop diameter. There will
inevitably be some variation; hence, tolerances are given on the median drop diameter from
the artificial rain source (ISO 140 Part 18). A tank of water with a perforated base forms a
suitable source as shown in Fig. 3.52. The number of drops, N, that fall upon the excitation
area in 1 s can then be calculated from the rainfall rate, Rr , in mm/h using:
N=
3
Rr
3 600 000 4πr 3
(3.115)
Natural rainfall has a statistical distribution of raindrop diameters. This can be modelled using
the exponential distribution given by Marshall and Palmer (1948) in terms of n(D) in mm−1 m−3 ,
n(D) = 8000 exp (−4.1DRr−0.21 )
(3.116)
where D is the raindrop diameter in mm and Rr is the rainfall rate (i.e. rain intensity) in mm/h.
Figure 3.53 shows distributions of raindrop diameters for typical rainfall rates. This emphasizes
the paucity of large diameter raindrops in the distribution. However, it is not the volume of water
that is relevant, it is the power input into the structure; and this tends to be dominated by the
small fraction of larger diameter drops that apply high forces due to their high terminal velocities
(Ballagh, 1990). The Marshall–Palmer distribution of drop diameters can be used to calculate
the number of drops, N(D), with drop diameters between D and D + δD (in mm) that fall upon
the excitation area in 1 s,
N(D) = n(D)vT δD
(3.117)
As with the ISO tapping machine, the impedance of an impacting body can affect the power
input; although the effect is generally much less significant with a rain drop. Rather than treat
the rain drop as a lump mass like the tapping machine hammer, a flow impedance is used to
describe the raindrop impedance (Petersson, 1995).
Zdrop = ρw πr 2 vT
308
(3.118)
Chapter 3
10 000
Rainfall rate
4 mm/h
1000
n(D) (mm⫺1 m⫺3)
15 mm/h
40 mm/h
100
10
1
0.1
0
1
2
3
4
Raindrop diameter, D (mm)
5
6
Figure 3.53
Marshall–Palmer distribution of raindrop diameters for different rainfall rates.
For artificial rainfall with single diameter drops, the power input into a plate with a driving-point
impedance, Zdp that is real-valued (i.e. an infinite plate) is calculated using:
Win = NSe |F(f )|2
1
Zdp + Zdrop
(3.119)
where N equals the number of drops that fall upon the excitation area, Se in 1 s.
For natural rainfall the distribution can be described by minimum and maximum drop diameters,
Dmin and Dmax , with a chosen diameter step (δD). The power input is calculated by summing
the individual power inputs from Eq. 3.119 for each diameter in the distribution using N(D)
instead of N.
The measurement Standard for rain noise (ISO 140 Part 18) is based upon the rain type classifications for intense and heavy rain in IEC 60721-2-2. The latter Standard can be used to define
moderate, intense, and heavy rain based on upper limits for the rainfall rate, drop diameter,
and fall velocity. Hence moderate, intense, and heavy rain are defined as having rainfall rates
of 4, 15, and 40 mm/h respectively, median drop diameters of 1, 2, and 5 mm respectively, and
fall velocities of 2, 4, and 7 m/s respectively. These can be used to define artificial rain for use
in the laboratory where drops with a median drop diameter often fall at velocities that are lower
than their terminal velocity. It is logical to try and compare these with natural rain. However,
it is not possible to definitively describe natural rain; the rainfall rate varies during a period of
rainfall; often with an initially high rate followed by lower rates. Hence the duration and the
rainfall rate are important, not just the depth in millimetres that falls over 1 h. For comparative
purposes we can create one example of natural rainfall (there are many possibilities) by using
the same rainfall rates as the artificial rain, the Marshall–Palmer model for the drop diameter
distribution, and terminal velocities calculated using Eq. 3.114. For temperate climates we
will assume Dmin = 0 mm, Dmax = 5 mm, and δD = 0.1. Other examples of natural rain can be
309
S o u n d
I n s u l a t i o n
70
Glass
ρ ⫽ 2500 kg/m3
cL ⫽ 5200 m/s
h ⫽ 0.006 m
60
Power input (dB)
50
40
30
20
Artificial rain: heavy (40 mm/h, 5 mm, 7 m/s)
Artificial rain: intense (15 mm/h, 2 mm, 4 m/s)
Artificial rain: moderate (4 mm/h, 1 mm, 2 m/s)
Example of natural rain: heavy (40 mm/h)
Example of natural rain: intense (15 mm/h)
Example of natural rain: moderate (4 mm/h)
10
0
50
80
125
200
315
500
800 1250 2000
One-third-octave-band centre frequency (Hz)
3150
5000
Figure 3.54
Predicted power inputs for artificial and natural rain falling upon a 1 m2 excitation area on 6 mm glass.
created by changing Dmin and Dmax ; note that reducing Dmax can significantly reduce the power
input and therefore requires justification against real rainfall statistics for any design work.
Figure 3.54 shows the power input (Eq. 3.119) for rainfall on 6 mm glass using the above
descriptions of artificial and natural rain. For other plates with a frequency-independent drivingpoint impedance (i.e. infinite plates) and negligible drop impedance, the curves will have a
similar shape. As with other impulses, the force spectrum is flat up to a cut-off frequency above
which the power input decreases with increasing frequency. For one-third-octave-bands this
results in an increasing power input up to the cut-off frequency which occurs in the mid- or highfrequency range. There is some similarity between artificial heavy and the natural heavy rain
because the power input tends to be dominated by larger diameter drops. Using Dmax = 5 mm
means that the natural intense and natural moderate rainfall have higher power inputs than
their artificial counterparts. For this particular example of natural rain, using artificial heavy
and intense rain could give a reasonable indication of the range from moderate through to
heavy natural rain. However, with some roof elements it is only possible to measure with
heavy artificial rain in the laboratory in order that the sound pressure levels are well-above the
background noise level.
310
Chapter 3
3.7.2
Radiated sound
Two methods are described in the relevant Standard (ISO 140 Part 18) to determine the sound
intensity or sound power radiated from underneath the test element; sound pressure level
measurements in a reverberant room or sound intensity measurements.
For homogeneous elements the radiated power from the test element is proportional to the
excitation area. The excitation area used in the measurements may not cover the entire surface
hence it is useful to present values in terms of the sound intensity level, LI . For sound pressure
level measurements in a reverberant room this is calculated using:
Se
V
− 10 lg
(3.120)
− 14 dB
LI = Lp + 10 lg
T
S0
where V is the volume of the receiving room, T is the reverberation time in the receiving room,
Se is the excitation area, and reference area, S0 = 1 m2 .
For sound intensity measurements,
LI = LIn + 10 lg
Sm
Se
(3.121)
where LIn is the temporal and spatial average normal sound intensity level over the measurement surface.
Using Eq. 3.120 or 3.121, the sound power radiated by a homogeneous element when excited
over its entire surface area can be calculated using:
S
LW = LI + 10 lg
(3.122)
S0
where S is the surface area of the test element.
3.7.3 Other measurement issues
This section discusses the comparison of measurements from different roof elements and the
application of measured data to in situ installations.
Example rain noise measurements on 6 mm float glass and an Insulating Glass Unit (IGU)
formed from two panes of 6 mm glass are shown in Fig. 3.55. For the 6 mm glass there is a
small peak in the sound power at the critical frequency. The critical frequency peak is only just
discernible for the IGU, but there is a significant peak at the mass–spring–mass resonance
frequency. Roof elements formed from single sheets of metal, plastic, or glass often have
their critical frequency in the mid- or high-frequency range. Below the critical frequency, the
radiation efficiency depends on the baffle orientation, boundary conditions, and the plate
dimensions. If these factors differ significantly between measurements on different elements
and/or they differ from in situ, it makes comparisons between different elements more awkward
unless the effect of these differences can be modelled (see Section 4.5).
As with measurements of airborne or impact sound insulation in transmission suites, total loss
factor measurement also needs consideration for rain noise measurements on solid homogenous plates. With rain excitation, the sound power radiated by a solid homogeneous plate is
311
S o u n d
70
I n s u l a t i o n
6 mm glass
6-12-6 insulating glass unit
60
Solid lines: sound power
Dotted lines: total loss factor (TLF)
(Open symbols correspond to the TLF
in the measurement frame used in the
rain noise measurement. Shaded
symbols correspond to the measured
TLF when resiliently suspended)
Sound power (dB)
50
40
Artificial heavy rain
Rainfall rate: 40 mm/h
Median drop diameter: 5 mm
Fall velocity: ≈ 7 m/s
Surface area, S: 1.875 m2
Excitation area, Se: 1 m2
30
20
0
50
80
125
200
315
500
800 1250 2000
One-third-octave-band centre frequency (Hz)
3150
Total loss factor (dB)
10
106
104
102
100
98
96
94
92
5000
Figure 3.55
Radiated sound power from artificial heavy rain on glazing using sound intensity measurements according to ISO 15186
Part 1. Test element dimensions: Lx = 1.5 m, Ly = 1.25 m. 6 mm glass: ρs = 15 kg/m2 , fc = 2021 Hz, mounted in a wooden
frame with putty. 6-12-6 insulating glass unit (air filled): fmsm = 200 Hz, mounted in the same wooden frame but without putty.
Both elements were orientated at an angle of 30◦ for water drainage. All total loss factors were determined from structural
reverberation time measurements using MLS shaker excitation. All values are shown with 95% confidence intervals (values for
resiliently suspended 6 mm glass are only shown below the critical frequency). Measured data from Hopkins are reproduced
with permission from BRE and DfES.
dependent on the total loss factor of the plate. For this reason, damping compounds are sometimes applied to lightweight metal roofs to increase the internal loss factor, and therefore the
total loss factor in order to reduce the radiated power (note that commonly used metals tend
to have very low internal loss factors). If the mounting conditions change the total loss factor, this will also change the radiated sound power. To give an indication of the effect of the
mounting, the total loss factors were measured once in the measurement frame used for the
rain noise measurements and once when resiliently suspended. For the resiliently suspended
6 mm glass there are no structural coupling losses, only internal and radiation losses. The
internal and radiation losses are low so the total loss factor increases significantly once it is
installed in the frame where the coupling losses predominantly determine the total loss factor.
For the 6 mm glass that forms one side of the IGU, the increase is much smaller because each
pane of the IGU is tightly coupled to the other; hence, the structural coupling losses of each
pane are already quite high when it is resiliently suspended.
312
Chapter 3
Note that for laminated glass the bending stiffness and internal loss factor vary significantly
with surface temperature. During the measurement the radiated sound power is therefore
affected by the water temperature as well as room temperature. This affects the comparison
of measurements as well as use of the measurement to determine the radiated sound power
in situ. Compared to float glass of the same thickness this makes it possible for laminated glass
to radiate more sound power at some frequencies and less sound power at others; this will
depend on its temperature.
3.8 Reverberation time
In reverberant sound fields encountered in typical rooms, the decay curves vary throughout the
space and it is necessary to determine the spatial average reverberation time. In Section 1.2.6.3
we looked at idealized decay curves in rooms; these were straight lines or smooth curves that
were unaffected by random fluctuations or background noise. The ability to gain good estimates
of the reverberation time from measured decay curves is determined by the combination of the
acoustic system under test, the measurement procedure, signal processing, and evaluation of
the decay curve.
There are two main methods that are used to determine the reverberation time in spaces or the
structural reverberation time on structures; the interrupted noise method, and the integrated
impulse response method (ISO 3382). The derivations in this section tend to refer to sound
pressure as this is the most common application; but they are equally applicable to vibration.
For accurate sound insulation measurements in rooms, we typically need to determine the
reverberation time in seconds to two decimal places. For structural reverberation times the
decays are much shorter and three decimal places are often required.
3.8.1 Interrupted noise method
The interrupted noise method uses random noise to create a steady-state level. After a steadystate has been achieved, the excitation is stopped and the subsequent decay of the sound
pressure level is recorded. An example decay curve is shown in Fig. 3.56. Compared to the
idealized decay curve previously shown in Fig. 1.19, the curve is not smooth; it is characterized
by random fluctuations due to the random nature of the excitation signal. As the decaying signal
level gets closer to the background noise level, the slope of the decay curve is altered by the
background noise and it is no longer representative of the actual decay. When this occurs
within the evaluation range, it prevents accurate determination of the reverberation time. Any
significant effects on TX due to background noise can be avoided by using a steady-state
level that is at least 15 + X dB above steady background noise, preferably more, because
background noise is not always steady. For example, the steady-state level should be at least
30 dB above background noise for T15 or at least 45 dB above background noise for T30 . The
starting point for the evaluation range is 5 dB below the steady-state level, and therefore the
end point will be at least 10 dB above the background noise level.
If measurement of a decay curve is repeated using exactly the same microphone and loudspeaker positions, then each time that the excitation is stopped, the modes will all have random
phases and amplitudes. The interaction between the decaying modes will therefore vary with
each measurement, which will result in different decay curves. For this reason, a single decay
curve measured with interrupted noise is not particularly useful; hence, we need to take more
313
S o u n d
I n s u l a t i o n
Steady-state level before
excitation is stopped at t ⫽ 0 s
Sound pressure level, Lp(t) (dB)
5 dB
60 dB
T
10 dB
Background noise level
t=0s
Time, t (s)
Figure 3.56
Example decay curve measured using the interrupted noise method.
than one measurement at each position to calculate an average value. This average can be
calculated in two ways; we can either arithmetically average the reverberation times from each
curve, or ensemble average the decay curves. Both are acceptable, but the latter is preferable
because it results in smoother decay curves. A smooth curve is beneficial when determining
the reverberation time, TX , because in order to carry out linear regression over a range of X dB
it is necessary to clearly identify the starting point and the end point that define the evaluation
range. If there are large random fluctuations it can be difficult to identify the starting point that
is 5 dB below the initial level and the end point that is X + 5 dB below the initial level. Ensemble
averaging the decay curves is also preferable because the ensemble average of an infinite
number of curves gives the same decay curve as the integrated impulse response for that
point in the room (Schroeder, 1965). An infinite number of measurements is clearly out of the
question, but by averaging several decay curves we obtain a single, much smoother, decay
curve. This requires each individual decay curve, Lp (t), to be synchronized at the time when
the excitation is stopped. The ensemble average decay curve, Lp,av (t), can then be calculated
from the N individual decay curves using the following:
N
1 Lp,n (t)/10
(3.123)
10
Lp,av (t) = 10 lg
N
n=1
To determine the spatial average reverberation time from different source and receiver positions, it is acceptable to arithmetically average the reverberation times, or to ensemble average
the decay curves. The benefits of the latter are the same as have just been discussed.
3.8.2 Integrated impulse response method
When an acoustic system is excited by a Dirac delta function the resulting response is the
impulse response of that system, h(t) (Section 1.2.2). An acoustic system is defined such that
314
Chapter 3
it includes the space or structure under test as well as the chain of measurement equipment.
In practice, h(t) is the signal received by the analyser from the measurement transducer, e.g.
sound pressure from a microphone in a room or acceleration from an accelerometer on a wall.
For a linear acoustic system, the impulse response completely describes that system; hence,
the output signal from the system can be calculated from any known input signal that excites it.
The integrated impulse response method involves generating an impulse, and recording or
sampling the response of an acoustic system to this impulse. Due to its infinite height and
infinitely narrow width, it is not possible to create a Dirac delta function in practice. However, it
is possible to create an impulse of sufficiently short duration that can represent the Dirac delta
function. On structures, it is possible to generate an impulse with a hammer blow. In rooms an
impulse can be generated with a gunshot from a starting pistol, balloon bursts, handclaps, or
noise bursts via a loudspeaker. Although these impulse sources have been used in rooms for
many years, not all of them are omnidirectional and are able to generate a flat spectrum at a
sufficiently high level whilst avoiding very high crest factors (ratio of peak to rms). The latter can
cause problems due to the limitations of the detector in the analyser. All of these problems can
be avoided when the system to be measured is linear and time-invariant (LTI). The required
impulse response can then be measured with swept-sine signals (Müller and Massarani, 2001)
or a signal referred to as a Maximum Length Sequence (MLS) (Section 3.9).
The integrated impulse response method was introduced by Schroeder (1965) and involves
signal processing of an impulse response using reverse-time integration. Using this method
gives the same decay curve that would be determined by averaging an infinite number of
decay curves from measurements using the interrupted noise method. This results in a decay
curve that truly represents the characteristics of the acoustic system. In addition, a single
measurement gives a decay curve without random fluctuations, which increases the accuracy
of reverberation time calculations. We recall that for interrupted noise measurements, more
than one measurement needs to be taken at each receiver position to reduce the effect of
random variations.
From Schroeder (1965), the proof for the integrated impulse response method is based around
interrupted noise measurements made with stationary white noise as the input signal, x(t). We
assume that the input signal was switched on at some time in the past, represented by time
t = −∞, and that the acoustic system reaches a steady state before t = 0. The input signal is
interrupted at t = 0 in order to perform the decay measurement. The acoustic system under
test has an impulse response, h(t), hence the output signal, y(t), at the receiver position is
given by the convolution integral,
y(t) =
0
h(t − u)x(u)du
(3.124)
h(w)x(t − w)dw
(3.125)
−∞
Replacing t − u by a variable, w, gives,
y(t) =
∞
t
Due to the random nature of white noise, it is necessary to average all realizations of the input
signal (also referred to as an ensemble average) to find the expected value of the squared
output signal,
∞ ∞
E[ y 2 (t)] =
(3.126)
h(t1 )h(t2 )E[x(t − t1 )x(t − t2 )]dt1 dt2
t
t
315
S o u n d
I n s u l a t i o n
E[x(t − t1 )x(t − t2 )] is the expected value of the product of the signal with itself at another point in
time; this equals the auto-correlation function. Hence, substituting the auto-correlation function
for white noise from Eq. 3.5 into Eq. 3.126 gives
∞
E[ y 2 (t)] = N
h2 (w)dw
(3.127)
t
where N is the power spectral density.
Equation 3.127 shows that the squared impulse response acquired from a single measurement
can be integrated to give the expected value of the squared decay, E[y 2 (t)] from a large
number of interrupted noise measurements.
Practical implementation of Eq. 3.127 involves recording or sampling the impulse response,
and then playing it backwards into an analyser where the signal is squared and then integrated.
The direction of signal analysis is from the end of the decay to the start of the decay; hence,
this process is referred to as reverse-time integration or backwards integration. Effectively,
the integration starts at the end of the squared signal. From this starting point, each sample
of the decaying signal is replaced by the sum of the mean-square sound pressure of that
sample and the summed mean-square sound pressure from all previous samples back to the
starting point. This results in a new decay curve of level versus time, the logarithmic integrated
impulse response. For a decay curve normalized to the total energy of the impulse response,
the process of reverse-time integration gives the logarithmic integrated impulse response as
⎞
⎛ t
⎞
⎛ ∞
2
2
h
(w)d(−w)
h
(w)dw
⎟
⎜
⎟
⎜ t
⎟ = 10 lg ⎜ ∞ ∞
⎟
Lp (t) = 10 lg ⎜
(3.128)
⎠
⎝
⎠
⎝ ∞
2
2
h (w)dw
h (w)dw
0
0
In practice, the tail end of any decaying signal is buried beneath the background noise. Therefore as reverse-time integration cannot actually start at t = ∞, it is necessary to find a starting
point which gives a decay curve that is not significantly altered by the presence of background
noise (Lundeby et al., 1995; Nilsson, 1992; Vorländer and Bietz, 1994). The effect of different
starting points on the logarithmic integrated impulse response can be calculated using an idealized impulse response. An idealized squared impulse response with an exponential decay
defined by the reverberation time is given by
−6t ln 10
h2 (t) = exp
(3.129)
T
It is assumed that during the impulse measurement there is steady background noise with a
2
mean-square pressure, pnoise
. As an example, if we have a background noise level that is
2
40 dB below the level at t = 0, then pnoise
= 10−40/10 .
The measured signal consists of the idealized impulse response combined with the back2
ground noise. Hence in Eq. 3.128 we need to replace h2 (w) with h2 (w) + pnoise
, and replace
the integration limit, ∞, with tmax . For the idealized impulse response this gives the logarithmic
integrated impulse response as
⎛
⎞
⎧
⎫
2
6pnoise
ln 10
⎪
⎪
−6t/T
−6tmax /T
⎪
⎪
10
−
10
+
(t
−
t)
⎪
⎪
max
⎜
⎟
⎪
⎪
⎨
T
⎟ for 0 ≤ t ≤ tmax ⎬
10 lg ⎜
⎝
⎠
2
(3.130)
Lp (t) =
6p
t
ln
10
max
⎪
⎪
1 − 10−6tmax /T + noise
⎪
⎪
⎪
⎪
⎪
⎪
T
⎭
⎩
−∞ dB
for t ≥ tmax
316
Chapter 3
60
Actual decay (T ⫽ 0.6 s)
50
Background noise (20 dB)
Logarithmic integrated
impulse response
Level (dB)
40
30
20
10
tmax⫽0.3 s
tmax⫽0.4 s
tmax⫽0.6 s
tmax⫽0.9 s
0
0
0.1
0.2
0.3
0.4
0.5
0.6
Time, t (s)
0.7
0.8
0.9
1
Figure 3.57
Example decay curves measured using the integrated impulse response method.
Examples of the logarithmic integrated impulse response calculated using Eq. 3.130 are shown
in Fig. 3.57 for different values of tmax . The actual decay curve has a reverberation time of 0.6 s
and has been shifted so that it starts at a level of 60 dB. The background noise level is 20 dB.
The later part of the measured decay becomes shorter or longer depending upon the value
of tmax . In this example, the level of the actual decay is equal to the background noise level
at t = 0.4 s. When tmax = 0.4 s the logarithmic integrated impulse response between t = 0 and
0.4 s is only slightly affected by the presence of background noise. Hence, despite the fact that
we can only estimate the actual background noise, suitable values of tmax can readily be found
that allow calculation of the reverberation time.
The approach used to determine tmax depends on the availability of the background noise level
during the impulse response measurement (ISO 3382). For steady background noise this level
can usually be estimated from the horizontal tail of the impulse response where the decaying
signal is well-below the background noise. As indicated in the example, the optimum value
for tmax is at the time where the actual decay curve equals the background noise. In practice,
we do not know the actual decay curve as this is the very reason that we are carrying out
the measurement in the first place. In addition, the fluctuations that exist in real background
noise mean that we cannot always obtain a good estimate from a short sample at the tail end of
the impulse response. For these reasons, the procedures used to determine the optimum value
tend to be based on iterative processes that are automated in the measurement equipment
(Lundeby et al., 1995; ISO 3382).
As with the interrupted noise method, different source and receiver positions are needed to
determine the spatial average reverberation time.
317
S o u n d
I n s u l a t i o n
3.8.3 Influence of the signal processing on the decay curve
With the interrupted noise method, the signal that is processed by the analyser passes through
the required filters before going to the detector. In the detector the signal is squared, and then
integrated over a specific time interval using either linear or exponential averaging to determine an average value. Averaging is necessary to reduce the fluctuations in the squared signal,
however if the averaging times are too long, it is not possible to accurately calculate the gradient of the decay curve. In the following discussion we are not concerned with the squaring
of the signal, hence all references to the detector will relate to its function as an averaging
device.
With the interrupted noise method, both the filters and the averaging device can alter the shape
of the decay curve. With the integrated impulse response method, the signal is filtered and
squared before using reverse-time integration to calculate the decay curve; hence, with this
method it is only the filters that can alter the shape of the decay curve.
The effect of the detector and the filters on the decay curve is thoroughly described by Jacobsen
(1987) and the same approach is used here by looking at the effect of the detector and the
filters separately.
3.8.3.1 Effect of the detector
The effect of the averaging device on the decay curve is determined by convolving the impulse
response of the detector, d(t), with the squared signal from an interrupted noise measurement,
x(t). This equals the output from the analyser, y(t), when we assume that it is only the detector
that affects the measured decay. Convolution of these two continuous time signals, d(t) and
x(t), is denoted as d(t)∗ x(t) and is defined by the convolution integral,
y(t) =
∞
−∞
d(u)x(t − u)du =
∞
−∞
d(t − u)x(u)du
(3.131)
where u is a dummy time variable.
The squared noise signal is interrupted at t = 0 and results in an idealized exponential decay.
This signal is defined using Eq. 3.129 to give
⎫
for t < 0 ⎬
−6t ln 10
x(t) =
⎩ exp
for t ≥ 0 ⎭
T
⎧
⎨1
(3.132)
where T is the actual reverberation time.
The impulse response for a linear averaging device with an integration (averaging) time, Tint , is
318
⎫
⎧
⎬
⎨ 1
for 0 < t < Tint
d(t) = Tint
⎭
⎩
0
for all other t
(3.133)
Chapter 3
For linear averaging, convolution gives the following decay curve,
⎫
⎧
⎞
⎛
−6t ln 10
⎪
⎪
⎪
⎪
T
−
T
exp
⎪
⎪
⎪
⎪
⎜
⎟
t
⎪
⎪
T
⎪
⎪
⎜
⎟
⎪
10
lg
1
−
for
0
≤
t
<
T
+
int ⎪
⎬
⎨
⎝
⎠
Tint
6Tint ln 10
Lp (t) =
⎪
⎪
⎪
⎪
⎪
⎪
⎪
⎪
⎪
⎪
6T
ln
10
T
−6t
ln
10
⎪
⎪
int
⎪
⎪
exp
−1
for t > Tint
⎭
⎩ 10 lg exp
T
T
6Tint ln 10
(3.134)
For exponential averaging, the impulse response is a decaying exponential curve that weights
the average towards the latter part of the time sample. The detector therefore has its own
decay time, Tdetector , which is defined as
Tdetector = 6τ ln 10
(3.135)
where τ is the time constant. When exponential averaging is implemented using an RC circuit
(resistor capacitor), τ is equal to RC.
The impulse response for an exponential averaging device is
⎫
⎧
⎨ 1 exp − t for 0 < t < ∞ ⎬
d(t) = τ
τ
⎭
⎩
0
for −∞ < t < 0
(3.136)
for which convolution with the idealized exponential decay gives the following decay curve,
⎫
⎧
6t ln 10
T
6τ ln 10
t
−6t ln 10
⎪
⎪
⎪
⎪
−
exp
−
10 lg exp
⎪
⎪
⎪
⎪
T
T − 6τ ln 10 T − 6τ ln 10
T
τ
⎪
⎪
⎪
⎪
⎪
⎪
⎨
T ⎬
Lp (t) =
for τ =
⎪
6 ln 10 ⎪
⎪
⎪
⎪
⎪
⎪
⎪
⎪
⎪
⎪
⎪
−6t
ln
10
6t
ln
10
T
⎪
⎪
⎭
⎩
10 lg exp
1+
for τ =
T
T
6 ln 10
(3.137)
Figure 3.58 shows the effect of linear and exponential averaging on the idealized decay using
Eqs 3.134 and 3.137. All the decay curves have been shifted so that they start at 60 dB. To
assess the effect of the detector on an idealized exponential decay, the decaying sound pressure level is plotted against the normalized time, t/T , so that it is applicable to any reverberation
time, T . If the effect of the detector on the idealized decay is negligible, the decay will have the
same gradient as the actual decay over the time interval, 0 ≤ t/T ≤ 1. The figures show that
the detector causes curvature of the decay curve, starting at the beginning of the decay. For
some of the curves produced by exponential averaging this is easier to see if a straight edge is
placed against them. The extent of the curvature over the time interval, 0 ≤ t/T ≤ 1, depends
on the properties of the linear or exponential averaging device.
The relevant Standards for reverberation time measurement (ISO 354 and ISO 3382) place
requirements on the detector in terms of Tint or τ, but also require that the evaluation shall start
5 dB below the initial sound pressure level at t = 0.
319
(b) Exponential averaging device
(a) Linear averaging device
60
60
5 dB
50
50
40
40
The time intervals
shown above
correspond to
0 ≤ t/T ≤ Tint/T
30
Level (dB)
Level (dB)
5 dB
The time intervals
shown above
correspond to
0 ≤ t/T ≤ 4τ/T
for T/Tdetector⫽1.448
and T/Tdetector⫽2.895
30
Actual decay
20
Actual decay
20
Linear averaging
T/Tint
T/Tdetector
2
1
4
10
Exponential averaging
1.448
10
7
2.171
12
2.895
0
0
0
0.1
0.2
0.3 0.4 0.5 0.6 0.7
Time, t, normalized to the actual
reverberation time, T (⫺)
Figure 3.58
Effect of linear and exponential averaging devices on the decay curve.
0.8
0.9
1
0
0.1
0.2
0.3 0.4 0.5 0.6 0.7
Time, t, normalized to the actual
reverberation time, T (⫺)
0.8
0.9
1
Chapter 3
With linear averaging, the detector causes significant curvature in the initial part of the decay
curve, 0 ≤ t < Tint . However, when t ≥ Tint , the decay curves shown in Fig. 3.58a have the same
slope as the actual decay curve. The evaluation range therefore needs to exclude the time interval, 0 ≤ t < Tint , with the requirement that T /Tint ≥ 4 (Jacobsen, 1987). Typical requirements in
the relevant Standards are that Tint shall be less than T /12 or less than T /7. Combining these
requirements with an evaluation range that starts 5 dB below the level at t = 0 means that there
is no need to refer to a specific time at which the evaluation should begin. This can be seen in
Fig. 3.58a where the requirement for an initial 5 dB drop excludes the initial non-linear part of
the decay that is caused by the detector for both T /Tint = 7 and T /Tint = 12.
With exponential averaging, the detector can also cause curvature of the decay curve. Figure
3.58b shows that the decay curve for T /Tdetector = 1 is significantly different to the actual decay
curve. Jacobsen (1987) proposed that the evaluation range should exclude the time interval,
0 ≤ t < 4τ (equivalent to 0 ≤ t < 0.29Tdetector ), provided that T /Tdetector > 2. Typical requirements
in the relevant Standards are that the time constant, τ, of an exponential averaging device
shall be less than, but as close as possible to (a) T /20 (where τ < T /20 is equivalent to
T /Tdetector > 1.448) or (b) T /30 (where τ < T /30 is equivalent to T /Tdetector > 2.171). An alternative requirement (used in ISO 354:1985 that has now been superseded) is that the equivalent
averaging time of an exponential averaging device, 2τ, shall be less than, but as close as possible to T /20. The condition 2τ < T /20 is equivalent to T /Tdetector > 2.895. By using an evaluation
range that starts 5 dB below the level at t = 0, the time interval, 0 ≤ t ≤ 4τ is excluded when
T /Tdetector > 2.895.
The time intervals that need to be excluded for linear and exponential averaging can be compared with the arrival time for the sound wave that travels directly from the sound source to the
microphone, and for the first reflected wave that arrives at the microphone. We are interested
in the decay of the reverberant field rather than propagation of the direct field. For this reason
it is necessary to exclude the time interval in which the last sound that was produced by the
source (i.e. just before it was switched off at t = 0) travels directly from the sound source to
the microphone. In addition, we want to exclude the time interval before the arrival of the first
reflected wave. This is because the mean-square pressure does not approximate an exponential decay until the arrival of the first-order reflections (Vorländer, 1995). As the first reflected
wave always arrives later than the direct wave, it is the arrival of the first reflected wave that
determines the time interval that should be excluded. By assuming a diffuse field, the mean
free path can be used to estimate the average path length from the source to the microphone
with a single reflection from a room surface. This assumes that the distance from the source to
the reflection point on the surface, and the distance from the microphone to the reflection point
on the surface are both equal to half the mean free path. The time taken for sound to travel a
distance equal to the mean free path is approximately 7 ms in a 50 m3 room and approximately
11 ms in a 200 m3 room. For the majority of rooms, the initial 7 or 11 ms will be contained within
the time intervals that are already excluded by using a starting point that is 5 dB below the level
at t = 0 and by satisfying the requirements on the detector in the relevant Standards (ISO 354
and ISO 3382). Hence, there is no need to exclude an additional time interval based upon the
position of the source and the microphone in relation to the room surfaces.
3.8.3.2 Effect of the filters
A filter has its own impulse response; hence, a filter also has its own decay time. To accurately
measure a decaying sound or vibration signal that is sent through a band-pass filter, the decay
321
S o u n d
I n s u l a t i o n
time of the filter must be shorter than the actual reverberation time that is to be measured.
This is necessary to ensure that we measure the reverberation time of the acoustic system
and not the reverberation time of the filter itself (sometimes referred to as ringing of the filter).
The normal measurement method is referred to as forward-filter analysis. This name is used
to distinguish it from reverse-filter analysis where the signal is played backwards through the
filter to reduce its influence on the decay curve.
3.8.3.2.1
Forward-filter analysis
For one-third-octave and octave-band filters that satisfy the relevant Standard (IEC 61260) the
impulse response is not symmetrical along the time axis. This can be seen from the impulse
responses for 6th order IIR Butterworth filters in Fig. 3.59. The asymmetry means that the filter
responds relatively quickly to an input signal in comparison to the time it takes for the response
of the filter to decay. This is a desirable feature for sound and vibration level measurements,
but it is not always ideal for reverberation time measurements. When a sinusoidal signal is
sent through a filter (where the frequency of the sinusoid is within the bandwidth of the filter)
there will be a short time delay of ≈1/B s before the filter responds to this sinusoid and outputs
a sinusoid with the same amplitude. The time interval on the decay curve, 0 ≤ t < 1/B, can
therefore be ignored; note that when BT > 12, this time interval occurs within the initial 5 dB
drop. We now need to look at how the relatively slow decay time of the filter impulse response
can affect the measured decay curve.
Filtering is implemented by convolving the impulse response of a filter, f (t), with the impulse
response of an acoustic system, x(t). The resulting impulse response is then processed using
reverse-time integration as described in Section 3.8.2 on the integrated impulse response
method.
The impulse response of a filter will vary depending upon the type of filter (e.g. Butterworth)
and its implementation in analogue or digital form. With digital implementation of a filter, its
impulse response can simply be determined by convolution with a unit impulse.
To define an idealized impulse response of an acoustic system it is convenient to use an
exponentially decaying sinusoid because this will give a decay curve that is a straight line. The
frequency of this sinusoid, f , is assumed to be equal to the band centre frequency of the filter.
This impulse response can therefore be considered as representing a single decaying mode
that lies within the filter passband. In practice this could occur on structures or in spaces in the
low-frequency range. However, there is usually more than one decaying mode in a band, and,
as we have seen in Section 1.2.6.3.2, this can result in decay curves that are not straight lines.
Therefore, it is easier to draw general conclusions on the effect of the filter by using a single
decaying sinusoid; so the idealized impulse response of the acoustic system is defined as
⎧
⎫
⎨ cos(2π ft) exp −3t ln 10 for t > 0
⎬
x(t) =
(3.138)
T
⎩
⎭
0
for all other t
where T is the actual reverberation time.
The effect of the filter on the decay curve depends upon the filter bandwidth and the actual
reverberation time. For this reason, it is useful to calculate decay curves for different values of
BT, where the time scale on the x-axis is normalized to the actual reverberation time (Jacobsen,
1987). The filter bandwidth, B, can be estimated using the band centre frequency, f , where
B = 0.23f , for one-third-octave-bands and B = 0.707f for octave-bands.
322
Chapter 3
(Linear units)
(a) One-third-octave-band filter
0
0
1
2
3
4
5
6
5
6
Normalized time, Bt (⫺)
(Linear units)
(b) Octave-band filter
0
0
1
2
3
Normalized time, Bt (⫺)
4
Figure 3.59
Impulse response of one-third-octave-band and octave-band filters.
323
S o u n d
I n s u l a t i o n
Figure 3.60 shows forward-filter analysis for one-third-octave-band filters. The plot of the decay
curve associated with the filter itself is shown alongside the plot of decay curves for the idealized
impulse response with values of BT between 1 and 32. From the impulse response of the filter
(Fig. 3.59), one would not expect its decay curve to be a straight line; in fact, the decay
curve contains a number of prominent ripples over the 60 dB decay range. When the actual
reverberation time is very short these ripples also appear in the filtered decay curve of the
exponentially decaying sinusoid as can be seen in Fig. 3.60 when BT = 1. As BT increases
from 1 to 32, the decay curve tends towards a straight line. We have previously seen that the
detector affected only the initial part of the decay curve; here we see that filtering not only
introduces curvature in the initial part of the decay curve but that it can also distort the main
part of the decay curve. When BT > 16 the main part of the decay curve is straight with minor
curvature in the initial 5 dB of the decay (Jacobsen, 1987). This minor curvature is unimportant
with interrupted noise measurements because evaluation of the decay curve starts 5 dB below
the level at t = 0 in order to minimize the curvature introduced by the detector. To show the
effect of the filter, the full 60 dB decay range is shown. However, the decay curve is almost
always evaluated within the initial 35 dB drop to calculate T10 , T15 , T20 , or T30 .
For octave-bands the decay curve associated with the filter has different ripples. However,
the same requirement that BT > 16 to give straight decays is applicable to octave-band
measurements.
For electroacoustic purposes, the design of band pass filters is primarily based on their
attenuation at frequencies inside and outside the passband, rather than their effect on decay
measurements (IEC 61260). The requirement that BT > 16 is appropriate for most one-thirdoctave and octave-band filters designed according to this Standard. However, this is quite a
strict requirement as it is based on the idealized situation of a single decaying sinusoid. In addition, the impulse response depends upon the individual filter, so manufacturer’s information
for the specific filter should be used whenever it is available.
In room acoustics measurements, the integrated impulse response method is often used to
determine the Early Decay Time (EDT) from the initial 10 dB of the decay. This method avoids
any distortion caused by the detector, but significant curvature in the initial part of the decay
curve caused by filtering needs to be avoided by ensuring that BT > 16. This requirement
is usually satisfied in spaces used for music performance because octave-bands are used
(i.e. large bandwidths) and the spaces tend to have quite long reverberation times. For sound
insulation measurements in both the field and the laboratory the BT > 16 requirement is stricter
than is necessary for the accuracy that is required; hence BT > 8 can be adopted instead (ISO
3382).
Figure 3.61 shows the minimum reverberation times that can be measured under the requirement that BT > 16 and BT > 8. We can look at the BT > 8 requirement in the context of
one-third-octave-band measurements in furnished dwellings. It is reasonable to assume that
furnished rooms have reverberation times of approximately 0.5 s. For reverberation times
≥0.35 s, the requirement that BT > 8 is always satisfied between 100 and 3150 Hz and it
is possible to measure much lower reverberation times in the mid- and high-frequency ranges.
If the requirement is not satisfied, the errors incurred by using forward-filter analysis depend
upon the evaluation range and the shape of the actual decay curve; hence, they are not easily
quantifiable. Errors can usually be avoided by using reverse-filter analysis that will be discussed
in the next section or by using octave-band measurements.
324
60
60
5 dB
50
50
40
40
Forward-filter analysis of an
exponentially decaying sine wave
with a one-third-octave-band filter
Level (dB)
Level (dB)
BT = 1
30
20
BT = 2
BT = 4
30
BT = 8
BT = 16
BT = 32
20
Decay curve of a
one-third-octave-band filter
calculated from its
impulse response using
forward-filter analysis
10
0
10
BT = 1
BT = 32
0
0
1
2
3
4
Normalized time, Bt (⫺)
(a)
0
1
2
3
Time, t, normalized to the actual reverberation time, T (⫺)
4
5
(b)
Figure 3.60
Effect of the filter on reverberation time measurement using forward-filter analysis. (a) Decay curve derived from the impulse response of the filter. (b) Decay curves for an exponentially decaying
sinusoid for different values of BT.
S o u n d
I n s u l a t i o n
1.4
Forward-filter analysis.
Minimum reverberation times
corresponding to the
requirement:
1.3
1.2
1.1
Minimum reverberation time (s)
one-third-octave-bands
1.0
BT > 16
octave-bands
0.9
0.8
0.7
one-third-octave-bands
0.6
BT > 8
octave-bands
0.5
0.4
0.3
0.2
0.1
0.0
50
80
125
200
315
500
800
1250
Band centre frequency (Hz)
2000
3150
5000
Figure 3.61
Requirements for the minimum reverberation times that can be measured using forward-filter analysis.
For the measurement of structural reverberation times it is useful to work in terms of loss
factors, hence the requirement that BT > 16 means that the maximum measurable loss factor
is 0.03161 for one-third-octave-bands and 0.09716 for octave-bands.
3.8.3.2.2
Reverse-filter analysis
Any problems encountered in satisfying the requirement that BT > 16 or BT > 8 for forwardfilter analysis can usually be overcome by using reverse-filter analysis (Jacobsen and Rindel,
1987). This approach is often essential when measuring the structural reverberation times of
building elements. It may also be needed for reverberation times in highly damped rooms such
as recording studios (Rasmussen et al., 1991).
To understand reverse-filter analysis it is useful to look at the convolution process from a
qualitative point of view. From the convolution integral (Eq. 3.131) we see that the output
from the filter is either determined by integrating the product of the filter impulse response and
the reverse-time impulse response of the acoustic system or by integrating the product of the
reverse-time filter impulse response and the impulse response of the acoustic system. We
have already noted that the impulse response of octave-band or one-third-octave-band filters
is asymmetric with a faster response time than decay time. Hence, if the impulse response
of the filter (or the impulse response of the acoustic system) can be reversed in time before
the convolution process, we can take advantage of the fast response time of the filter to
326
Chapter 3
measure shorter reverberation times than with forward-filter analysis. Reverse-time analysis is
implemented on analysers by storing the measured impulse response of the acoustic system,
and then playing it backwards through the filters.
Figure 3.62 shows reverse-filter analysis for one-third-octave-band filters; this can be compared
to forward-filter analysis previously shown in Fig. 3.60. Forward-filter analysis delays the signal,
whereas reverse-filter analysis results in a negative time delay. However, this does not prevent
determination of the reverberation time, which is determined using relative time, rather than
absolute times. The decay curve of the filter is significantly shorter than with forward-filter
analysis and does not contain prominent ripples. For this reason, the main unwanted effect
of reverse-filter analysis is curvature in the initial part of the decay curve. When BT > 4 there
is some curvature in the initial 5 dB of the decay, but after this the decay curve is straight
(Jacobsen, 1987). Although this example uses one-third-octave-bands, the requirement that
BT > 4 equally applies to octave-band measurements.
Another benefit of reverse-filter analysis is that the distortion of the decay curve caused by an
exponential averaging device is reduced. The requirement on the detector that T /Tdetector > 2
can be changed to T /Tdetector > 0.25 when the evaluation range starts 5 dB below the level at
t = 0 (Jacobsen and Rindel, 1987).
Figure 3.63 shows the minimum reverberation times that can be measured under the requirement that BT > 4. This requirement corresponds to a maximum measurable loss factor of
0.12643 for one-third-octave-bands and 0.38864 for octave-bands.
3.8.4
Evaluation of the decay curve
Measured decay curves are rarely perfect straight lines; hence, they need to be evaluated in
such a way as to minimize errors in the calculated reverberation time. Evaluation is usually
automated and carried out by the analyser to give a straight line that best fits the data in the
chosen evaluation range. The simplest form of linear regression, and the most commonly
used, is the least-squares method which is described in general textbooks on statistics. An
alternative method to calculate the reverberation time is to use a weighting function. This can
reduce the effect of random fluctuations in the decay curve at both ends of the evaluation range.
Vigran and Sørdal (1976) show that symmetric weighting functions can give a low variance
for the spatial average reverberation time, with negligible change to the mean value. To use a
weighting function, W (z), the reverberation time, TX , is calculated from the decay curve using
(Vigran and Sørdal, 1976)
TX =
60
X
∞
−∞
W
Lp (t) − L−5 dB
L−(X +5)dB − L−5 dB
dt
(3.139)
where L−5 dB is the starting point of the evaluation range (5 dB below the initial sound pressure
level), and L−(X +5)dB is the end point.
An example of a symmetric weighting function is a triangular function described by
⎫
⎧
⎪
for 0 ≤ z ≤ 0.5 ⎪
⎬
⎨ 4z
W (z) = 4(1 − z) for 0.5 < z ≤ 1
⎪
⎪
⎩0
for all other z ⎭
(3.140)
327
60
60
50
50
40
40
Level (dB)
Level (dB)
5 dB
30
20
Decay curve of a
one-third-octaveband filter calculated
10
from its impulse
response using
reverse-filter analysis
0
⫺3
⫺2
⫺1
0
30
Reverse-filter analysis of an
exponentially decaying sine wave
with a one-third-octave-band filter
BT ⫽ 1
20
BT ⫽ 2
BT ⫽ 32
BT ⫽ 4
10
BT ⫽ 8
BT ⫽ 1
BT ⫽ 16
BT ⫽ 32
1
2
0
⫺1
Normalized time, Bt (⫺)
0
Time, t, normalized to the actual reverberation time, T (⫺)
(a)
(b)
1
Figure 3.62
Effect of the filter on reverberation time measurement using reverse-filter analysis. (a) Decay curve derived from the impulse response of the filter. (b) Decay curves for an exponentially decaying
sinusoid for different values of BT.
Chapter 3
0.40
Reverse-filter analysis.
Minimum reverberation times
corresponding to the
requirement BT > 4
0.35
Minimum reverberation time (s)
0.30
one-third-octave-bands
octave-bands
0.25
0.20
0.15
0.10
0.05
0.00
50
80
125
200
315
500
800 1250
Band centre frequency (Hz)
2000
3150
5000
Figure 3.63
Requirement for the minimum reverberation times that can be measured using reverse-filter analysis.
3.8.5 Statistical variation of reverberation times in rooms
For interrupted noise measurements, the reverberation time varies between different measurement positions as well as between repeated measurements at the same position. For the
integrated impulse response method, only a single measurement is needed at each position;
so it is only the spatial variation within the room that needs to be considered. In rooms with
diffuse fields, the standard deviation for these sources of variation can be calculated for the
situation where interrupted noise measurements are taken using a point source driven with a
Gaussian white noise signal (Davy et al., 1979). In a diffuse field the decay curve is a straight
line. However, the random nature of white noise means that the measured decay curve will
have random fluctuations about this straight line; the approach taken by Davy et al. (1979)
to determine the reverberation time is to fit a straight line to the decay curve using the leastsquares method. It is assumed here that the filters and the detectors have no effect on the
shape of the decay curve within the evaluation range.
For the spatial variation of the reverberation time, the standard deviation, σT ,s , is (Davy
et al., 1979)
σT ,s =
720T
Bs X 3
10
ln 10
2
F
ln 10
X
10
(3.141)
329
S o u n d
I n s u l a t i o n
where X is the evaluation range (dB) used to calculate the reverberation time, Bs is the statistical
bandwidth of the filter, and the function F( ) is
F(z) = 1 −
3
12
12
(1 + exp(−z)) − 2 exp(−z) + 3 (1 − exp(−z))
z
z
z
(3.142)
For the interrupted noise method using an exponential averaging device with its own decay
time, Tdetector , the standard deviation, σT ,r , for repeated measurements at the same microphone
position is (Davy et al., 1979).
720T
T
10 2
ln 10
(3.143)
σT ,r =
F
X
Bs X 3 ln 10
10 Tdetector
For the interrupted noise method using a linear averaging device there are a larger number
of variables that determine the standard deviation. For a device which averages N samples
during an integration time, Tint , with a time interval, a, between the start of successive averages
(a ≥ Tint ), the standard deviation, σT ,r , for repeated measurements at the same microphone
position is (Davy et al., 1979)
!
"
" 10 2 1
2
"
+
" ln 10
Bs Tint
N
"
σT ,r = "
(3.144)
# X3
X2
10aX
+
+
720aT
4T 2
T3
Octave-bands will have lower standard deviations than one-third-octave-bands, although the
latter are most commonly used in sound insulation measurements.
The filter bandwidth used in the equations is the statistical bandwidth of the specific type of filter.
For Butterworth filters this can be calculated from the effective bandwidth using Eq. 3.9 (Davy
and Dunn, 1987). However, the statistical bandwidth is not always known or available; hence,
an estimate can be made by using the band centre frequency, f , which gives a bandwidth of
0.23f , for one-third-octave-bands and 0.707f for octave-bands.
For the interrupted noise method with Nr measurements at each of Ns different microphone
positions, the total standard deviation in seconds can be estimated using:
!
"
" 1
σT2 ,r
2
#
σT =
(3.145)
σT ,s +
Ns
Nr
For the integrated impulse response method, the total standard deviation in seconds can be
estimated using:
σT2 ,s
σT =
(3.146)
Ns
The spatial variation in terms of σT ,s is shown in Fig. 3.64 for one-third-octave-bands with a
reverberation time of 1 s. This illustrates how the standard deviation decreases with increasing frequency and with increasing evaluation range. However, it should not be inferred that
measurements should always use the largest evaluation range (i.e. T60 ) in order to minimize
the standard deviation. This model only applies to the diffuse field situation where there is a
straight line decay. For non-diffuse fields where the decay curve is not a straight line, T10 , T15 ,
330
Chapter 3
0.5
Reverberation time, T ⫽ 1 s
Measurement parameter
T10
Standard deviation, sT, s (s)
0.4
T20
0.3
T30
T60
0.2
0.1
0.0
50
80
125
200
315
500
800
1250
2000
3150
5000
One-third-octave-band centre frequency (Hz)
Figure 3.64
Standard deviation for the spatial variation of the reverberation time with different evaluation ranges.
or T20 needs to be used to relate the reverberant sound pressure to the sound power radiated
into a room (Section 1.2.7.5.2); note that σT ,s calculated in this section is only applicable to
diffuse fields.
To assess the total standard deviation for interrupted noise measurements we will take an
example using exponential averaging where the time constant, τ, of an exponential averaging
device is required to be less than, but as close as possible to T /30; to satisfy this requirement
we will use τ = 0.99 T /30. For parallel filter measurement of all frequency bands (using either
linear or exponential averaging), the choice of averaging time should be based on an estimate
of the smallest reverberation time to be measured. To look at the effect of increasing the
number of repeated measurements at each microphone position we will consider a 20 dB
evaluation range. With parallel filter measurements used for sound insulation, T20 can be used
for measurements over the entire building acoustics frequency range (ISO 140 Parts 3 & 4).
From Fig. 3.65 it is clear that there is little to be gained by taking large numbers of repeat
measurements, hence Nr = 2 is commonly used for reverberation time measurements that are
intended for sound insulation calculations.
The standard deviation in seconds (Eqs 3.145 and 3.146) increases with increasing reverberation time. Hence, it may be inferred that more measurements are needed for longer
reverberation times. However, to calculate the airborne or impact sound insulation, we need
to know the total standard deviation in decibels that corresponds to the correction terms,
10 lg(T /T0 ) and 10 lg(A/A0 ); this can be estimated from the normalized standard deviation using
σT (dB) ≈ 4.34
σT
T
(3.147)
where T is the mean value of the reverberation time.
Examples of σT (dB) are shown in Figs 3.66 and 3.67 for the interrupted noise method and the
integrated impulse response method respectively. This indicates that the standard deviation in
331
S o u n d
I n s u l a t i o n
Total standard deviation, sT (s)
0.20
Reverberation time, T ⫽ 1s
Number of measurements at
Measurement parameter: T20
each position, Nr
Number of different positions, Ns ⫽ 6
1
Exponential averaging: τ ⫽ 0.99T/30
2
0.15
4
8
0.10
16
0.05
0.00
50
80
125
200
315
500
800
1250
2000
3150
5000
One-third-octave-band centre frequency (Hz)
Figure 3.65
Total standard deviation for the interrupted noise method (exponential averaging) with different numbers of repeat
measurements at each position.
1.0
Measurement parameter: T20
Standard deviation, sT (dB) (dB)
0.9
Number of different positions, Ns ⫽ 6
0.8
Number of measurements
at each position, Nr ⫽ 2
0.7
Exponential averaging: τ ⫽ 0.99 T/30
0.6
Reverberation time, T
0.5
0.5
0.4
1
0.3
1.5
0.2
2
0.1
0.0
50
80
125
200
315
500
800
1250 2000
One-third-octave-band centre frequency (Hz)
3150
5000
Figure 3.66
Standard deviation in decibels for the interrupted noise method using exponential averaging.
decibels increases with decreasing reverberation time and that the integrated impulse response
method has slightly lower standard deviations than the interrupted noise method.
The standard deviations only apply to diffuse fields; hence the values in the low-frequency range
will not be applicable to many typical rooms. Davy et al. (1979) note that Eqs 3.141, 3.143, and
3.144 tend to overestimate the standard deviation in the low-frequency range. Davy (1988) used
332
Chapter 3
1.0
Measurement parameter: T20
Standard deviation, sT (dB) (dB)
0.9
Number of different positions, Ns ⫽ 6
0.8
0.7
0.6
Reverberation time, T
0.5
0.5
0.4
1
0.3
1.5
0.2
2
0.1
0.0
50
80
125
200
315
500
800
1250 2000
One-third-octave-band centre frequency (Hz)
3150
5000
Figure 3.67
Standard deviation in decibels for the integrated impulse response method.
measurements in reverberant chambers (100–600 m3 ) to derive empirical correction factors to
account for this overestimation based on the statistical modal overlap. The correction factors
were derived using T30 where the decay curve could be considered as linear over the evaluation
range; hence they are suited to reverberant chambers rather than typical rooms.
The choice of a suitable probability distribution for the spatial variation of reverberation time
is less well-defined than for the spatial variation of mean-square sound pressure. Bodlund
(1976) has shown that either a gamma or a normal (Gaussian) probability distribution could
be assumed for diffuse fields in empty rooms with hard surfaces. However, the presence of
absorption (particularly when concentrated on one surface) makes it harder to identify a single,
simple probability distribution that is appropriate.
3.9 Maximum Length Sequence (MLS) measurements
When airborne sound insulation is measured using broad-band noise, the sound pressure level
to be measured in the receiving room is not always sufficiently high above the background
noise. In these situations it is necessary to increase the power output from the loudspeaker;
this can be done by limiting the broad-band noise to an individual frequency band. However, the
increase in level may still not be sufficient. This problem can be overcome by using a signal that
is commonly referred to as MLS (Schroeder, 1979). This allows the measurement of signals
(e.g. sound pressure, vibration) at low levels, as well as reverberation times in the presence
of high background noise. MLS measurements are well-suited to reverberation time measurements as they determine the impulse response of an acoustic system; this facilitates use of
the integrated impulse response method as well as reverse-filter analysis (Section 3.8.3.2.2).
The signal processing for MLS measurements is more complex than with broad-band noise
measurements. Modern analysers usually automate the MLS measurement process so that it is
not essential (although it is beneficial) to have an in-depth knowledge of the signal processing.
333
S o u n d
I n s u l a t i o n
To reap the benefits of using MLS with sound insulation measurements, it is necessary to
accept longer measurement times, restrict changes in the environmental conditions during the
measurement, and prohibit the use of moving microphones and loudspeakers. This section
starts with an overview of MLS in its application to sound insulation measurements and then
reviews the limitations of the method. These limitations are due to the requirement for the
system under test to be linear and time-invariant.
More details on the generation and processing of MLS signals can be found in Rife and
Vanderkooy (1989).
3.9.1 Overview
MLS is a deterministic signal; in other words, it is exactly repeatable. This is in contrast to white
noise, which is a random signal and is therefore described by its statistics. However, like white
noise, the shape of the MLS frequency spectrum is white. An MLS signal theoretically has a
crest factor of unity and this low value is beneficial because it gives high signal-to-noise ratios.
An MLS signal consists of a periodic binary sequence where each value in the sequence is
either +1 or −1 and is equally spaced in time. One MLS period contains 2n − 1 values where
n is an integer and describes the order of the sequence. For example, an MLS period of order
n = 15 comprises a sequence of 215 − 1 values. The values in the sequence are generated
from digital shift registers (Schroeder, 1979). The length of the MLS period, TMLS , in seconds,
is calculated by dividing the number of values in the sequence, by the sampling rate of the
clock, fclock , that is used to generate the sequence,
TMLS =
2n − 1
fclock
(3.148)
To send the MLS signal to a loudspeaker, it is converted to an analogue signal of bi-polar pulses.
Part of an MLS sequence and its analogue form is shown in Fig. 3.68. The MLS sequence of
+1 and −1 values can be viewed as a sequence of impulses with different polarities.
The MLS measurement process is used to determine the impulse response of a system, h(t),
where the system is defined as the combination of the measurement equipment and the acoustic system that is being measured. For a linear and time-invariant system with an impulse
response, h(t), the input signal x(t) is related to the output signal, y(t), by the convolution of
x(t) and h(t).
y(t) = x(t) ∗ h(t)
(3.149)
The impulse response is determined from the cross-correlation function, Rxy (τ) between the
input and the output signals. Cross-correlation describes the relationship between the input and
output signals in terms of the time displacement, τ, between these two signals after the output
signal has been modified by the measurement equipment and the room and/or structure under
test. It is particularly useful in extracting signals from noise where the input signal is related
to the output signal. This is the situation in which MLS can be used to overcome the problem
of high background noise. We recall that the MLS signal is deterministic; hence, we know
everything about the signal that leaves the analyser to allow us to correlate it with the signal
received back at the analyser from the microphone. The repeatable nature of the MLS signal
means that when averaging the response of a linear and time-invariant system to an MLS
signal it will be the same each time, whereas unwanted background noise will be uncorrelated
334
Chapter 3
(a)
⫹1
⫺1
Time
(b)
⫹1
⫺1
Time
Figure 3.68
Part of an MLS period: (a) comprising a sequence of impulses with different polarities and (b) the analogue version of the
MLS signal.
at different points in time. Cross-correlation can also be carried out with output signals of
random noise by using Fast Fourier Transform (FFT) techniques. However, the random nature
of the output signal means that more averages are needed than with MLS, and time-windowing
is necessary because unlike MLS the output signal is not periodic.
The cross-correlation function is determined from the auto-correlation function of the input,
Rxx (τ), convolved with the impulse response, and is given by,
Rxy (τ) = Rxx (τ) ∗ h(τ)
(3.150)
For the MLS input signal, the auto-correlation function, Rxx (τ), is (almost) the Dirac delta
function, δ(τ). This allows us to make use of the fact that convolution of an impulse with the
impulse response of the system will reproduce the impulse response of that system. Hence
the impulse response, h(τ), is equal to the cross-correlation function, Rxy (τ).
Rxy (τ) = δ(τ) ∗ h(τ) = h(τ)
(3.151)
For MLS signals, Rxy (τ) can be calculated using the Fast Hadamard Transform (FHT), an efficient cross-correlation algorithm. The measured sequence of impulses that are received from
the microphone occur at different times, and have different polarities due to the +1 and −1
values used to create the sequence. The Hadamard transformation takes all the impulse
responses and shifts them so that they are all aligned to occur at the same time with the
335
S o u n d
I n s u l a t i o n
Reverberation
time
FHT
Filter or FFT
MLS
(analogue form)
Level
Figure 3.69
Schematic diagram of MLS measurement.
Squared impulse response, h2 (t) (dB)
Peak
Noise
Majority of the signal energy
t⫽0
t1
t2
t3
Time, t (s)
Figure 3.70
Squared impulse response of an acoustic system.
same polarity. This results in the impulse response, h(t). The impulse response is then processed to give the spectrum or the decay curve. An outline of the MLS measurement process
is shown in Fig. 3.69.
For level measurements, the squared impulse response, h2 (t) shown in Fig. 3.70 is used to
calculate the energy that is due only to the signal whilst avoiding the parts that are affected
by background noise. The mean-square signal level, Esignal , is calculated by integrating h2 (t)
between 0 and t1 . Between t1 and t2 , the signal level is affected by background noise and
between t2 and t3 , there is only background noise. Note that Esignal does not account for the
signal energy embedded in the noise between t1 and t2 . However, this will not cause errors
as long as t1 is chosen such that the signal energy between t1 and t2 is negligible compared
to the energy between 0 and t1 . The requirement in the relevant Standard (ISO 18233) is
that t1 ≥ T /3, where T /3 corresponds to the 20 dB down point from the peak of the squared
impulse response. To quantify this error we use an idealized squared impulse response with
336
Chapter 3
an exponential decay that is calculated from the reverberation time of the system, T .
h2 (t) = X exp
−6t ln 10
T
(3.152)
where X is the peak value of the squared impulse response.
Note that in describing features of the MLS measurement process, we need to refer to the reverberation time of the system, T . For measurements where the loudspeaker is in the same room
as the microphone, use of the term reverberation time is unambiguous. When we measure the
signal in the receiving room with the loudspeaker in a different room, we can still measure a
reverberation time from the impulse response, but this particular reverberation time is not used
in sound insulation calculations.
The estimate of the total signal energy is obtained by integrating the squared impulse response
from 0 to t1 . The total signal energy is determined by integrating from 0 to T . Therefore the
total signal energy is underestimated by E in decibels, where,
−6t ln 10
E = 10 lg 1 − exp
T
(3.153)
For t1 = T /3, E is −0.04 dB. This error is sufficiently small that it is reasonable to assume
that the majority of the energy is contained in the initial part of the squared impulse response
between 0 and T /3 s.
The sound pressure level, Lp,MLS , at a single microphone position in either the source or the
receiving room is calculated from the squared impulse response using:
Lp,MLS
t1
⎞
2
W
h
(t)dt
0
⎟
⎜
0
⎟
= 10 lg ⎜
⎠
⎝
Cref
⎛
(3.154)
where W0 is a constant describing the signal power per unit bandwidth of the MLS signal, and
Cref is an arbitrary reference value.
The background noise level is estimated by integrating the energy in the squared impulse
response between t2 and t3 .
Lbackground,MLS
⎞
t3
1
2
h (t)dt ⎟
⎜t −t
2 t2
⎟
⎜ 3
= 10 lg ⎜
⎟
⎠
⎝
Cref
⎛
(3.155)
To gain an improved estimate of the sound pressure level, Lp∗ ,MLS , this background noise can
be ‘removed’ from the signal energy that is contained between 0 and t1 .
⎛
⎜
⎜
Lp∗ ,MLS = 10 lg ⎜
⎜
⎝
t1
0
[W0 h2 (t)] −
&
1
t3 − t2
Cref
t3
t2
'
⎞
h2 (t)dt dt ⎟
⎟
⎟
⎟
⎠
(3.156)
337
S o u n d
I n s u l a t i o n
With broad-band noise signals, a correction can be made for the background noise using
Eq. 3.75. This correction must not be used with MLS measurements as account has already
been taken of the background noise in the MLS measurement process.
By using several stationary microphone positions in the source and receiving rooms, the spatial
average sound pressure levels Lp1 and Lp2 can be calculated and used to calculate the level
difference needed to determine the airborne sound insulation.
The reverberation time is determined from the integrated impulse response in the same way
as with other impulse measurements, such as a pistol shot, by the application of reverse-time
integration (Section 3.8.2).
The length of the MLS period, TMLS , should be equal to, and preferably longer than the reverberation time of the system being measured, T . This ensures that the MLS period will excite every
resonance in the system (Bjor, 1995). It is usually necessary to average the measured impulse
responses from several MLS periods to provide the required signal-to-noise ratio. Hence, this
minimum period length also avoids problems with the tails of the measured impulse responses
overlapping into the next MLS period, known as time aliasing. It therefore appears that we need
to know the result of the measurement, i.e. the reverberation time, before we can set the MLS
period length on the analyser to actually measure the reverberation time. This is not a problem
in practice because the maximum reverberation time that is typically encountered in rooms, particularly in dwellings, is less than 8 s. Therefore the order of the MLS sequence and the sampling
rate can be set to give a valid period length for the majority of sound insulation measurements.
The MLS measurement process uses one MLS period to excite the system so that it is in a
steady state before using subsequent MLS periods to make the measurement. In contrast to
measurements with broad-band noise, the MLS measurement process determines the background noise during the measurement period, which is the period of time that we are actually
interested in. Both stationary and transient background noise energy is uniformly distributed in
time over the squared impulse response. This ‘smearing’ of transient background noise energy
over time means that MLS effectively shows a degree of immunity to unwanted transients (Rife
and Vanderkooy, 1989).
For level measurements it is useful to assess the improvement in the signal-to-noise ratio,
SNR, that is gained by using the MLS measurement process compared to stationary random
noise such as broad-band noise commonly used for sound insulation measurements (Vorländer
and Kob, 1997). To do this the MLS signal is considered in two different ways: firstly using
the MLS measurement process, and secondly by treating the MLS signal as a stationary noise
signal and using it for measurements in the same way as with broad-band noise. By taking
advantage of the fact that MLS is deterministic, several MLS periods can be synchronized and
averaged to improve the signal-to-noise ratio. Between these MLS periods there will be correlation between the impulse responses, but there will be no correlation between the background
noise during these periods. By averaging N MLS periods, the estimated improvement in the
signal-to-noise ratio in decibels by using MLS compared to broad-band noise is (Vorländer and
Kob, 1997; ISO 18233):
SNR = 10 lg
NTMLS
t1
(3.157)
Hence doubling the number of MLS periods in the averaging process will increase the signalto-noise ratio by 3 dB.
338
Chapter 3
As noted previously, the majority of the signal energy is contained in the initial part of the
squared impulse response. Hence for the squared impulse response derived from the MLS
measurement process it is more appropriate to describe the signal-to-noise ratio for the MLS
measurement process using the peak-to-mean-square noise ratio (Vorländer and Kob, 1997).
The peak of the impulse response is indicated in Fig. 3.70, along with the portion of the signal
used to estimate the mean-square noise.
The estimated improvement in the signal-to-noise ratio in decibels by using MLS compared
to broad-band noise, i.e. the ratio of the peak-to-mean-square noise ratio for MLS, to the
signal-to-noise ratio for broad-band noise, is (Vorländer and Kob, 1997):
6TMLS ln 10
SNR = 10 lg
T
TMLS
= 10 lg
T
+ 11.4 dB
(3.158)
Because TMLS must at least equal T , the minimum value of SNR is 11.4 dB. A typical value
of TMLS to cover the building acoustics frequency range is about 8 s, so with a reverberation
time of 1 s, SNR is 20.4 dB. Such large improvements in the signal-to-noise ratio mean
that the MLS measurement process allows measurement of the airborne sound insulation in
circumstances where high background noise would otherwise prevent the use of broad-band
noise. MLS measurements can therefore be attempted in situations where the signal-to-noise
ratio for broad-band noise measurements would be close to 0 dB.
3.9.2
Limitations
There are many advantages to using the MLS measurement process for sound insulation
measurements, but there are limitations that need consideration in any measurement. These
stem from the requirement that the system being measured must be linear and time-invariant.
The relevant elements in this system are the chain of measuring equipment, the environmental
conditions in which sound pressure level measurements are taken, and all the physical elements, such as the walls, floors, or ground, that affect sound propagation and transmission
from the loudspeaker to the microphone.
In airborne sound insulation measurements the main factor that affects linearity is distortion
from the loudspeaker and/or the amplifier when they are driven too strongly. Non-linearity
is identifiable by large spikes in the tail of the impulse response between time t2 and t3 .
This problem is straightforward to solve by reducing the level of the output signal, and, where
necessary, increasing the number of averages to achieve the required signal-to-noise ratio.
Hence, it is the time-invariant requirement, rather than linearity, that is the more important
issue for sound insulation measurements (Bjor, 1995; Svensson and Nielsen, 1999; Vorländer
and Kob, 1997). In contrast, measurements using broad-band noise are relatively immune
to time-variance and there are fewer problems with time-variance if the impulse response is
determined using a swept-sine signal instead of MLS.
MLS measurement relies on the Hadamard transformation to shift the measured impulse
responses so that they all occur at the same time with the same polarity. Therefore, any
variable that changes sound propagation or transmission between the loudspeaker and the
microphone is a potential source of error if it changes the impulse response of the system during the measurement. Common causes of such time-variance are air movement, variation in
temperature, variation in humidity, moving loudspeakers, and moving microphones. Two types
339
S o u n d
I n s u l a t i o n
of time-variance need consideration: variation within an MLS period (intra-periodic) and variation between MLS periods (inter-periodic). For sound insulation measurements we typically
use MLS periods that are a few or several seconds in length, hence both types are relevant.
Time-variance has the potential to change the frequency, amplitude, and the phase of a wave.
However, it only takes a small change to the speed of sound to introduce significant phase
shifts; hence, this tends to be the more important factor (Vorländer and Kob, 1997). Both air
movement and variations in temperature change the speed of sound. The resulting phase
shift reduces the correlation between the impulse response of the system and the MLS signal,
causing a decrease in the measured sound pressure level.
There tend to be fewer problems due to time-variance in the laboratory compared to the
field because there is usually more control over the environmental conditions. However, it
is good practice to monitor the environmental conditions in both the laboratory and the field
during measurements with long averaging times. It is also beneficial to avoid using longer
measurement times than are actually needed to achieve the required signal-to-noise ratio.
3.9.2.1 Temperature
An idealized model for the effect of temperature changes in a room during a measurement is
described by Vorländer and Kob (1997). This theoretical model for single frequencies is useful
in illustrating important features associated with a linear change in temperature. Three single
frequencies are considered here: 100, 500, and 1000 Hz for which the actual reverberation
time for each frequency is chosen to be 1 s. The inter-periodic temperature increase is chosen
to be 0.0033◦ C per MLS period. Hence, by assuming an MLS period length that is equal to
the reverberation time, and measuring over 600 periods, there will be a total increase of 2◦ C
during a 10 min measurement. The calculated decay curves are shown in Fig. 3.71 alongside
the actual decay of the system. For sound insulation measurements, the reverberation time
60
Actual decay
50
Sound pressure level, Lp(t) (dB)
100 Hz
40
500 Hz
30
1000 Hz
20
10
0
⫺10
⫺20
⫺30
⫺40
0
0.1
0.2
0.3
0.4
0.5
0.6
Time, t (s)
0.7
0.8
0.9
1
Figure 3.71
Predicted decay curves for MLS measurements at three single frequencies when there is a change in temperature during the
measurement. The actual reverberation time at each frequency without temperature drift should be 1 s.
340
Chapter 3
is often calculated from the first 35 dB decrease in the decay curve. During this time, the
change in temperature can cause the initial decay rate to be slower than the later decay rate.
A change in temperature tends to cause larger changes in the measured reverberation time at
higher, rather than lower frequencies. These features have been observed in practice where
measurements have been made in rooms undergoing changes in temperature (Bradley, 1996).
An important feature identified by the Vorländer and Kob model is the minima in the decay
curves. As with most single frequency models, it is reasonable to assume that the predicted
minima will be deeper than would occur with frequency band measurements that are used in
practice. However, the time at which the first minimum occurs in the decay curve can be used
to establish guidance for the maximum allowable change in temperature, θmax in ◦ C, during a
measurement (Vorländer and Kob, 1997). For both level and reverberation time measurements
this results in an equation of the form,
X
θmax ≤
(3.159)
fT
where T is the reverberation time.
Hence, the allowable change in temperature is smallest with long reverberation times and/or
at high frequencies. The guidance in the relevant Standard (ISO 18233) defines the maximum
allowable change in temperature over the measurement period as X = 1300 for level measurements, and X = 200 for reverberation time measurements using the decay range from 0
to 30 dB.
3.9.2.2 Air movement
Air movement occurs indoors due to HVAC systems, fans, or open windows. However, these
sources are usually controllable; hence the main issue is with outdoor measurements of sound
pressure levels in the presence of wind. In the measurement of façade sound insulation, the
sound pressure level is measured outside the building; hence we are interested in the effect of
wind on sound propagation from the loudspeaker to the microphone and to the test element.
The wind changes the speed of sound and introduces phase shifts, whether the wind occurs
in gusts or in more steady conditions.
Façade sound insulation measurements in a variety of wind conditions show that both the
external sound pressure level and the signal-to-noise ratio are reduced by the presence of
wind (Horvei et al., 1998). The effect is particularly significant in the high-frequency range.
It therefore tends to be important with façade elements, such as windows and ventilators for
which the high-frequency range is often important in determining the single-number quantity
for the façade sound insulation. Field tests from Horvei et al. (1998) have been used as an
empirical basis upon which to set the limits on wind speeds in the relevant Standard (ISO
18233). This defines suitable wind conditions for external measurements by monitoring the
wind speed close to the loudspeaker to ensure that average wind speeds are less than 4 m/s
and the wind speed during gusts is less than 10 m/s.
3.9.2.3
Moving microphones
Sound pressure level and reverberation time measurements using broad-band noise can be
made with moving microphones and, in the laboratory, moving loudspeakers. Whilst this is
suitable for broad-band noise, any moving equipment such as microphones, loudspeakers, or
diffusers, causes problems in satisfying the time-invariant requirement for MLS measurements.
341
S o u n d
I n s u l a t i o n
Change in sound pressure level (dB)
0
⫺5
⫺10
⫺15
⫺20
⫺25
50
80
125
200
315
500
800
1250
One-third-octave-band centre frequency (Hz)
2000
3150
5000
Figure 3.72
Measured change in the sound pressure level when using a rotating boom (sweep radius of 1 m and a 64 s sweep period)
instead of stationary microphone positions with MLS measurements. Measured in a 206 m3 reverberation room with an MLS
period of order n = 15, TMLS = 1.491 s, 1000 periods used for averaging. Measured data are reproduced with permission from
Weise and Schmitz (2000).
To determine spatial average sound pressure levels or reverberation times it is common to
use microphones on a rotating boom rather than a number of stationary microphone positions.
With MLS measurements, movement of the microphone introduces both intra-periodic and
inter-periodic time-variance as the impulse response changes in magnitude and phase within
an MLS period and between MLS periods. This time-variance reduces the correlation between
the input and the output signals. The result is a reduction in the measured signal level compared
to the level that would be measured by using stationary positions along the path of the rotating
microphone (Bietz et al., 1997; Weise and Schmitz, 2000). The reduction is dependent upon the
sweep radius of the rotating microphone as well as frequency. Examples of measured data are
shown in Fig. 3.72 for a 1 m sweep radius that is commonly used in airborne sound insulation
measurements where there is no synchronization between the sweep time and the MLS period
(Weise and Schmitz, 2000). This shows the general trend that the reduction is largest at high
frequencies. However, for airborne sound insulation we are interested in a sound pressure
level difference rather than absolute levels. It is possible to gain a reasonable estimate of the
level difference between the source and receiving rooms if exactly the same sweep radius and
sweep period are used in both rooms (Bietz et al., 1997). In practice, the significant reduction
in level at high frequencies along with the potential errors in setting up identical rotating booms
means that the use of moving microphones with MLS is not a viable option (Weise and Schmitz,
2000). Hence, only stationary microphones are considered for MLS measurements.
3.10 Sound intensity
Instantaneous sound intensity is a vector quantity, and is equal to the product of the sound
pressure and the particle velocity. We are usually interested in measuring the time-averaged
342
Chapter 3
intensity in stationary sound fields; this vector quantity describes the net flow of sound energy
passing through a unit area that lies normal to the measurement surface. This can be used to
calculate the sound power radiated by various surfaces in rooms.
If we were to measure the intensity of a propagating plane wave in an anechoic environment,
this could be done simply by using measurements of the time-averaged mean-square sound
pressure (Eq. 1.19). In practice we want to measure sound intensity in a variety of different
sound fields in and around buildings, often where the radiated sound is reflected back from
other surfaces and where there are other sound sources present. For example, measurement
of sound intensity radiated by a separating wall in a reverberant room where there is flanking
transmission from the surrounding walls and floors. In these situations, a single microphone
cannot be used to give the magnitude and direction of the intensity, it is necessary to determine
both the sound pressure and the particle velocity.
To illustrate the issues pertaining to sound intensity measurement in the presence of reflected
waves it is simplest to look at the superposition of two plane waves travelling in opposite directions (Fahy, 1989). In a one-dimensional interference field, the temporal and spatial variation
of the sound pressure is described by:
p(x, t) = [p̂+ exp(−ikx) + p̂− exp(ikx)] exp(iωt)
(3.160)
where p̂+ and p̂− are the amplitudes of the two waves.
Equation 3.160 can be re-written in terms of spatially varying amplitude and phase terms, p̂(x)
and φ(x) as:
p(x, t) = p̂(x) exp[i(ωt + φ(x))]
The particle velocity can now be determined from Euler’s equation,
∂p
1
u(x, t) = −
dt
ρ0
∂x
(3.161)
(3.162)
which gives:
u(x, t) =
i
dφ(x)
i ∂p
dp̂(x)
=
+i
p̂(x) exp[i(ωt + φ(x))]
ωρ0 ∂x
ωρ0 dx
dx
(3.163)
The instantaneous sound intensity (sometimes called energy flux density), I(x, t), in stationary sound fields is given by the product of sound pressure (Eq. 3.161) and particle velocity
(Eq. 3.163). It has two components, the active intensity, IA , and the reactive intensity, IR ,
I(x, t) = p(x, t)u(x, t) = IA + iIR
(3.164)
Active intensity describes the net flow of sound energy and is proportional to the phase gradient
where,
2
IA = −
[p̂(x)] dφ(x)
cos2 (ωt + φ(x))
ωρ0
dx
(3.165)
It is the active component that is of most importance, and for stationary sound fields we need the
time-averaged value. The time-averaged active intensity is proportional to the phase gradient,
and is given by:
2
IA t = −
[p̂(x)] dφ(x)
2ωρ0 dx
(3.166)
343
S o u n d
I n s u l a t i o n
Reactive intensity describes non-propagating energy that is moving back and forth, and is
proportional to the gradient of mean-square pressure,
2
IR = −
1 d[p̂(x)]
sin [2(ωt + φ(x))]
4ωρ0 dx
(3.167)
The instantaneous reactive intensity can take non-zero values, but the time-averaged reactive
intensity, IR t , is zero. Quantifying the reactive intensity is of limited practical use. However,
its existence helps to explain the difficulties that occur when trying to measure the active component. In a propagating plane wave there is only active intensity. In the one-dimensional
interference field described above, there would only be reactive intensity if there was no dissipation of sound energy (i.e. no sound absorption). This does not occur in reality as there
will always be some dissipation of energy, which will result in an active component. In threedimensional space it is clear that there will also be a reactive sound field in reverberant rooms,
but because surfaces always absorb some fraction of the incident sound intensity, there will
also be an active component. The sound field is highly reactive in the nearfield of a vibrating
plate or very close to a sound source, but it fades quite rapidly with distance which allows
active intensity to be measured quite close to a vibrating surface, often between 0.1 and 0.3 m
from the surface. For sound insulation, we need to measure the active intensity in sound fields
which have various degrees of reactivity. The word ‘active’ is commonly omitted and we simply
refer to intensity.
A thorough overview of sound intensity in theory and in practice can be found in the book by
Fahy (1989).
3.10.1
p–p sound intensity probe
To measure sound intensity there are two main types of probe: the p–p and the p–u probe. The
p–p probe comprises two pressure microphones and uses a finite-difference approximation to
estimate the particle velocity. The p–u probe uses direct measurement of the pressure and
the particle velocity. The most common type is the p–p probe; this typically has two microphones in a face-to-face configuration, separated by a solid spacer of length, d (see Fig. 3.73).
We will focus on the measurement of active intensity using a p–p probe. This probe comprises
two microphones, No.1 and No. 2, which measure the sound pressures, p1 and p2 respectively.
Positive intensity is defined for a plane wave propagating in the x-direction from microphone
No.1 to No. 2.
For a wave propagating in the x-direction, the particle velocity, ux , is related to the pressure
gradient, ∂p/∂x, by Euler’s equation (Eq. 3.162). The acoustic centres of the two microphones
are separated by a distance, d, hence the pressure gradient can be determined using a finitedifference approximation,
p2 − p1
∂p
=
∂x
d
(3.168)
The microphones are equally spaced about the point at which the intensity estimate is made,
so the sound pressure at the mid-point between the microphones is
p=
344
1
(p1 + p2 )
2
(3.169)
Chapter 3
Microphone No. 1
Direction of incident sound
corresponding to positive intensity
Spacer
Microphone No. 2
p1
p2
d
Figure 3.73
Sound intensity measurement: p–p probe.
Propagation direction
θ
p1
x
p2
d
Figure 3.74
Orientation of a p–p sound intensity probe in a propagating plane wave field.
Therefore the time-averaged sound intensity in the x-direction, Ix , can be determined using
Ix = pux t =
9
1
1
(p1 + p2 )
2
ρ0 d
(p2 − p1 ) dt
:
(3.170)
t
A p–p probe measures the net intensity component along the axis of the probe, not the complete vector. The importance of this fact becomes clearer if we consider a single plane wave
propagating at an angle, θ, to the x-axis (see Fig. 3.74). The probe axis is aligned with the
x-axis. For a wave with intensity, I, the probe measures the component, I cos θ. For θ < 90◦ ,
the intensity component along the probe axis is 10 lg (cos θ) dB below the level at θ = 0◦ ; at
60◦ the intensity component along the axis of the probe is 3 dB below the level corresponding
to 0◦ . At θ = 90◦ the wave propagates perpendicular to the probe axis, and there is no intensity
component along the probe axis.
To determine the three-dimensional intensity vector it is necessary to measure in three mutually
perpendicular directions to give the resultant intensity vector, I,
I = Ix i + I y j + I z k
(3.171)
345
S o u n d
I n s u l a t i o n
3.10.1.1 Sound power measurement
To determine the sound power of a radiating object or surface, it is necessary to define a
measurement surface, SM , that encloses a volume around the object or surface. The basis for
this lies in Gauss’s divergence theorem; this concerns the fact that a field enclosed within a
volume can only change by flow into or out of the volume. Application of this theorem means that
the sound intensity inside a specified volume is determined by the stationary signal emitted
by the sound source(s) enclosed within this volume, and by the sound intensity that flows
into the volume due to stationary signals from sound sources outside it (if any). Hence the net
sound power, Wnet , of the source(s) enclosed within the volume can be found by using Gauss’s
theorem to transform a volume integral into a surface integral over the measurement surface,
Wnet =
∇.I dV =
I.n dS
(3.172)
V
S
where n is the unit normal vector of the measurement surface that encloses the volume, with
the vector pointing out of the volume enclosed by this surface.
The axis of the p–p probe must therefore be perpendicular to the measurement surface so that
the intensity component being measured is normal to this surface. This component is denoted
as In , and is given by
In = I.n
(3.173)
The temporal and spatial average value over the measurement surface gives the normal sound
intensity level, LIn , as
|In |
(3.174)
LIn = 10 lg
I0
where I0 = 10−12 W/m2 . Note that the magnitude is used to give an unsigned value of In . This
means that LIn does not give information on the direction of the measured intensity; In would
be positive for a net intensity component flowing out of the measurement surface and negative
when flowing into the surface.
3.10.1.1.1 Measurement surfaces
Example measurement surfaces are shown in Fig. 3.75. For the wall there are narrow strips
around the edge of the measurement surface on the side walls, floor, and ceiling; it is important
that these strips radiate insignificant sound power. We are only trying to measure one source
(in this case, a wall) hence these strips are unwanted sound sources that are enclosed within
the volume by the measurement surface. This is not usually an issue in the laboratory where
flanking transmission is suppressed. In the field, measurements are almost always taken in
the presence of flanking. Therefore, the sound power radiated by the test element should be
more than 10 dB higher than the total sound power radiated by these narrow strips (ISO 15186
Part 2). If this is not satisfied then the sound radiated by these strips needs to be reduced by
covering them with an additional lining. However, it is difficult to get an accurate measurement
of the sound power radiated by these narrow strips, particularly when they are only ≈ 0.1 m
wide, so a single scan along each strip is usually considered sufficient to estimate their sound
power (ISO 15186 Part 2). Another problem can occur in the field when these strips absorb
significant sound power. In this situation, the measured sound power will underestimate the
actual value. When the strips are highly absorbent, they need to be covered with a lining that
makes the combination of the lining and the wall/floor construction reflective.
346
Chapter 3
(a)
(b)
Figure 3.75
Measurement surface, SM (shaded area) used to determine the net sound power radiated by (a) a wall in a room and (b) a
window flush with the façade. The dashed lines indicate how the measurement surface could be divided into sub-areas for
scanning measurements (e.g. four sub-areas for the wall and five sub-areas for the window).
Horizontal pattern
Vertical pattern
Figure 3.76
Scanning patterns for intensity measurements.
3.10.1.1.2 Discrete point and scanning measurements
Measurements for sound insulation purposes are taken with stationary random signals, so
intensity measurements can either be taken at discrete points on the measurement surface, or
by scanning the probe across the surface. Both methods are commonly used in the laboratory,
but scanning is usually more practical in the field.
For scanning measurements it is common to split the measurement surface into a number
of sub-areas. The area or sub-area is scanned at a constant speed between 0.1 to 0.3 m/s,
using both a horizontal and vertical scanning pattern (see Fig. 3.76); the arithmetic average
of the measured intensity from the two patterns is usually deemed to be acceptable when the
difference between the two patterns is less than 1.0 dB (Jonasson, 1991, 1993; ISO 15186
Parts 1 & 2). A suitable distance between the lines in the scanning pattern depends on the
347
S o u n d
I n s u l a t i o n
variation in the intensity over the measurement surface; this distance is usually the same as
the distance of the probe from the surface (ISO 15186 Parts 1 & 2).
Calculation of the net sound power for discrete point and scanning measurements is described
in the relevant Standards along with other checks on the sound intensity measurement.
3.10.1.2 Error analysis
To assess the errors involved in intensity measurement with a p-p probe it is simplest to look at
single frequencies under the assumption that the results are generally applicable to frequency
bands.
The finite-difference approximation introduces a systematic error in both the pressure at the
mid-point and in the pressure gradient. These errors combine to give a systematic error in
the intensity, and depend upon the sound field that is being measured. As before, we will
assume a plane wave propagating at an angle, θ, to the x-axis (Fig. 3.74). In a plane wave
field where the intensity is measured in the x-direction, the normalized error, eFD (Ix ), due to
the finite-difference approximation is (Fahy, 1989; Pavić, 1977):
eFD (Ix ) =
1
Ix − I
≈ − kd cos θ
I
6
2
(3.175)
where Ix is the value measured using the finite-difference approximation and I is the actual
value.
Normalized errors in linear values are converted to decibels using
10 lg (1 + e(Ix ))
(3.176)
Hence, as eFD (Ix ) is negative, the finite-difference approximation causes the measured intensity
level to be an underestimate of the actual value.
For spherical waves from a point source, Eq. 3.175 also applies where kr >> 1, i.e. in the
far-field (Fahy, 1989). Another source of particular interest is a vibrating plate such as a wall
or floor; however, for finite plates it is difficult to quantify the error, even more so with nonhomogeneous, isotropic plates. This does not cause problems in practice if the probe is at
least 0.1 m away from the vibrating surface (ISO 15186 Parts 1, 2, & 3).
There is also a measurement error due to phase-mismatch in the equipment. For the propagating plane wave shown in Fig. 3.74, the phase difference between the two points is kd cos θ.
However, there will also be an unwanted phase difference, ±φPM , between the two measurement channels. This occurs because for each channel there will inevitably be minor differences
between the microphones and other hardware or signal processing equipment. Therefore the
impulse response for each measurement channel will have a slightly different phase response.
The total phase difference measured in a propagating plane wave field will be kd cos θ ± φPM ,
hence the normalized error, ePM (Ix ), due to phase mismatch is
ePM (Ix ) =
±φPM
Ix − I
≈
I
kd cos θ
(3.177)
where Ix is the value measured with phase-mismatch but without any other errors.
From Eq. 3.177 it is seen that phase-mismatch errors are smallest when the axis of the probe
is aligned with the propagation direction.
348
Chapter 3
With increasing frequency and increasing microphone spacing, d, the magnitude of the normalized error for the finite-difference approximation increases, whereas the magnitude of the
normalized error for the phase-mismatch decreases. The combined normalized error due to
the finite-difference error and phase-mismatch is
e(Ix ) ≈ −
1
kd cos θ
6
2
±
φPM
kd cos θ
(3.178)
Equation 3.178 can be used to assess what length of microphone spacer is needed to cover the
different parts of the building acoustics frequency range. It can be seen that there will be some
cancellation of the individual errors when the phase-mismatch is positive. Hence the error is
calculated for both positive and negative phase-mismatch. To make an assessment of suitable
spacer lengths it is necessary to set a tolerable error; this can be taken as a normalized error
of ±5%, which corresponds to limits of ±0.2 dB (Fahy, 1989).
For the band centre frequencies we will now assume the following phase-mismatch values:
φPM = ±0.05◦ between 20 and 250 Hz, and φPM = ±f/5000◦ between 315 and 6300 Hz. These
would satisfy the relevant requirement for a Class 1 p–p probe according to IEC 1043. As the
error is also angle-dependent we will consider a plane wave propagating at angles of 0◦ and
60◦ to the x-axis. The combined normalized errors are shown in Fig. 3.77 for two common
spacer lengths, 50 and 12 mm. With a 50 mm spacer the combined error is within the ±0.2 dB
limits between 50 and 500 Hz; above 500 Hz it is the finite-difference error that causes the
large error. With a 12 mm spacer the combined error is just within the ±0.2 dB limits between
0.4
Finite difference
Phase-mismatch (⫹f)
Phase-mismatch (⫺f)
0.4
Combined error (⫹f)
Combined error (⫺f)
0.3
Normalized error (dB)
Normalized error (dB)
0.3
0.2
0.1
0.0
⫺0.1
⫺0.2
d ⫽ 50 mm
θ ⫽ 0°
⫺0.3
0.2
0.1
0.0
⫺0.1
⫺0.2
d ⫽ 50 mm
θ ⫽ 60°
⫺0.3
⫺0.4
⫺0.4
50
80
125 200
315
500 800 1250 2000 3150 5000
50
80
0.4
0.4
0.3
0.3
0.2
0.2
0.1
0.0
⫺0.1
⫺0.2
d ⫽ 12 mm
θ ⫽ 0°
⫺0.3
500 800 1250 2000 3150 5000
0.1
0.0
⫺0.1
⫺0.2
d ⫽ 12 mm
θ ⫽ 60°
⫺0.3
⫺0.4
125 200 315
One-third-octave-band centre frequency (Hz)
Normalized error (dB)
Normalized error (dB)
One-third-octave-band centre frequency (Hz)
⫺0.4
50
80
125 200 315
500 800 1250 2000 3150 5000
One-third-octave-band centre frequency (Hz)
50
80
125 200 315 500 800 1250 2000 3150 5000
One-third-octave-band centre frequency (Hz)
Figure 3.77
Normalized errors for a p–p probe in a propagating plane wave (± 0.2 dB corresponds to a normalized error of 5%).
349
S o u n d
I n s u l a t i o n
200 and 2000 Hz. However, another important error needs to be considered, the effects of
scattering and diffraction from the probe itself. Fortuitously for half-inch microphones with a
12 mm spacer in a face-to-face configuration, these effects compensate for the finite-difference
error; this means that the 12 mm spacer can be used up to 10 000 Hz (Jacobsen et al., 1998).
Hence it is possible to cover the entire building acoustics frequency range using a 50 and
a 12 mm spacer. From Fig. 3.77 it is clear that if the phase mismatch could be accurately
corrected, then measurements could be made over the entire building acoustics frequency
range with a single 12 mm spacer (Jacobsen, 1991; Ren and Jacobsen, 1992); some intensity
systems incorporate phase correction to achieve this.
The intensity probe is usually placed in an unknown sound field where the actual phase difference is φ. The measured phase difference will therefore be φ ± φPM . In order to assess the
validity of intensity measurements, it is useful to compare the measured phase difference to
the corresponding phase difference in an idealized sound field. For this purpose, we use the
phase difference in the direction of a propagating plane wave (θ = 0◦ ), for which the phase
difference is kd. Now we can look at the ratio of the phase difference in the direction of a propagating plane wave to the measured phase difference. For reasons that will shortly become
apparent, this ratio in decibels is referred to as the surface pressure-intensity indicator, FpI ,
(and sometimes as the pressure-intensity index or field indicator),
kd
FpI = 10 lg
(3.179)
φ ± φPM
As the phase difference is proportional to the intensity, FpI can be re-written in terms of the
intensity for a propagating plane wave relative to the measured normal intensity,
2
p t,s
ρc
FpI = 10 lg 0 0
In t
−6 2
= Lp − |LIn | + 10 lg (20 × 10 ) − 10 lg (ρ0 c0 )
10−12
(3.180)
where p2 t,s for the propagating plane wave is set equal to the measured temporal and spatial average mean-square sound pressure that is measured by the probe during the intensity
measurement in the actual sound field.
The last two terms in Eq. 3.180 cancel out, and FpI is given by,
FpI = Lp − |LIn |
(3.181)
Measurements will only give a non-zero value of FpI when the sound field is not a plane wave
propagating in the direction of the probe and/or there is significant phase-mismatch. It may
appear that this indicator is too far removed from any practical situation to be of any use. After
all, the actual sound field is rarely known. In and around buildings we assume that propagating
plane wave fields are not common because there are usually reflecting surfaces nearby; these
result in some kind of reverberant field. Even when we have a propagating plane wave, it is
likely to be propagating at an angle other than θ = 0◦ ; note that kd was only used in Eq. 3.179
to represent the maximum phase difference. The next step is to see how FpI can be used in
conjunction with another measurement to make it useful in practice.
The term ‘residual intensity’ is used to describe the intensity value that is reported by an
analyser purely due to phase-mismatch. When FpI is measured with the probe orientated in a
350
1.0
25
24
23
22
21
20
19
18
17
16
15
14
13
12
11
10
0.9
dpI0 (d ⫽ 50 mm)
0.8
0.7
0.6
0.5
dpI0 (d ⫽ 12 mm)
0.4
0.3
Phase-mismatch, fPM (°)
Pressure-residual intensity index, dpI0 (dB)
Chapter 3
0.2
Phasemismatch
50
80
125 200 315 500
800 1250 2000 3150
One-third-octave-band centre frequency (Hz)
0.1
0.0
5000
Figure 3.78
Example phase-mismatch errors for a p–p probe and the corresponding pressure-residual intensity indices for 50 and 12 mm
spacers.
field of uniform sound pressure such that φ = 0, the actual intensity is zero and the analyser
only measures the residual intensity. In this situation, FpI is referred to as the pressure-residual
intensity index, δpI0 , for a specific spacer distance.
kd
δpI0 = 10 lg
(3.182)
±φPM
The pressure-residual intensity index is measured in a small cavity specially designed for the
purpose. In this cavity the microphone diaphragms are exposed to the same sound pressure
but they are not separated by the solid spacer; the spacer distance is merely entered into the
analyser for calculation of δpI0 according to Eq. 3.182.
For a specific spacer distance the pressure-residual intensity index should ideally be as large
as possible. Figure 3.78 shows the pressure-residual intensity index for spacer lengths of 50
and 12 mm corresponding to phase-mismatch errors of φPM = ±0.05◦ between 50 and 250 Hz,
and φPM = ±f /5000◦ between 315 and 5000 Hz.
Both FpI and δpI0 use a propagating plane wave as a point-of-reference, hence the value,
δpI0 − FpI , can be used to give a measure of the error caused by phase-mismatch in the actual
sound field,
φ ± φPM
(3.183)
δpI0 − FpI = 10 lg
±φPM
Using δpI0 − FpI allows requirements to be set for negligible phase-mismatch error; the higher
the value of δpI0 − FpI , the lower the error. For precision or engineering grade accuracy, the
error in the measured intensity level is within ±0.5 dB when δpI0 − FpI > 10 dB. For survey
grade accuracy, it is within ±1 dB when δpI0 − FpI > 7 dB. These requirements are sometimes
described by introducing another term, the dynamic capability index, Ld (ISO 9614 Part 1),
Ld = δpI0 − K
(3.184)
351
S o u n d
I n s u l a t i o n
where K is the bias error factor. K is 7 dB for precision or engineering grade accuracy, or 10 dB
for survey grade accuracy. The requirement can then be stated as Ld > FpI .
Having established a requirement for negligible phase-mismatch error by using δpI0 − FpI , an
additional requirement can be set using FpI by itself. This can indicate when a sound field is
reactive and sufficiently different to a propagating plane wave as to render the measurement
invalid. For intensity measurement of building elements in the laboratory, the requirement in the
relevant Standards depends upon whether the surface that is being measured with the intensity
probe is reflective or absorbent. The requirement is FpI ≤ 10 dB for sound reflecting test elements and FpI ≤ 6 dB for sound absorbing test elements (ISO 15186 Parts 1 & 3). The requirement is more stringent for sound absorbing elements because the intensity probe measures net
flow of sound energy through unit area. If the surface of the test element that is being measured
has a much higher absorption area than the total absorption area of the room, measurement of
the net sound intensity will underestimate the sound power actually transmitted by the test element, and overestimate the airborne sound insulation (Roland et al., 1985; van Zyl et al., 1987).
When FpI is too high there are two common causes: the sound field is too reverberant and/or
the probe is too close to the element such that it is in the nearfield. These problems can often
be solved by adding sound absorbent material in the space where the measurements are being
carried out, and by moving the probe several centimetres further away from the test element.
This highlights an advantage of splitting the measurement surface into sub-areas when using
scanning sound intensity measurements. It is often easier to gain reliable measurements on
some sub-areas than others; using sub-areas sometimes avoids having to re-measure the
entire measurement surface.
Although it is not necessary to have a defined receiving room for sound intensity measurements, they are often carried out in the same reverberant receiving room that is used for
standard sound insulation tests. Therefore it is useful to have a rule-of-thumb that indicates
how much absorption is needed in this room to satisfy the requirement, FpI ≤ 10 dB. This can
be determined using the definition of FpI in Eq. 3.180. It is assumed that there is a diffuse field
in a room where waves are incident from all possible angles upon a perfectly reflecting and
rigid test element (Section 1.2.7.1.1). The sound pressure level very close to the surface of the
element will be 3 dB higher than the temporal and spatial average sound pressure level in the
central zone of the receiving room, p22 t,s , hence
p2 t,s = 2p22 t,s
(3.185)
The normal intensity, In t can be estimated from the power transmitted into the room by the test
element (Eq. 3.36) by assuming that all the transmitted sound is radiated by the test element,
therefore
In t ≈
p22 t,s A
4ρ0 c0 S
(3.186)
where S is the area of the test element and A is the absorption area in the room.
Assuming negligible phase-mismatch, an estimate of FpI can now be calculated from Eqs 3.180,
3.185 and 3.186 giving
FpI ≈ 9 + 10 lg
352
S
A
(3.187)
Chapter 3
This gives the rule-of-thumb that S/A < 1.25 is needed to satisfy the requirement FpI ≤ 10 dB.
For S = 10 m2 , the receiving room will therefore need A > 8 m2 , so additional absorbent material
(e.g. foam or mineral wool) usually has to be added in both the field and the laboratory. Rectangular blocks of absorbent material are a practical option in both the field and the laboratory
as they can be stacked on top of each other. These blocks can be placed a short distance
away from walls and on top of the floor as long as these surfaces are not being measured with
the intensity probe.
For ≈10 m2 test elements in rooms without additional absorbent material (field or laboratory),
FpI is typically between 5 and 15 dB over the building acoustics frequency range.
3.11 Properties of materials and building elements
The availability of measured dynamic or acoustic properties facilitates the choice of materials
at the design stage. They are particularly useful in situations where consideration is being given
to substituting one material or element in a construction with a different one.
3.11.1
Airflow resistance
Airflow resistance can be used to predict sound absorption by, and sound transmission through
porous materials. The fundamental parameters describing resistance to airflow are given in
Section 1.3.2.1.2.
Two methods to determine the airflow resistance are described in ISO 9053. The most common
method is the direct airflow method (Brown and Bolt, 1942). This uses an air supply (or vacuum)
to produce a pressure drop across a specimen of porous material for which there is no temporal
variation under steady-state conditions. Two measurements are required: the pressure drop
between the two faces of the test specimen, and the volumetric airflow rate. To ensure a
simple representative relationship between pressure and velocity for acoustic purposes it is
necessary to ensure there is laminar (rather than turbulent) airflow. An example test cell is
shown in Fig. 3.79; the dimensions and other details are described in ISO 9053.
An alternative method is the alternating airflow method (Venzke et al., 1972). This requires a
device to produce a slowly alternating airflow in order to measure the alternating component
of pressure in a volume enclosed by a sample of porous material. Equipment includes a 2 Hz
calibrating pistonphone (non-standard apparatus), a 2 Hz alternating airflow device, and a condenser microphone to measure the alternating pressure. The direct airflow method is simpler
to implement and tends to be more widely used.
Airflow resistance usually varies with linear airflow velocity (Brown and Bolt, 1942). Hence
the airflow resistance is typically measured with a low linear airflow velocity; 0.5 × 10−3 m/s is
recommended in ISO 9053. This corresponds to the particle velocity in a plane wave with a
sound pressure level of 80 dB.
Care needs to be taken when installing test specimens to ensure that there are no air gaps
between the specimen and the test cell. For fibrous materials there can be a wide variation in
the airflow resistance between samples taken from a single sheet, as well as between different
sheets. A sufficiently large number of test specimens need to be measured to determine the
average value. In addition, each test specimen should be weighed, and its thickness in the test
353
S o u n d
I n s u l a t i o n
Air supply
Rotameter/
flowmeter
Porous
material
Differential
pressure
measuring
device
Figure 3.79
Basic measurement set-up for the airflow resistance of porous materials using the direct airflow method.
cell must be measured. This is particularly important if the intention is to establish a relationship
between airflow resistance and bulk density; it is not always appropriate to assume a single
value for all samples using the bulk density quoted by the manufacturer.
3.11.2
Sound absorption
Measuring the sound absorption is relevant to both rooms and cavities. Only a brief
overview of the measurements is given here because there are detailed descriptions in the
Standards themselves, or in other books on room acoustics or sound absorption (e.g. Mechel,
1989/1995/1998).
3.11.2.1 Standing wave tube
The standing wave tube can be used to determine the normal incidence sound absorption coefficient (ISO 10534 Parts 1 & 2). It is particularly useful for comparing small samples of material
during product development or in quality control. It can also be used to assess absorbent material placed at the ends of long cavities in the low-frequency range where the normal incidence
absorption coefficient is relevant to the axial modes. However, in rooms there is a wide range
of angles of incidence and values at normal incidence are of limited use. If the absorber is
locally reacting then the statistical sound absorption coefficient can be calculated using the
measured impedance from the standing wave tube. However for many locally reacting materials, empirical laws can be used to calculate the normal incidence and statistical absorption
coefficients (Section 1.3.5.2.1).
3.11.2.2 Reverberation room
A reverberation room typically has non-parallel walls, hard reflective surfaces, and diffusers to
give a sound field that approximates a diffuse field in the steady-state, and during the decay
process. Reverberation room measurements can be used to determine the sound absorption
354
Chapter 3
coefficient or absorption area of an absorber (ISO 354). Due to the restrictions on what can
be tested in a standing wave tube, the absorption values determined using the reverberation
room method have greater practical application, particularly to the calculation of reverberation
times in situ.
The absorption area of the reverberation room excluding air absorption is calculated from
Eqs 1.83 and 1.97, which gives
A = AT − Aair =
24V ln 10
− 4 mV
c0 T
(3.188)
Reverberation time measurements are carried out with and without the absorber in the room.
Equation 3.188 is then used to calculate the absorption area in both cases to give the absorption
areas, Awith and Awithout . The absorption area of the test specimen, Aabsorber is then calculated
from
Aabsorber = Awith − Awithout
(3.189)
The sound absorption coefficient, αs , for an absorber placed over a wall or floor in the room is
αs =
Aabsorber
S
(3.190)
where S is the area of the wall or floor that is covered by the absorber.
Note that whilst αs can be compared with predicted values for the statistical sound absorption
coefficient, αst , it is possible for αs to take values that are larger than unity. This occurs due to
diffraction at, or absorption by, the edges of the absorber. Diffraction occurs because sound
waves that impinge upon the wall or floor effectively see two very different impedances adjacent
to each other; the hard reflective wall/floor surface and the absorber. Edge effects from the
finite size of the absorber need to be considered along with a decaying sound field that can be
significantly different to a diffuse field due to the presence of the absorber(s). To allow a fair
comparison of different absorbers the relevant Standard (ISO 354) therefore prescribes the
surface area of a plane absorber, its position in the room, and mounting conditions.
3.11.3
Dynamic stiffness
Vibration isolation between plates is usually achieved by incorporating resilient elements; for
example a resilient layer in a floating floor or wall ties in a cavity masonry wall. These resilient
elements can often be treated as simple springs that are described by their dynamic stiffness;
it is worth noting that there is not usually a simple relationship between dynamic stiffness
and static stiffness. In the laboratory, the dynamic stiffness can be quantified using mass–
spring systems by connecting the resilient element to one or more masses. The test set-up is
arranged so that each mass acts as a simple lump mass that does not support wave motion
in the frequency range that contains the resonance frequency of the mass–spring system.
When resilient elements are installed as part of a wall or floor it is assumed that they will
undergo small displacements. Hence they will act as linear springs obeying Hooke’s law with
a linear relationship between stiffness and displacement. Whilst this is a reasonable assumption for airborne sources, it is not always appropriate for structure-borne sound sources (e.g.
heavy impacts on a floor). To simplify the comparison of dynamic stiffness for different resilient
elements, and to give values that can be used in prediction models, it is useful to quantify the
355
S o u n d
I n s u l a t i o n
Linear spring
Hardening spring
Frequency (Hz)
Frequency (Hz)
Frequency (Hz)
Response
Softening spring
Figure 3.80
Mass–spring system: Frequency response with linear and non-linear (softening and hardening) springs. For the non-linear
springs, the arrows indicate the response curve that occurs when the excitation frequency is either increased or decreased.
dynamic stiffness when they act as linear springs. Unfortunately, many resilient elements act as
non-linear springs with increasing or decreasing stiffness with increasing input force; these are
described as hardening or softening springs respectively. In addition, it is not always possible to
identify the region that can be considered as linear. Figure 3.80 shows the frequency response
at and near resonance for linear and non-linear springs in a mass–spring system. For a linear
spring the peak of the curve occurs at the resonance frequency. Non-linear springs are more
complicated because at some frequencies the theoretical curve has more than one value. For
swept sine excitation, the response in these unstable regions jumps between different points
on the curve depending upon whether the excitation frequency is increasing or decreasing.
Difficulties in measuring the resonance frequency of non-linear springs can be overcome by
measuring the resonance frequency with very low input forces at a number of different force
levels within a defined range. When these input forces are sufficiently low, linear regression can
be used to determine a resonance frequency at zero input force (Pritz, 1987). This resonance
frequency is then used to calculate the dynamic stiffness corresponding to zero input force.
Whilst this is a step removed from reality, it is a pragmatic way of dealing with non-linear
springs. It is also quite adequate for calculating the resonance frequencies of walls or floors
that can be represented as mass–spring systems (see Sections 4.3.5.1 and 4.4.4).
In lightweight walls and floors, spring-like elements such as resilient bars, channels, or hangars
are often used to provide isolation. The dynamic properties of these elements can be affected by
the static load; for example a ceiling formed by resilient channels supporting one or two layers
of plasterboard of different weights. In addition, these elements cannot always be treated as
simple springs, and measurement of their dynamic properties usually requires more complex
measurement procedures (Brunskog and Hammer, 2002).
3.11.3.1 Resilient materials used under floating floors
A test set-up to determine the dynamic stiffness of resilient materials used under floating floors
is specified in ISO 9052 Part 1. This Standard is intended for materials subjected to static loads
in the range 0.4 to 4 kPa whilst the actual test uses a static load of 2 kPa; this range makes it
suitable for many floating floors such as those made from concrete screed or layers of sheet
material. It is mainly intended for the comparison of resilient materials used in continuous layers
(not small individual resilient mounts) under floating floors. However the measured dynamic
stiffness can be used to estimate the mass–spring resonance frequency of floating floors and
356
Chapter 3
Shaker
Force
Plaster of Paris
Velocity
Load plate
Resilient material
Figure 3.81
Measurement set-up for the dynamic stiffness of resilient materials.
wall linings as well as to predict the sound insulation at higher frequencies. Table A3 in the
Appendix has some examples of measured dynamic stiffness values.
No pre-loading of the resilient material is applied before the measurement, i.e. applying a static
load on top of the material for a certain time. Pre-loading can change the dynamic stiffness
such that it is no longer representative of the situation when it is actually installed under a
floating floor (Metzen, 1996).
3.11.3.1.1 Measurement
The simplest test set-up uses a mass–spring system as shown in Fig. 3.81. Measurement of
the mass–spring resonance frequency can be used to calculate the dynamic stiffness (Cremer
et al., 1973). A sample of resilient material is placed on a heavy rigid base. This is covered with a
thin sheet of waterproof film before applying a layer of plaster of Paris to account for any surface
irregularities and ensure excitation over the entire surface of the sample. For measurements
according to ISO 9052 Part 1, the dimensions of this sample are 200 × 200 mm. A steel load
plate with the same dimensions is then bedded down onto the plaster of Paris, and left to set.
The mass of the load plate is 8 kg, which corresponds to a 2 kPa static load. The load plate
must act as a lump mass, therefore the fundamental bending mode of the load plate must be
well-above the mass–spring resonance frequency; this mode will usually be above 1500 Hz for
a 200 × 200 × 26 mm steel load plate.
For porous samples (e.g. mineral wool, open-cell foams) it is important that the air is free
to move in and out of the sides of the sample during the test; therefore the waterproof film
and plaster of Paris must not cover the sides of the sample. The effect of air contained within
a porous material on the calculation of dynamic stiffness will be discussed shortly. For nonporous samples (e.g. closed-cell foams) there is no air movement in and out of the sample.
However, in the test set-up there can be air movement via the joint along the perimeter of
the non-porous sample and the heavy rigid base; hence it is necessary to seal this joint with
petroleum jelly (ISO 9052 Part 1).
A force transducer is used to measure the input force, and an accelerometer is positioned
adjacent to the force transducer to measure the vertical vibration of the load plate. The excitation force must be applied to the centre of the load plate so that there is only a vertical
component to the vibration. To prevent lateral and rotational forces being applied to the force
357
S o u n d
I n s u l a t i o n
iωm
F
v
m
v
k
R
k
iω
F
R
Figure 3.82
Mass–spring system representing the measurement set-up for the dynamic stiffness of resilient materials and its equivalent
electrical circuit.
transducer, a drive rod or stinger is used to connect the shaker to the force transducer (Mitchell
and Elliott, 1984). This must provide lateral flexibility whilst providing very high stiffness in the
axial direction; piano wire often forms a suitable stinger for dynamic stiffness measurements.
The force and acceleration signals are taken to a dual-channel FFT analyser to calculate
autospectra, and the Frequency Response Function that corresponds to the driving-point mobility, Ydp . Calibration of the measurement system is carried out using a freely suspended mass
(ISO 7626 Part 2). The actual system under test is likely to be non-linear with various different
damping mechanisms depending upon the material. However, because low force levels are
used it is still helpful to look at a linear mass–spring system with idealized damping. This can
be used to calculate the driving-point mobility from the equivalent circuit shown in Fig. 3.82,
which yields:
Ydp =
v
1
=
k
F
iωm + iω
+R
(3.191)
where m is the mass of the load plate and k is the spring stiffness (alternatively they may
represent the mass per unit area and stiffness per unit area). For a linear system with a
viscous damper, the damping constant, R, is related to the loss factor by:
√
R = η km
(3.192)
From Eq. 3.191, the magnitude of the driving-point mobility will be largest at the frequency
where,
iωm +
k
=0
iω
which is defined as the resonance frequency, fms , of the mass–spring system,
1
k
fms =
2π m
(3.193)
(3.194)
The resonance frequency is determined at a number of different input force levels within a
defined range; 0.1 to 0.4 N if s′ ≤ 50 MN/m3 , and 0.2 to 0.8 N if s′ > 50 MN/m3 (ISO 9052
Part 1). Linear regression can then be used to find the resonance frequency at zero input
force. For each input force, the resonance frequency can be found from the magnitude and/or
the phase of the driving-point mobility. The idealized linear mass–spring system described by
Eq. 3.191 gives an indication of the important features in the driving-point mobility (see
Fig. 3.83). At the resonance frequency, the magnitude reaches a peak and the phase is 0◦ .
For lightly damped springs, the resonance frequency can be identified from the peak in the
magnitude of the driving-point mobility or the peak in the vibration spectrum. High damping can
358
Chapter 3
Driving-point mobility-phase (°)
Driving-point mobility-magnitude (dB)
90
Increasing
η
60
30
0
⫺30
⫺60
⫺90
fms
fms
Frequency (Hz)
Frequency (Hz)
Figure 3.83
Driving-point mobility for a linear mass–spring system in the vicinity of the resonance frequency.
make it difficult to discern this peak and it is more accurate to identify the resonance frequency
from the phase of the driving-point mobility. When measuring phase it is worth being aware
that phase shifts can be introduced by charge amplifiers, and by changing the accelerometer
orientation by 180◦ . In addition the charge amplifier may not have a flat phase response with
frequency. To obtain the velocity signal from the accelerometer, a flat (but possibly phaseshifted) response can usually be obtained at low frequencies by taking the acceleration signal
from the charge amplifier, and using post-processing to integrate the signal to give the velocity.
The excitation signal can be a sinusoid (often using an automated sine sweep), white noise, or
some form of impulse (e.g. instrumented force hammer, MLS signal). ISO 9052 Part 1 states
that sinusoidal signals form the reference method in the case that there is dispute over different
resonance frequencies obtained with different signals. As it is quite common to measure slightly
different resonance frequencies with different signals, there is little motivation to use anything
other than a sinusoidal signal for the purpose of Standardization. With a force hammer it can
be difficult to generate a flat force spectrum with a low input force; the force levels typically
needed for a flat spectrum can drive the system into non-linear response.
3.11.3.1.2 Calculation of dynamic stiffness
The relevant parameter for resilient materials is the dynamic stiffness per unit area; this is
specific to the thickness of material under test. To make the link between dynamic stiffness
and the resonance frequency at zero input force requires consideration of any air movement
in and out of the sample during the test. This needs to be related to the in situ situation where
the resilient material is installed as a continuous layer under a floating floor, and the air only
moves within the resilient material.
The dynamic stiffness of porous resilient materials is effectively determined by two springs
connected in parallel; one spring representing the skeletal frame of the material, and another
spring representing the air contained within the material that surrounds the skeletal frame
359
S o u n d
I n s u l a t i o n
(Cremer et al., 1973). These springs add together in series to give the dynamic stiffness per
unit area of the installed resilient material, s′ , in N/m3 .
If the lateral airflow resistivity of the material is not too high, the small size of the test sample will
allow air to move freely in and out of its sides during the measurement. So in this test set-up,
the stiffness of the air within the sample does not come into play. Therefore the test only gives
an estimate of the dynamic stiffness for the skeletal frame; this is referred to as the apparent
dynamic stiffness per unit area, st′ , and is calculated using
2
st′ = 4π2 ρs fms
(3.195)
where ρs is the mass per unit area of the load plate and fms is the mass–spring resonance
frequency at zero input force.
For a porous resilient material in situ, the air contained within it acts as a spring with dynamic
stiffness per unit area, sa′ . For most porous resilient layers used under floating floors, the
mass–spring resonance frequency occurs in the low-frequency range or below it. At these low
frequencies it can be assumed that there is isothermal compression of the air. The dynamic
stiffness per unit area of the air is calculated from
sa′ =
K
φd
(3.196)
where K is the bulk compression modulus of air (for isothermal compression,
K = P0 = 1.013 × 105 Pa), φ is the porosity of the porous resilient material, and d is the thickness
of the sample under the load plate.
For porous resilient materials with lateral airflow resistivity in the range, 10 ≤ r < 100 kPa.s/m2
(ISO 9052 Part 1), the dynamic stiffness per unit area of the installed resilient material, s′ , can
be calculated from
s′ = st′ + sa′
(3.197)
Figure 3.84 shows sa′ for typical porosities and thicknesses of resilient material. Uncertainty in
estimates of the porosity usually results in a negligible change to sa′ , but it is important to note
that the stiffness of the contained air often forms a significant percentage of s′ . For layers of
mineral wool typically used under floating floors, sa′ is often between 25% and 250% of s′ .
For non-porous materials or porous materials with very high lateral airflow resistivity,
r ≥ 100 kPa.s/m2 , there is no need to include the sa′ term, hence s′ = st′ (ISO 9052 Part 1).
For porous resilient materials with lateral airflow resistivity, r < 10 kPa.s/m2 , then s′ = st′ only
if sa′ ≪ st′ , otherwise s′ cannot be quoted as satisfying the Standard (ISO 9052 Part 1). For
some reconstituted foams, r < 10 kPa.s/m2 , but sa′ ≪ st′ is not satisfied; yet the test can still
yield useful estimates by including the sa′ term (Hall et al., 1996).
Having quantified the dynamic stiffness with this test set-up it would be ideal if the internal loss
factor could be calculated from the same measurement; for example, using the 3 dB bandwidth
of the resonance peak. However, the measured loss factor is only likely to be relevant to this
particular test set-up. For porous materials, the measured loss factor has two components:
damping due to the skeletal frame, and damping due to airflow in and out of the sides of
the sample. The latter is often significant; hence the loss factor measured in the test set-up
may not be relevant to the in situ situation where the lateral dimensions of the layer are usually
much larger. For non-porous materials it is possible to rank order the loss factor measured with
360
Chapter 3
22
Dynamic stiffness of the air contained
within a porous material, s⬘a (MN/m3)
20
Porosity, φ
18
0.90
16
0.95
14
0.99
12
10
8
6
4
2
0
5
10
15
20
25
30
35
Thickness, d (mm)
40
45
50
Figure 3.84
Dynamic stiffness per unit area for the air contained within a porous material.
different materials, but it should be noted that the use of petroleum jelly around the perimeter
can also change the measured loss factor.
This test set-up only provides a single value of dynamic stiffness and provides no information
on frequency dependence. Other measurements show that the dynamic Young’s modulus
and loss factor of the skeletal frame (measured in a vacuum) are independent of frequency for
mineral wool (Pritz, 1986), but frequency-dependent for some foams (Pritz, 1994). For practical
calculations it is rarely necessary to measure frequency-dependent values when it is only the
low-frequency range that is of interest. It is usually sufficient just to be aware that the dynamic
stiffness of some materials will be frequency-dependent.
3.11.3.2 Wall ties
Wall ties in masonry cavity walls are rarely simple rectangular strips or cylinders; they usually
have a drip in the centre that is formed by twists, notches, or kinks. This allows water to drip
down rather than pass between the leaves of a cavity wall. Differences between these twists,
notches, or kinks, as well as the material properties, cross-sectional dimensions, and cavity
depth give rise to different values of dynamic stiffness. Table A4 in the Appendix has some
examples of measured dynamic stiffness values.
3.11.3.2.1 Measurement
The dynamic stiffness of wall ties can be determined by casting each end of a wall tie into a
concrete cube, and measuring the axial mass–spring–mass resonance frequency (Craik and
Wilson, 1995). The dynamic stiffness determined by the measurement is specific to the spacing, X mm between the cubes; this spacing corresponds to the depth of the wall cavity in which
the tie is to be used. Wall ties with a highly resilient material at the centre may have a dynamic
stiffness that is only determined by this material, so the stiffness will be independent of the
361
S o u n d
I n s u l a t i o n
Force
Wall
tie
Shaker
Velocity
Mass, m1
Mass, m2
Figure 3.85
Measurement set-up to determine the transfer mobility for the dynamic stiffness of wall ties.
cavity depth (Wilson, 1992); these ties are sometimes used in high performance cavity walls
in studios.
Standard concrete cube moulds can be used to make nominally identical 100 mm cubes using
a concrete mix that rigidly holds the wall tie (Hall and Hopkins, 2001). The test set-up is shown
in Fig. 3.85. To allow free vibration of the mass–spring–mass system, each cube is supported
along its centre line by a loop of cord.
The measurement equipment is the same as described for resilient materials used under floating floors. Excitation is applied to the centre point on the outer surface of the cube. The force
transducer measures the input force on cube mass, m1 , with an accelerometer to measure
vibration in the horizontal direction on cube mass, m2 . The force and acceleration signals are
taken to a dual-channel FFT analyser to calculate the Frequency Response Function that corresponds to the transfer mobility, Ytr . Wall ties tend to act as non-linear springs, some of which
are hardening and others are softening; an example of a wall tie acting as a hardening spring is
shown in Fig. 3.86 (Hopkins et al., 1999). Problems with measuring non-linear springs can be
overcome by applying low input forces and extrapolating to a resonance frequency at zero force
input. Therefore it is still useful to look at a linear mass–spring–mass system with idealized
damping using the equivalent circuit shown in Fig. 3.87; this gives the driving-point mobility as
Ydp =
k
iωm2 + iω
+R
v1
= $
% $
k
F
iωm1 iωm2 + iω + R + iωm2
k
iω
+R
%
k
iω
+R
%
and the transfer mobility as
Ytr =
v2
= $
F
iωm1 iωm2 +
k
iω
k
iω
+R
% $
+ R + iωm2
(3.198)
(3.199)
The magnitude of the transfer mobility is largest at the frequency where the denominator of
Eq. 3.199 is smallest. This gives the resonance frequency, fmsm , of the mass–spring–mass
system as
!
1 "
k
"
fmsm =
(3.200)
#
m1 m2
2π
m1 +m2
362
Chapter 3
⫺45
Transfer mobility-magnitude (dB)
⫺50
⫺55
ESD: 0.0012 N2 s/Hz
⫺60
⫺65
⫺70
⫺75
⫺80
⫺85
ESD: 0.0342 N2 s/Hz
⫺90
164.5
167
172
177
182
187
192
197
202
207
169.5 174.5 179.5 184.5 189.5 194.5 199.5 204.5 209.5
Frequency (Hz)
Figure 3.86
Measured transfer mobility of a mass–spring–mass system with a butterfly wall tie. Impulse excitation from an instrumented
force hammer is applied at various excitation levels (shown in terms of the Energy Spectral Density, ESD). Measured data
are reproduced with permission from Hopkins (1999).
iωm1
v1
v2
v1
k
F
m1
m2
k
iω
v2
iωm2
F
R
R
Figure 3.87
Mass–spring–mass system representing the measurement set-up for the dynamic stiffness of wall ties and its equivalent
electrical circuit.
An anti-resonance can be found in the driving-point mobility where the numerator of Eq. 3.198
is smallest. This occurs at the frequency,
1
k
(3.201)
2π m2
The magnitude and phase of the driving point or transfer mobility for an idealized linear mass–
spring–mass system (Eqs 3.198 and 3.199) are shown in Fig. 3.88. The peak in the magnitude
and/or the phase passing through 0◦ can be used to identify the axial mass–spring–mass resonance frequency. It is often necessary to use the phase when the resonance is highly damped.
The aim is to excite only the axial mass–spring–mass resonance. However, twisting or oblique
resonances may be excited with some wall ties, so it is necessary to identify the axial resonance with an additional measurement (Hopkins et al., 1999; Wilson, 1992). If there is doubt
as to which peak in the transfer mobility corresponds to the axial resonance, then the single
363
S o u n d
I n s u l a t i o n
180
Transfer mobility
150
Transfer mobility
Driving-point mobility
120
Driving-point mobility
Mobility-magnitude (dB)
90
Mobility-phase (°)
60
30
0
⫺30
⫺60
⫺90
⫺120
⫺150
⫺180
fmsm
Anti-resonance
Frequency (Hz)
Anti-resonance
fmsm
Frequency (Hz)
Figure 3.88
Transfer mobility and driving-point mobility for a linear mass–spring–mass system in the vicinity of the resonance frequency.
accelerometer on cube mass, m2 , can temporarily be replaced by two accelerometers. These
are positioned near the edges of the cube at equal distances from the centre point. The
accelerometers are aligned along the centre line of the cube in the horizontal plane, and then
in the vertical plane in a subsequent measurement. At the axial mass–spring–mass resonance
frequency the phase difference between these two accelerometers in both the horizontal and
vertical planes will be zero; at any other resonance frequency the phase difference will be
non-zero.
The resonance frequency is determined at a number of different input force levels in the range,
0.01 to 0.1 N (Hopkins et al., 1999). Linear regression is then used to find the resonance
frequency at zero input force.
3.11.3.2.2 Calculation of dynamic stiffness
From Eq. 3.200, the dynamic stiffness of a wall tie, sX mm , in N/m for a cavity depth of X mm is
2
mav
sX mm = 2π2 fmsm
(3.202)
where fmsm is the mass–spring–mass resonance frequency at zero input force, and the average
cube mass, mav , is calculated from the following:
mav =
(m1 + m2 + mtie ) − mtie
2
(3.203)
where the term in brackets corresponds to the complete test specimen.
3.11.3.3
Structural reverberation time
Structural reverberation times for bending wave vibration on walls and floors are usually short.
In addition, many building elements have low modal density and evaluation of the decay curves
364
Chapter 3
can be more awkward than with rooms. To relate the decay time to the reverberant vibration
level it is usually necessary to use T10 , T15 , or T20 and measure to three decimal places.
To measure a smooth decay curve that is unaffected by the reverberation time of the filters
it is generally best to use the integrated impulse response method with reverse-filter analysis
(Sections 3.8.2 and 3.8.3.2.2). The response of the structure is measured using an accelerometer. The impulse needed to excite bending waves can be provided by a single hammer hit
from a plastic-headed hammer (e.g. see Craik, 1981) or by using a signal such as MLS from
a shaker pushed up against the wall or floor (e.g. see Meier and Schmitz, 1999).
The measured reverberation time can be used to calculate the total loss factor. This is the sum
of the internal loss factor and all the coupling loss factors (radiation and structural losses). If
the mounting conditions for the structure are carefully arranged it is possible to use the total
loss factor to estimate the internal loss factor, or the coupling loss factor due to sound radiation
(see Sections 3.11.3.4 and 3.11.3.9).
As noted in Section 2.6.3, the internal loss factor of building materials can be non-linear if
measured with high-vibration amplitudes. The response is generally expected to be linear for
coupling loss factors between masonry/concrete elements that are connected at rigid junctions.
However, non-linearity could potentially occur with certain junctions; for example, where there
are resilient connections and the resilient material or resilient connecting element effectively
acts as a non-linear spring. There is some evidence from measurements on a masonry wall
in a transmission suite that hammer excitation gives higher total loss factors than MLS shaker
excitation (Meier and Schmitz, 1999). This may possibly be attributed to non-linearity, but it has
not yet been proved explicitly. Hammer excitation of walls and floors has been used for many
years and the results generally correspond to theoretical predictions. It is possible that state-ofthe-art measurement technology has exposed an issue that was previously hidden by difficulties
in correctly evaluating the decay curves. Whatever the reason for these discrepancies, it is
best to avoid driving a structure into non-linear response and to try and measure structural
decays using vibration levels that are representative of the situation in practice. For lightweight
and heavyweight walls excited with a hammer, it is possible to cause lateral deflections that
are unrepresentative of those induced by airborne excitation. In contrast, the impact sound
insulation of floors is usually measured with hammer excitation from the ISO tapping machine.
It is worth noting that non-linearity of the total loss factor has not yet been identified as an issue
with tapping machine measurements on bare concrete floors. To avoid any problem, a cautious
approach is to use MLS shaker excitation instead of a plastic-headed hammer. However, the
latter is quicker and more convenient than the former. If measurements are regularly needed
on a specific type of element in the laboratory (e.g. masonry walls) it is worth investing time
to compare both methods to see if they can be considered as equivalent within the range of
measurement uncertainty.
3.11.3.4 Internal loss factor
For structure-borne sound transmission on plates and beams, the internal loss factor quantifies
the damping due to the conversion of vibrational energy into heat energy. We will only look
at measuring the internal loss factor for bending waves as these are the most important. Two
methods are generally used to determine the internal loss factor. If the material comes in sheet
form, measuring the structural reverberation time is usually the most convenient approach.
Alternatively, short, narrow strips of material can be used to form beams; these are excited by
365
S o u n d
I n s u l a t i o n
a shaker and the loss factor can then be determined at the various resonance frequencies of
the beams.
For measurements on plates, the structural reverberation time can be used to determine the
total loss factor in frequency bands over the building acoustics frequency range. To minimize
errors in the evaluation of the decay curves, there should ideally be at least five bending
modes in the frequency bands of interest. If the coupling losses due to structural connections
and sound radiation are negligible, the total loss factor provides an estimate of the internal loss
factor. If the loss factor does not vary significantly with frequency, an average loss factor can
be determined from a chosen frequency range below the critical frequency. To avoid problems
with non-linearity, low levels with MLS shaker excitation can be applied.
For sheet materials such as plasterboard, chipboard, or OSB the plate can be suspended vertically by two loops of resilient cord. This resilient suspension minimizes the coupling losses,
but the radiation losses from the plate are inherent in the measurement. Fortunately, homogeneous plates without stiffeners do not radiate particularly efficiently below the critical frequency.
For most sheet materials the critical frequency is in the high-frequency range; this gives a relatively wide frequency range below the critical frequency that can be used to estimate the internal
loss factor. It may be necessary to avoid the frequency band immediately below the critical
frequency because of high radiation losses. These losses usually cause a very steep slope
in the early part of the decay curve. Note that trying to estimate the radiation efficiency and
subtract the calculated radiation losses from an unbaffled plate with free boundaries is highly
likely to introduce errors. Below the critical frequency the decay curves are usually reasonably
straight and T15 or T20 can be calculated accurately. Figure 3.89 shows example data from
sheet material below the critical frequency. The internal losses for these particular materials
can generally be considered as being independent of frequency. An arithmetic average value
can therefore be calculated from frequency bands below the critical frequency; this usually
109
Plywood (Birch) ηint = 0.0157
108
107
Plasterboard (Natural gypsum) ηint = 0.0141
Internal loss factor (dB)
106
Plasterboard (Combination of flue gas gypsum and
natural gypsum) ηint = 0.0125
105
104
103
102
101
100
99
98
97
96
50
80
125
200
315
500
800
1250 2000
One-third-octave-band centre frequency (Hz)
3150
5000
Figure 3.89
Internal loss factor of sheet materials determined from structural reverberation time measurements on plates. The legend
contains arithmetic average values from the frequency range below the critical frequency. Measured data from Hopkins are
reproduced with permission from ODPM and BRE.
366
Chapter 3
gives a reasonable estimate for use in prediction models over the building acoustics frequency
range.
For masonry blocks it is useful to measure the internal losses when the blocks are mortared
together as they are in full-size walls. Unfortunately their weight often makes it difficult to
structurally isolate the test element to minimize the coupling losses. One possibility is to create a
beam of several mortared blocks and determine the loss factor at the resonance frequencies of
the beam (Kuhl and Kaiser, 1952). Another possibility is to measure the structural reverberation
time on full-size walls that are structurally isolated, so that the total loss factor is primarily
determined by internal and radiation losses (Craik and Barry, 1992; Craik and Osipov, 1995).
Structural isolation can be achieved by building a free-standing, full-size wall on top of a resilient
material. However for structural stability it is necessary to have some support at points along
the sides of the wall; contact at these points also needs to be made via a resilient material to
minimize these unwanted structural coupling losses.
For laminated materials such as laminated glass, the internal loss factor usually varies with
temperature. For this reason it is convenient to take measurements on short strips of material
inside a small temperature-controlled chamber (Yoshimura and Kanazawa, 1984). The dimensions of the strip are chosen so that it acts as a beam with bending modes at or near the
frequencies of interest. For the comparison of laminated glass products, the strip dimensions
are standardized to 25 mm wide × 300 mm long (ISO/PAS 16940). This method can also be
used with other materials, not just laminates.
The strip is excited at the mid-point using random noise from a shaker and the input impedance
is measured at the excitation point using an impedance head (see Fig. 3.90). FFT analysis is
then used to determine a transfer function corresponding to the input impedance. The resulting
input impedance spectrum contains peaks at the various resonance frequencies of the strip.
The frequency resolution must be sufficiently fine to accurately determine the resonance frequencies and their 3 dB down points. The strip is usually lightly damped so it is important to
ensure that the transfer function accurately reproduces the resonance peaks (Randall, 1987).
The internal loss factor can be calculated at each resonance frequency, fi , using
f3 dB,i
ηint,i =
(3.204)
fi
where f3 dB,i is the half-power bandwidth.
In comparison with measuring structural reverberation times on a plate, this method using
beams is advantageous because it allows estimates of the internal loss factor near and above
the critical frequency. However, it only gives internal loss factors at specific frequencies, so it is
less efficient when there is significant variation with frequency. It is important to note that measured values for the internal loss factor depend on the mode shape; this means that there can
be significant differences between values determined on beams and plates (Dunn et al., 1983).
When the internal loss factor varies significantly with frequency, better estimates can be found
by using both methods to cover the building acoustics frequency range. Figure 3.91 shows an
example of the internal loss factor for laminated glass using both methods; note that the internal
loss factor increases significantly with increasing frequency and increasing temperature.
3.11.3.5
Quasi-longitudinal phase velocity
In Section 2.3 we saw that the phase velocity for bending, transverse shear, and torsional
waves can be related to the phase velocity for quasi-longitudinal waves, cL . This makes the
367
S o u n d
I n s u l a t i o n
(a) Test set-up
Material
Impedance head
Velocity
Force
Shaker
Input impedance (dB)
(b) Example spectrum for the input impedance
Frequency (Hz)
Figure 3.90
Measurement of the internal loss factor using short strips of material acting as beams.
109
108
20°C
107
Internal loss factor (dB)
106
15°C
105
104
11°C
103
102
101
100
99
98
97
96
50
80
125
200
315
500
800
1250
2000
3150
5000
One-third-octave-band centre frequency (Hz)
Figure 3.91
Measured internal loss factor of 6.4 mm laminated glass (3 mm–0.4 mm PVB–3 mm) at three different surface temperatures.
Open symbols connected by lines correspond to measurement of the structural reverberation time on a plate shown with 95%
confidence intervals. Shaded symbols correspond to individual resonance frequencies measured on short beams (average
from two beams). The critical frequency is estimated to lie in the 2000 Hz one-third-octave-band; measurements on the plate
are only shown below this frequency. Measured data from Hopkins are reproduced with permission from BRE.
368
Chapter 3
quasi-longitudinal phase velocity a very useful property. Table A2 in the Appendix has some
examples of measured values.
In principle it is possible to estimate the quasi-longitudinal phase velocity by identifying the
critical frequency from the coincidence dip in the measured sound reduction index of a plate.
In practice, many plates are orthotropic to some degree, the coincidence dip is not always
easy to identify and the lowest point in the coincidence dip is not always close to the critical
frequency; hence this method is prone to error. For small homogeneous specimens (e.g. solid
masonry blocks) the quasi-longitudinal phase velocity can be measured using ultrasound (BS
1881-203:1986) or determined from modal analysis (Maysenhölder and Horvatic, 1998). The
latter usefully allows estimates of both the Young’s modulus and Poisson’s ratio. In practice it
is usually the properties of an entire wall or floor that are of interest, rather than the properties
of individual blocks in a masonry wall. An impulse from a hammer can be used to excite quasilongitudinal waves, and measure the time-of-flight across a beam or a plate (Craik, 1982a). To
excite only quasi-longitudinal waves requires access to the end of a beam, or the edge of a
plate. This can usually be arranged in the laboratory, but not always in the field. The advantage
of this method is that measurements can be carried out over the length of a beam, or over the
length and width of a plate; hence for modular elements such as masonry walls, the effect of the
mortar joints can be included in the measurement. It can also be used for homogeneous beams
or plates that are either isotropic or orthotropic. However, it is not appropriate for sandwich
plates (e.g. plasterboard – rigid foam – OSB) or laminates (e.g. laminated glass) for which it is
less meaningful to consider quasi-longitudinal waves, as well as being more difficult to excite
them.
The measurement set-up is shown in Fig. 3.92. Two accelerometers are positioned along a
measurement line separated by a distance, d. These must be aligned in the same direction to
respond to in-plane vibration. An impulse from a hammer is used to excite quasi-longitudinal
waves in the plane of the test element. The signals from the accelerometer are sent to a dualchannel analyser (or oscilloscope) to measure signal voltage against time. Quasi-longitudinal
phase velocities are usually greater than 1400 m/s, hence the time resolution of the analyser
and the distance between accelerometers (ideally >1 m) must be chosen to give sufficient
accuracy (±5% is usually acceptable). The only part of the signal that is of interest is the initial
rising slope, after this the signal is affected by waves reflected from the boundaries, some of
which may have been converted into other wave types. The quasi-longitudinal wave velocity
is calculated from
cL =
d
t
(3.205)
where d is the distance between accelerometers and t is the time between nominally identical
points on the initial rising slope of the response.
Measured values from beams, cL,b , and plates, cL,p , can be used to calculate Young’s modulus
using Eqs 2.20 and 2.21 respectively.
Measurements can either be taken in situ or on isolated elements in the laboratory. For
heavyweight elements such as masonry/concrete walls and floors it is convenient to take measurements in situ. There is rarely any need for them to be decoupled at the boundaries, other
than to allow excitation along one edge for isotropic plates or two edges for orthotropic plates.
For lightweight elements it is necessary to make a distinction between cL measured on individual materials (e.g. a sheet of chipboard, a timber batten) and on combinations of materials
369
S o u n d
I n s u l a t i o n
d
Voltage
Time
∆t
Voltage
Time
Figure 3.92
Measurement of the quasi-longitudinal wave velocity.
(e.g. sheets of chipboard screwed to timber battens to form a floating floor). Most lightweight
walls or floors are formed from lightweight sheet material connected to beams. These beams
effectively act as stiffeners; therefore the quasi-longitudinal phase velocity is usually different
in the two orthogonal directions. To quantify cL for individual materials, it is necessary to measure them in isolation. Lightweight sheet materials (e.g. plasterboard) can be suspended from
resilient cord for the measurement. For heavier elements such as timber joists, it is usually
sufficient to rest them on a resilient material on the floor. Note that problems can occur if they
are placed directly onto a rigid surface (e.g. a concrete floor) because structure-borne waves
may also be excited in this other structure that will affect the measurement.
3.11.3.6
Bending phase velocity
For homogeneous thin plates, the simplest option to determine the bending phase velocity
is to calculate it from measurements of the quasi-longitudinal phase velocity. However, there
are times when it is not possible to gain access to the edge of the element, and/or the plate
acts as a thick plate at the frequencies of interest, and/or the element is non-homogeneous
(e.g. hollow blocks). In these situations the frequency-dependent bending phase velocity can
be determined by measuring the phase difference between two point on a plate (Nightingale
et al., 2004; Rindel, 1994; Roelens et al., 1997). These measurements are more complex and
tend to be required on reverberant plates which can make it difficult to accurately determine the
phase difference of the propagating wave. In addition these plates are not always large enough
370
Chapter 3
to take measurements that are out of the nearfield. For these reasons it is not always possible
to measure values across the entire building acoustics frequency range and the results can be
highly dependent upon the signal processing.
3.11.3.7 Bending stiffness
The bending stiffness can be determined with small strips of material acting as beams; this
uses the same test set-up as for internal loss factor measurements (Section 3.11.3.4). When
the stiffness varies with temperature, this measurement can be arranged inside a small
temperature-controlled chamber. The peaks in the input impedance spectrum correspond to the
resonance frequencies, fi . The bending stiffness for a plate of the material can be calculated
at each resonance frequency from these beam measurements (Yoshimura and Kanazawa,
1984; ISO/PAS 16940)
2
πL2 fi
Bp,i = ρs
(3.206)
2Ci2
where L is the length of the beam, and Ci is a constant depending on the resonance frequency.
(C1 = 1.87510, C2 = 4.69410, C3 = 7.85476 : NB Above the third resonance frequency it can
be difficult to achieve sufficient accuracy using this approach.)
3.11.3.8 Driving-point mobility
Driving-point mobility of structures is needed to calculate the power input for point force excitation of plates and beams. This is particularly relevant to the power input from the ISO tapping
machine into a floor. Whilst the driving-point mobility can be predicted for simple homogeneous
plates, it is not always possible for non-homogeneous plates, plates with low modal density,
or plates connected to a framework of beams. The driving-point mobility can also be used to
estimate the modal density, the loss factors of individual modes, and mode frequencies.
Driving-point mobility is given by the ratio of the velocity to the excitation force at the point of
excitation. This is measured using dual-channel FFT analysis to give the required Frequency
Response Function. For thin plates with a low mass per unit area, the force and velocity can be
measured at the same point using an impedance head. Many plates that form building elements
are too heavy and stiff to use impedance heads and require use of a separate accelerometer
and a force transducer. It is not particularly convenient to use excitation from a shaker because
of the difficulty in fixing the force transducer to the wall or floor (ISO 7626 Part 2). Hence the
force transducer is usually mounted in an instrumented force hammer (ISO 7626 Part 5).
The bandwidth of the impulse is affected by the tip attached to the force transducer as well as
the hammer mass. On most building elements, a plastic tip (rather than rubber or metal) can be
used to give a sufficiently flat force spectrum up to ≈1000 Hz. Calibration of the measurement
system is carried out using a freely suspended mass as described in ISO 7626 Part 2.
To measure driving-point mobility, the excitation point and the accelerometer are ideally positioned exactly opposite each other, on either side of the plate. However, for floors and walls
there may only be access to one side, so the accelerometer can only be positioned adjacent
to the excitation point. Even when there is access to both sides, the inconvenience and the
positional errors that occur in fixing an accelerometer on the opposite side of a large wall or
floor make it preferable to use the same side. By measuring the acceleration at a position
371
S o u n d
I n s u l a t i o n
adjacent to the excitation point we are effectively measuring the transfer mobility between two
points. However, when the bending wavelength is large compared to the distance, d, between
the accelerometer and the excitation point (i.e. kB d ≪ 1), a good estimate of the driving-point
mobility can be obtained. State-of-the-art accelerometers are sufficiently small that the distance between the excitation point and the centre of the accelerometer can usually be kept to
d ≤ 20 mm. For most plates that form walls and floors in buildings this will allow measurements
over the low- and mid-frequency range.
Mobility measurement using FFT analysis outputs both real and imaginary parts. It is usually
the real part that is of interest as this relates to power input and to the modal density. The
imaginary part and magnitude may be affected by the contact stiffness of the material at the
excitation point acting as a spring (Eq. 3.97).
To minimize phase errors from the charge amplifiers, no integration of the accelerometer
signal takes place before FFT processing. Therefore, the analyser gives accelerance which is
converted to mobility through a single integration (i.e. division by iω). Peaks in the spectrum
of the driving-point mobility correspond to resonant modes; therefore if the modal density
and damping are low, deep troughs (anti-resonances) will occur in the spectrum between the
resonant peaks. Coherence values at anti-resonances are often low due to the weak response
signal being contaminated with noise. This is usually apparent at low frequencies where the
mode spacing is large. However, it is the response at the resonant peaks that is of most interest.
If the FFT lines are used to calculate the average mobility in a frequency band containing
one or more modal peaks, then the low coherence at anti-resonances usually has negligible
effect on the frequency-average mobility. Frequency Response Function H1 (Channel A: Force;
Channel B: Acceleration) is usually used to give the optimum estimate because most noise is
in the acceleration signal, rather than the force signal.
Spatial averaging is carried out by arithmetically averaging the linear values of the driving-point
mobility. The standard deviation and 95% confidence interval can also be calculated; these
can help decide whether a simple infinite plate model for the driving-point mobility would be
adequate over part of the building acoustics frequency range. It is often worth storing all the
individual measurements and only carrying out the averaging after checking each individual
measurement for errors. For this reason it is useful to store a repeat measurement at each
position. Errors can occur due to the operator not producing a single ‘clean’ impulse from the
hammer hit. This can occur on crumbly, dusty, or powdery surfaces that can also cause weak
accelerometer fixing.
Frequency-averaging to give the mobility (real part) in one-third-octave or octave-bands is
needed for calculations of the power input or to determine the modal density. This is calculated
by averaging the discrete FFT lines in each frequency band using the following:
Re{Ydp }f =
1
fu − f l
fu
fl
Re{Ydp }df
(3.207)
where fl and fu are the lower and upper limits of the frequency band respectively.
Estimating the modal density by counting modal peaks in the response of an isolated plate
is generally prone to error; there may be modes with the same frequency, modes very close
together, or modes that are not discernable when the damping is high. For homogenous plates,
better estimates can often be calculated from the spatial average driving-point mobility (real
part). (This is not appropriate for plates with attached beams where the driving-point mobility is
372
Chapter 3
significantly different above the beams, compared to between the beams.) The modal density
for a frequency band with centre frequency, f , is (Clarkson and Pope, 1981; Cremer et al., 1973)
n(f ) = 4mRe{Ydp }f ,s
(3.208)
where m is the mass of the plate or beam (kg) and f ,s indicates a frequency and spatial
average value.
Loss factors of individual modes can either be calculated from the half-power bandwidth
(Section 2.8) or with a more accurate approach from modal analysis such as using vector
diagrams (e.g. see White, 1982). Mode frequencies can be identified from the peaks in the
driving-point mobility. However, unless the plate is uncoupled and isolated from other structures, it can be difficult to relate these to the local mode frequencies that assume idealized
boundary conditions.
In existing buildings, information is not always available on the thickness or density of walls and
floors. If a simple length or mass measurement is not possible then the driving-point mobility
can be measured in order to estimate either the thickness or the density (one of them must be
known). This also requires a reasonable estimate for the quasi-longitudinal phase velocity. If
the plate is homogenous then the driving-point mobility at high frequencies will tend towards
that for an infinite plate and Eq. 2.190 can be used to estimate either the thickness or the
density. One application of this approach is with concrete ground floors where there is no
access underneath. However, caution is needed when a concrete slab has been cast directly
onto the ground/hardcore because the slab will not necessarily act independently of what lies
beneath it.
3.11.3.9 Radiation efficiency
Accurate prediction of the radiation efficiency is not possible for all plates and for all boundary
conditions; unfortunately, reliance on measurement will not always provide the solution. This is
mainly because radiation efficiency is dependent upon the type of excitation (e.g. mechanical
point force excitation, bending waves impinging upon a boundary, plane sound waves). It is
also specific to the plate dimensions and the arrangement and type of stiffening elements;
hence measurements on one plate are not always transferable to other plate sizes.
Two relatively simple methods of measuring the radiation efficiency are considered here: one
uses measurement of plate vibration and radiated sound power, the other estimates the radiation efficiency from measurement of the structural reverberation time. There are also more
complex methods such as acoustical holography (Villot et al., 1992).
Radiation efficiency makes the link between the radiated sound power and the mean-square
bending wave velocity on the plate (Eq. 2.198). For this reason it is applicable to resonant transmission rather than non-resonant transmission; these terms are introduced in Section 4.3.1.
With airborne excitation, non-resonant transmission tends to be the dominant sound transmission mechanism at frequencies well-below the critical frequency; hence, measurement of plate
vibration and radiated sound power tends to result in an overestimate of the radiation efficiency
(Macadam, 1976). As it is usually the radiation efficiency below the critical frequency that is of
most interest, it is necessary to use mechanical excitation.
The radiation efficiency is usually measured with the intention of using it in an SEA model. For
this reason we ideally want to excite the plate modes using statistically independent excitation
373
S o u n d
I n s u l a t i o n
forces. This should provide equipartition of modal energy and no correlated modal response
(Section 4.2). However, statistically independent excitation forces are not easily provided. It
is much more convenient to use point force excitation from an electrodynamic shaker. The
disadvantage of mechanical excitation at a point is that it gives a correlated modal response
and depending on the excitation point, some modes will be excited and not others (Fahy,
1985). Point excitation also causes the vibration at the excitation point to be higher than the
spatial average plate vibration. To try and minimize these effects it is necessary to average the
plate response by taking measurements at a number of different excitation positions. Below
the critical frequency there will also be sound radiation from the nearfield at the excitation point.
This can be calculated as described in Section 2.9.7. Fortunately, the power radiated by the
nearfield is often negligible compared to power radiated by the bending modes. In principle,
the sound power radiated by the nearfield on homogeneous plates could be calculated and
extracted from the measured sound power; in practice this is prone to introducing errors due
to the uncertainty in the plate impedance and input force.
The measurement set-up requires baffles around the perimeter of the plate to be arranged
so that they are representative of in situ. If the plate is not baffled, there will be a degree of
cancellation between pressures on opposite sides of the plate. Below the critical frequency the
absence of baffles tends to reduce the radiation from plate modes acting as corner or edge
radiators.
For a test element installed in a transmission suite, the radiation efficiency is conveniently
determined by exciting the wall with a shaker in one room (source room) and measuring the
radiated sound pressure in the other room (receiving room). This avoids measuring the selfgenerated noise of the shaker. By measuring the plate vibration along with the reverberation
time in the receiving room, the radiated sound power and the mean-square plate velocity can be
used to calculate the radiation efficiency from Eq. 2.198. As with the measurement of Rresonant
(Section 3.5.1.2.2) a signal such as MLS can be used to overcome problems in measuring low
sound pressure levels in the receiving room. When the shaker is simply pushed up against
the wall, the power input may vary between excitation positions. Lpv,T can then be calculated
according to Eq. 3.49, and the radiation efficiency can be calculated from the following:
10 lg σ = Lpv,T
20 × 10−6
+ 20 lg
10−9
6V ln 10
− 10 lg T0 + 10 lg
Sρ02 c03
(3.209)
It is usually the radiation efficiency below and in the vicinity of the critical frequency that is
of most interest. The radiation efficiency above the critical frequency tends towards unity for
homogeneous isotropic plates. When measuring at frequencies above the critical frequency
and above the thin plate limit, the measured radiation efficiency sometimes becomes significantly lower than unity; this is due to thick plate effects in combination with point excitation. The
resulting values are not representative of the radiation efficiency under airborne excitation.
Example measurements on a solid homogeneous masonry wall with mechanical point force
excitation are shown in Fig. 3.93. These measurements with point excitation indicate that the
radiation efficiency can be slightly higher than unity at and above the critical frequency, even
though predictions for masonry/concrete plates with airborne excitation are usually better when
unity is assumed (Section 2.9.4.3).
For plates with high radiation efficiencies, very low internal losses, and a critical frequency
above the frequency range of interest, it is possible to estimate the radiation efficiency using
374
Chapter 3
Radiation efficiency, σ (⫺)
10
1
100 mm masonry (aircrete) wall (3.53 ⫻ 2.63 m)
f11⫽19 Hz, fc⫽341 Hz, fB(thin) ⫽ 1040 Hz
0.1
0.01
Measured
Method No. 1
Method No. 3
0.001
50
80
125
200
315
500
800
1250 2000
One-third-octave-band centre frequency (Hz)
3150
5000
Figure 3.93
Measured radiation efficiency for a 100 mm masonry wall in a transmission suite along with predicted values (method
Nos 1 & 3) assuming simply supported boundaries and baffles in the same plane as the wall. Measured data from Hopkins
are reproduced with permission from BRE Trust.
structural reverberation time measurements to give the total loss factor. An example application
is for plates with high critical frequencies that are bonded to a porous material for which the
radiation efficiency into the porous material is often much higher than into air (Craik et al.,
2000). The total loss factor is determined from measurements of the structural reverberation
time as described in Section 3.11.3.3. The plate needs to be isolated at its boundaries so
that coupling losses are negligible. If the internal loss factor is known (or can be measured),
this can be subtracted from the measured total loss factor to estimate the coupling loss factor
due to sound radiation from both sides of the plate (note that this is twice the value given by
Eq. 4.21). If the radiation efficiency is the same for both sides of the plate, then the result is
simply divided by two. If they are different (i.e. one side radiates into a porous material and the
other into air) then the estimate of the higher radiation efficiency into the porous material may
be improved by subtracting the estimated coupling loss factor for radiation into air on one side.
3.12 Flanking transmission
Measurements relating to flanking transmission for either airborne or impact sound insulation
can be considered in the following categories:
(1) Quantifying flanking sound transmission via building elements that form a single flanking path between rooms such as suspended ceilings and access floors
(laboratory measurements).
(2) Quantifying flanking sound transmission for individual flanking paths across
a junction where there is more than one possible flanking path (laboratory
measurements).
375
S o u n d
I n s u l a t i o n
(3) Quantifying flanking transmission between two or more connected elements to
provide input data for SEA or SEA-based prediction models to calculate the in situ
performance (mainly laboratory measurements, but sometimes made in the field).
(4) Quantifying the sound insulation provided by combinations of separating and
flanking elements by simulating the in situ construction in a flanking laboratory.
(5) Identifying and quantifying the effect of flanking transmission on the sound insulation that is achieved in situ. This provides information that can be used to try and
improve the sound insulation.
This section looks at measurements that fit into these five categories. The basic theory needed
to arrange such measurements in the laboratory, as well as to apply and interpret the results
is covered in Chapters 4 and 5.
3.12.1
Flanking laboratories
We recall that laboratory measurement of direct sound transmission across a building element
requires a transmission suite with suppressed flanking transmission. In a similar way, laboratory
measurement of flanking transmission requires a facility and/or measurement technique in
which all flanking paths can be suppressed other than the flanking path(s) of interest. This
increases the complexity of the test set-up. It also means that the type of laboratory facility
and/or measurement technique may differ for different types of construction.
3.12.1.1
Suspended ceilings and access floors
Laboratory facilities for suspended ceilings and access floors are described in the relevant
Standards (ISO 140 Parts 9 & 12). The basic principle is that all sound is transmitted between
the two rooms via the ceiling/floor and the plenum (see Fig. 3.94). The dividing wall must have
sufficiently high sound insulation that there is no significant transmission across it; the dividing
wall is usually isolated from the suspended ceiling or access floor by a resilient material to
minimize vibration transmission between them. One side wall and both end walls of the plenum
are lined with highly absorbent material.
Plenum
Suspended ceiling
Source
room
Dividing
wall
Receiving
room
Figure 3.94
Outline sketch of a transmission suite used to measure airborne flanking transmission via a suspended ceiling (or access floor
if viewed upside down).
376
Chapter 3
Suspended ceilings and access floors are usually considered as introducing a single flanking path between two rooms. However, in terms of an SEA model there is more than one
sound transmission path by which airborne or impact sound is transmitted between these two
rooms. These paths can often be described by a plate–cavity–plate system (see Sections 4.3.5
and 5.3.1.1.1) if the sound field in the plenum can be considered as reverberant (i.e. not too
highly damped). The transmission paths involve the sound field in the cavity and any structural
coupling between the plates where they meet along the line of the dividing wall. The sound
insulation measurement is specific to the test set-up with the plenum/floor/ceiling dimensions
and plenum absorption that are used in the laboratory. The only normalization that can be
carried out is to the absorption area in the receiving room.
The normalized flanking level difference, Dn,f , for suspended ceilings or access floors is (ISO
140 Parts 9 & 12):
A
Dn,f = D − 10 lg
(3.210)
A0
where the reference absorption area, A0 , is 10 m2 . Note that for suspended ceilings this is
normally quoted as Dn,c but it is calculated in the same way as Eq. 3.210.
The normalized flanking impact sound pressure level, Ln,f , for access floors (ISO 140 Part 12)
is given by:
A
(3.211)
Ln,f = Lp + 10 lg
A0
3.12.1.2 Other flanking constructions and test junctions
Requirements for the laboratory facilities and measurements on other flanking constructions
and test junctions are described in the relevant Standards (ISO 10848). The type of facility that is
required depends on whether sound pressure level measurements are being used to determine
Dn,f and Ln,f (as with suspended ceilings and access floors), or vibration measurements are
being used to determine parameters such as the coupling loss factor or vibration reduction
index for subsequent use in SEA or SEA-based models.
Sound pressure level measurements can be useful in quantifying flanking transmission paths
where the test junction is formed from plates that are not well-suited to inclusion in SEA or
SEA-based models. Examples include plates with a significant decrease in vibration level with
distance, plates where the vibration field on each element is not a reverberant bending wave
field, and plates where vibration transmission cannot be directly related to measurement of the
lateral surface velocity on each element. Measurements in a laboratory facility are relatively
simple to arrange when the elements that form the test junction lie in a single plane as previously seen in Fig. 3.94 for suspended ceilings or access floors. This set-up is appropriate
when the separating wall or floor has negligible influence on transmission between the flanking
elements that lie in the same plane; this could apply to some façade elements. However, with
many wall and floor junctions the separating and flanking elements are perpendicular to each
other and they all play a role in vibration transmission across the junction; hence shielding of
one or more elements is required as indicated in Fig. 3.95. To measure transmission via any
specific flanking path the connection at the junction must remain unchanged so that its role
in the vibration transmission process is unaltered. It is therefore necessary to shield one or
more elements to prevent them from being directly excited by airborne sound in the source
room and to significantly reduce the sound radiated into the receiving room. Airborne sound
377
S o u n d
I n s u l a t i o n
Source
room
Receiving
room
Element, i
Element, j
Flanking path
chosen for
measurement
Shielded element of
the T-junction
(shielding is achieved
by using efficient wall
linings on both sides)
Figure 3.95
Illustration of shielding used to measure airborne sound transmission via one specific flanking path. In this example the test
junction installed in the laboratory is a T-junction. The walls of the junction could be lightweight or heavyweight construction
and different to each other.
insulation measurements can then be used to determine the normalized flanking level difference (Eq. 3.210) for each specific flanking path. With structure-borne excitation from the
ISO tapping machine it is only necessary to shield elements in the receiving room if it can be
assumed that the airborne sound generated in the source room does not cause an unwanted
flanking path. This can be used to determine the normalized flanking impact sound pressure
level (Eq. 3.211) for each specific flanking path.
The use of structure-borne excitation and vibration measurements is well-suited to elements
with a reverberant bending wave field and no significant decrease in vibration with distance.
This allows use of vibration measurements to determine parameters such as the coupling loss
factor or vibration reduction index that can be used in SEA or SEA-based models. For such
measurements it is not necessary to have a laboratory with reverberant rooms commissioned
for sound pressure level measurement. This allows some flexibility in choosing where measurements are carried out. However, the way in which a test junction is installed in a laboratory
facility affects the flow of vibrational energy between the test elements as well as the way in
which those elements transmit energy into the laboratory structure.
There are many different types of test junction and many possible ways of installing a junction
within a laboratory. For this reason, additional measurements are needed to assess flanking
transmission between elements of the test junction via the laboratory structure. To make a full
assessment of flanking transmission via all possible paths would require a prediction model
of the test junction and the laboratory; however if such a model is available there is no need
to make measurements. It is possible to overcome this problem if it can be assumed that the
laboratory structure can be described by SEA plate subsystems. An assessment can then
be made based on principles of energy flow from SEA (Section 4.2). As an example of this
we consider a junction of three walls installed as a free-standing element in a laboratory as
shown in Fig. 3.96. The aim of the measurement is to determine the velocity level difference,
Dv,ij between two elements i and j with excitation of element, i. If elements i and j are positioned on the same side of the vibration break in the laboratory floor there will be an unwanted
flanking path between the separating wall and the flanking wall via the floor. For this unwanted
flanking path to have negligible effect on Dv,ij , when element, i, is excited, there must be
378
Chapter 3
Receiving element, j
(flanking wall)
Power
input
Laboratory
element, k
Source element, i
(separating wall)
Figure 3.96
Example assessment of flanking transmission due to connections between the test junction and the laboratory. All test
elements and laboratory elements are assumed to act as SEA subsystems. The solid arrow indicates the flanking path of
interest for which vibration measurements are being used to give the velocity level difference, Dv,ij . The dotted arrow indicates
an unwanted flanking path via laboratory element, k.
a positive net flow of vibrational energy from receiving element, j, to laboratory element, k.
This assumes that elements i, j, and k all act as SEA plate subsystems and there is only
bending wave motion. Hence when element i is excited by a structure-borne sound source,
element, j must have higher modal energy than element, k; this can be written as
Ej
Ek
>
nj (f )B
nk (f )B
(3.212)
where E is energy (Eq. 2.237), n(f ) is the bending wave modal density (Eq. 2.139), and B is
the bandwidth of the frequency band.
Equation 3.212 can now be written in terms of a velocity level difference, Dv,jk , between
elements j and k that is measured during excitation of element i. This gives the flanking
criterion as
ρs, j nk (f )
Dv, jk + 10 lg
> 0 dB
(3.213)
ρs,k nj (f )
This is more conveniently written in terms of the quasi-longitudinal phase velocity or critical
frequency of the plate elements to give the following inequalities (ISO 10848 Part 1)
ρs,j fc,k
ρs, j cL,j
(3.214)
> 0 dB or Dv, jk + 10 lg
> 0 dB
Dv, jk + 10 lg
ρs,k cL,k
ρs,k fc, j
The above approach is not appropriate when each element in the test junction is connected to
laboratory walls/floors that are too small to be considered as SEA plate subsystems.
Constructions can be built into flanking laboratories in different ways. One extreme is where
all elements in the test junction are rigidly connected to the laboratory structure on all sides to
form two (or more) box-shaped rooms. The other is where each element in the test junction is
isolated from the laboratory structure with resilient materials or is supported on small laboratory floor elements that are all isolated from each other by vibration breaks. Neither option is
ideal for all types of test element. For heavyweight elements the total loss factor of each element plays an important role in determining the vibration transmission between the elements.
The effect of the total loss factor for each element on the coupling loss factor or vibration
reduction index is discussed later with examples in Section 5.2.3.
379
S o u n d
3.12.2
I n s u l a t i o n
Ranking the sound power radiated from different surfaces
In completed buildings where the sound insulation needs to be improved, it is necessary to
identify which room surfaces need remedial treatment, such as a wall lining or an independent
ceiling. Vibration and sound intensity measurements are useful diagnostic tools for this purpose. For both measurements the radiated sound power is determined for the various surfaces
that face into the receiving room. In many situations these measurements can be used to rank
order the sound power radiated by different surfaces in the room. This allows remedial treatments to be prescribed for the surfaces that radiate high sound power levels, and an estimate
can be made of the potential improvement in the sound insulation.
3.12.2.1 Vibration measurements
For bending wave vibration, the sound power radiated by a surface is calculated from the temporal and spatial average mean-square velocity over the surface, v 2 t,s , using the following:
W = Sρ0 c0 σv 2 t,s
(3.215)
where σ is the radiation efficiency.
Due to spatial variation of plate vibration over the plate surface (Section 2.7), a sufficiently
large number of accelerometer positions are needed to calculate the spatial average value.
Even for masonry walls in the high-frequency range, a single position can be up to 10 dB higher
than the spatial average level due to material imperfections. Therefore the potential for errors
is large if only a single accelerometer position is used to estimate the radiated power.
The radiation efficiency that is needed for Eq. 3.215 can be calculated for homogeneous
isotropic plates when the critical frequency is known (Section 2.9.4). Above the critical frequency it can usually be assumed that the radiation efficiency is unity. The critical frequency
of a solid masonry/concrete plate tends to occur in the low-frequency range, so a large part
of the building acoustics frequency range can often be covered with this simple assumption.
In contrast this will only cover a few frequency bands in the high-frequency range for many
lightweight walls and floors.
The limitations of this method are that it cannot be applied to homogeneous walls or floors
unless the critical frequency can be estimated, nor to walls and floors that do not have a
reverberant bending wave field. This is awkward because construction details in an existing
building that has been decorated can be difficult to ascertain. In addition it assumes that
non-resonant transmission below the critical frequency is insignificant compared to resonant
transmission; this can be assessed from the theory and examples in Section 4.3.1. A more
practical limitation is that attaching accelerometers to walls and ceilings in the field tends to
mark or damage decorated surfaces, and they cannot be attached to floors with fixed coverings
such as carpets. All the above limitations can be overcome by using sound intensity.
3.12.2.2
Sound intensity
Sound intensity is a powerful tool with which to assess flanking transmission in buildings, and
it avoids the limitations of vibration measurements in quantifying the radiated sound power
(Villot and Roland, 1981). However, measurements over the complete building acoustics frequency range can take several hours if the majority of receiving room surfaces need to be
measured and reactive sound fields make it difficult to take valid measurements. Field and
380
Chapter 3
laboratory measurements almost always require additional absorption to be introduced into
the receiving room. In the field it is rarely possible to measure every single frequency band
between 100 and 3150 Hz whilst satisfying the requirements on FpI and δpI0 − FpI (ISO 15186
Part 2). Even in the laboratory there are often single bands in the low- or high-frequency range
that are difficult to measure (usually ≤125 Hz and ≥2000 Hz). Depending on the time available
for the field test it is not uncommon to leave the test site with valid measurements and positive
intensity readings in less than half the frequency bands over the building acoustics frequency
range for the majority of room surfaces. Fortunately it is possible to identify solutions to most
flanking transmission problems without needing measured data in every single frequency band
(e.g. see Carman and Fothergill, 1990; Hongisto, 2001).
In the field, scanning is easier to carry out than discrete point measurements. The sound power
for each surface is calculated from the temporal and spatial average normal sound intensity
level, LIn , and the measurement surface area, SM , using
LW = LIn + 10 lg SM
(3.216)
At the outset it is important to establish which surfaces are likely to be the dominant radiating
surfaces; there will not always be enough time to measure every surface in a room. A short
measurement on all surfaces with the intensity probe will indicate which ones are likely to have
the highest sound power, and on which surfaces it will be difficult to satisfy the requirements
on FpI and δpI0 − FpI . For each surface, the most practical option is to slowly scan across a
single diagonal line from corner to corner (Pettersen et al., 1997).
Scanning speeds are usually between 0.1 and 0.3 m/s (ISO 15186 Part 2), hence long measurement times needed to scan all the walls and floors can be problematic in the field. It is
unwise to increase the scanning speed, but time savings can sometimes be made by increasing the distance between the lines in the scanning pattern. With elements such as windows
and doors this is not appropriate because there are often leaks at edges and over the surface
so the sound radiation is rarely uniform. However for masonry/concrete elements where the
air paths have been removed by sealing the surface with a bonded surface finish (e.g. plaster),
it is often possible to double the distance between the lines in the scanning pattern without
significant loss of accuracy.
Both during and after the measurements, it is useful to check whether all the dominant radiating
surfaces have been measured. This is relatively simple if it can be assumed that there is a
diffuse sound field in the receiving room. Reverberation times are needed to calculate the total
absorption area, and the spatial average sound pressure level from the central zone of the
receiving room. Essentially we need to check whether the room energy due to sound radiated
from surfaces measured with the intensity probe is equal to the room energy calculated from
the reverberant sound pressure level measurement. Note that the latter accounts for all sound
transmission. To carry out this check it is necessary to account for higher energy density near
the room boundaries by using the Waterhouse correction to estimate the room energy from
the reverberant sound pressure level; this is more important for the low-frequency range. The
result can be expressed as an energy level difference, E, in decibels given by
N
ST λ
4
LWi /10
E = 10 lg
− Lp + 10 lg 1 +
(3.217)
+ 10 lg
10
A
8V
i=1
where LWi is the sound power level for radiating surface i. Note that E is the same as K2
defined in an informative annex of ISO 15186 Part 2.
381
S o u n d
I n s u l a t i o n
3
2
1
0
∆E (dB)
⫺1
⫺2
⫺3
⫺4
⫺5
⫺6
⫺7
Mean with 95% confidence intervals (N ⫽ 14)
⫺8
Minimum and maximum values
⫺9
50
80
125
200
315
500
800
1250
One-third-octave-band centre frequency (Hz)
2000
3150
Figure 3.97
E values from a flanking laboratory with 14 different masonry test constructions. The receiving rooms used for the
measurements had volumes of ≈ 50 m3 with one separating element and one, two, or three flanking elements. Measured
data from Hopkins are reproduced with permission from ODPM and BRE.
The factors that affect interpretation of E are measurement uncertainty in the sound pressure,
sound intensity, and reverberation time. In addition there are errors due to the assumption of
a diffuse field and the associated fact that the Waterhouse correction is only usually a good
estimate for empty box-shaped rooms with a minimum volume of 50 m3 . However, when E
equals 0 dB and the intensity measurements satisfy the requirements on FpI and δpI0 − FpI it is
reasonable to assume that all the dominant radiating surfaces have been measured accurately.
Negative values of E are an indication that not all the radiated sound has been accurately
measured using sound intensity. Positive values of E can occur in the low-frequency range
due to a non-diffuse sound field which invalidates the Waterhouse correction and increases
the measurement uncertainty.
Example values of E from masonry/concrete test constructions in a flanking laboratory are
shown in Fig. 3.97. It is usually easier to measure under laboratory conditions than in the field;
hence it will rarely be possible to achieve E = 0 dB over a wide frequency range in the field.
Based on these measurements, an acceptable range for field measurements can be taken as
−2 < E ≤ 0 dB. It can sometimes be difficult to achieve values in this range with lightweight
constructions (ISO 15186 Part 2). This is partly due to the complexity of sound radiation from
lightweight walls and floors that form flanking elements. When lightweight walls and floors
are excited along one junction line, there is often a decrease in vibration level with distance
away from the junction; this gives a corresponding decrease in the radiated sound power
(Nightingale, 1996). This causes some measurement problems. Firstly, the volume enclosed
by the measurement surface for the separating wall may contain strips of flanking wall/floor
that radiate significant sound power; if so, additional linings will be required to reduce their
sound radiation. Secondly, it may be difficult to accurately quantify the sound power that is
radiated by a flanking element if there is a rapid decrease in the radiated sound intensity level
with increasing distance from the excited junction.
382
Chapter 3
Field measurements are often taken in furnished buildings and it is not always possible to fix
additional linings to the strongly radiating surfaces that are close to the measurement surface.
If the airborne sound insulation is predominantly determined by a separating element with
relatively low sound insulation then it can be difficult to measure radiation from flanking walls
and floors. A common example is an internal floor (Rw ≈ 40 dB) that needs to be upgraded to
a separating floor in flat conversions. Quantifying the flanking transmission from the existing
walls with sound intensity measurements may not be possible until the sound insulation of the
floor has been upgraded.
Building façades are often formed from several elements that are adjacent to each other (e.g.
wall, infill panels, curtain walling, window, door). Before using sound intensity as a diagnostic tool to identify the dominant radiating surface(s) it may be necessary to shield the other
surfaces using additional linings (e.g. see Vermeir et al., 1996). Care needs to be taken in
shielding the various elements because the dominant sound source will sometimes be the
frame or gap in-between the elements.
It is important to note that sound intensity measurements quantify sound radiation from separating and flanking elements; they do not automatically quantify the sound transmitted along a
certain transmission path. If we consider sound transmission between two rooms in terms of
the paths Dd, Fd, Df, and Ff (previously defined in Fig. 3.14) it is clear that the sound power
radiated by a separating or flanking element can result from more than one transmission path.
However, if the test construction is arranged in a similar way to the laboratory set-ups in
Figs 3.94 and 3.95 and there is only one dominant flanking path, then measurements can
indeed be used to quantify transmission via this path.
Two examples are now used to illustrate the ranking of sound power from different surfaces.
They also indicate the influence of flanking transmission at levels of airborne sound insulation
commonly required in building regulations. The overall sound insulation is not only determined
by the separating wall or floor; there are many transmission paths involving the separating and
flanking elements that determine the sound insulation in situ. The first example compares the
radiated sound power from intensity and vibration measurements. This is shown in Fig. 3.98
for a masonry separating wall in a masonry building. The intensity measurements show that
the separating and flanking walls radiate similar levels of sound power over a wide frequency
range and at some frequencies the flanking wall radiates more than the separating wall. Above
the critical frequency of each wall, the sound intensity and vibration measurements are in
close agreement. Below the critical frequency there will be errors in the vibration measurements; these are mainly due to the predicted radiation efficiency and partly due to non-resonant
transmission below 100 Hz. For heavyweight walls and floors (without lightweight linings) the
required insight into the flanking transmission can often be gained from either measurement
method. For lightweight walls and floors, reliance is placed on sound intensity measurements
due to the complexity in predicting the radiated sound from vibration measurements. The second example in Fig. 3.99 shows the sound powers for a separating timber floor and flanking
walls in a timber frame building. In this example E generally falls in the range −2 < E ≤ 0 dB;
note that the remaining flanking wall that contained the entrance door was not scanned due to
a lack of time. At low frequencies it was only the separating floor (surface 1) that gave positive intensity satisfying the requirements on FpI and δpI0 − FpI . In the mid-frequency range, the
sound power radiated by the separating floor was approximately equal to the combined sound
power radiated by all the flanking walls (surfaces 2 to 6).
383
S o u n d
I n s u l a t i o n
75
Separating wall (Sound intensity)
Flanking wall (Sound intensity)
70
Separating wall (Vibration)
Flanking wall (Vibration)
65
Sound power (dB)
60
55
50
45
Separating wall
Flanking
wall
40
DnT,w = 56 dB
35
Both walls form part of
masonry cavity walls
Separating wall:
ρs ⫽ 170 kg/m2
fc in 250 Hz band
Flanking wall:
ρs ⫽ 70 kg/m2
fc in 250/315 Hz band
30
50
63
80 100 125 160 200 250 315 400 500 630 800 1000 1250 1600 2000
One-third-octave-band centre frequency (Hz)
Figure 3.98
Example measurement of sound power radiated by separating and flanking walls in a heavyweight construction. Comparison
of measurements using sound intensity and vibration. Sound power derived from vibration measurements is shown with the
95% confidence intervals for the spatial average velocity level. Measured data from Hopkins are reproduced with permission
from ODPM and BRE.
3.12.3
Vibration transmission
Vibration transmission can be quantified using measurement of structural intensity or the velocity level difference; the latter is the simpler method. These measurements are usually required
to calculate coupling parameters for use in an SEA or SEA-based prediction model. The two
most widely used parameters are the coupling loss factor and the vibration reduction index. As
these parameters have their basis in SEA, the basic assumptions made in SEA (Section 4.2)
need to be satisfied by the elements being measured.
3.12.3.1 Structural intensity
Structural intensity measurement can be used to quantify the net structural power; sometimes
called structure-borne sound power. This gives an insight into the net energy flow that lies
‘underneath’ the reverberant vibration field. Bending waves tend to be the most important
wave type for sound insulation in buildings and it is unusual to need measurements of in-plane
wave intensity. For this reason only structural intensity measurement of bending wave vibration
is considered here. As with sound intensity, it is the active component of intensity that is useful
rather than the reactive component. The magnitude and direction of the active intensity vectors
changes with time, hence the time-averaged value for stationary vibration signals is needed
for most practical purposes. In contrast to sound intensity measurement, we cannot place a
384
Chapter 3
80
Source room is directly above
surface No.1 (separating floor)
DnT,w = 63 dB
70
60
6
5
3
2
Sound power (dB)
Surface No.
50
1
4
(both sides)
2
40
3
4
5
30
6
1
20
10
⫺1
∆E (dB)
0
∆E
0
100
125
160
200
250
315
400
500
630
One-third-octave-band centre frequency (Hz)
800
1000
⫺2
1250
Figure 3.99
Example measurements of sound power radiated by a separating floor and flanking walls in a lightweight construction using
sound intensity. Measured data from Hopkins and Turner are reproduced with permission from ODPM and BRE.
structural intensity probe at a point in space where the energy flow takes place; this lies within
the structure. However, by taking measurements on the surface of a structure, the results can
be related to energy flow within its cross-section. Structural intensity is therefore quoted in
terms of the net power per unit width (W/m) across a line on the surface.
Structural intensity measurement for bending waves on beams or plates with a twoaccelerometer probe was formulated and implemented by Noiseux (1970). The assumptions
in this approach limit its validity to propagating bending waves on thin plates where nearfields
are negligible. A more rigorous approach to structural intensity measurement that avoids the
assumption of measuring in the far-field was developed by Pavić (1976). This requires intensity
probes that comprise arrays of four accelerometers for beams and eight accelerometers for
plates. However, as the size of the array increases beyond two accelerometers, errors in the
signal processing and the positioning of the accelerometers tends to reduce the advantage
of using a large array to give a full description of the structural intensity (Bauman, 1994). For
practical measurements on building structures it is usually sufficient to use a two-accelerometer
probe as shown in Fig. 3.100 (Kruppa, 1986).
By deriving the structural intensity for bending wave motion in the x-direction on a thin homogeneous isotropic plate, the result can be simply adapted to thin beams. We will use the sign
conventions and variables for bending waves on thin beams and plates defined in Section 2.3.3
for positive power flow in the positive x-direction. Referring back to this bending wave theory,
385
S o u n d
I n s u l a t i o n
Accelerometers
No.1 No.2
z
d
x
Figure 3.100
Structural intensity measurement (bending wave motion): a–a probe.
we see that it is most practical to formulate all variables in terms of the lateral displacement,
η; this displacement is relatively large and easily determined from acceleration measurements
using accelerometers.
The time-averaged structural intensity in the x-direction, Ix , has two components: a force
component, IxF , and a moment component, IxM (Noiseux, 1970) such that,
Ix = IxF + IxM
(3.218)
9
:
∂ ∂2 η
∂η
∂2 η
IxF = Qx νz t = Bp
+
∂x ∂x 2
∂y 2
∂t t
(3.219)
where the force component is
and the moment component which comprises bending and twisting moment contributions, IxB
and IxT , is
9
2
2
2 :
4
3
∂ η
∂2 η
∂2 η
∂ η
∂ η
IxM = IxB + IxT = Mxy ωy + Mxx ωx t = −Bp
+
ν
+
B
(1
−
ν)
−
p
∂x 2
∂y 2
∂t∂x
∂x∂y
∂t∂y t
(3.220)
Hence the time-averaged structural intensity in the x-direction is
9 2
2
:
∂ ∂ η
∂ η
∂2 η ∂η
∂2 η ∂2 η
∂2 η ∂2 η
Ix = Bp
+
−
+
ν
−
(1
−
ν)
∂x ∂x 2
∂y 2 ∂t
∂x 2
∂y 2 ∂t∂x
∂x∂y ∂t∂y t
(3.221)
The same approach applies to the structural intensity in the y-direction, Iy , which yields:
2
:
9 2
∂ η
∂2 η ∂η
∂2 η ∂2 η
∂2 η ∂2 η
∂ ∂ η
+
−
+
ν
−
(1
−
ν)
(3.222)
I x = Bp
∂y ∂x 2
∂y 2 ∂t
∂y 2
∂x 2 ∂t∂y
∂x∂y ∂t∂x t
Hence the resultant intensity vector, I, is
I = Ix i + Iy j
(3.223)
where the magnitude is
|I| =
Ix2 + Iy2
(3.224)
and the direction in terms of the angle, θ, from the x-axis is
θ = arctan
386
Iy
Ix
(3.225)
Chapter 3
For a thin beam we only need to consider the structural intensity in the x-direction. This is
found from Eq. 3.221 by replacing Bp with Bb and removing all terms involving the y-direction
to give:
Ix = Bb
9
∂3 η ∂η
∂2 η ∂2 η
− 2
3
∂x ∂t
∂x ∂t∂x
:
(3.226)
t
Having formulated the equations that describe the structural intensity, we now look at how
a two-accelerometer probe can be used to estimate this intensity. To simplify the measurement procedure, Noiseux (1970) noted that for a plane bending wave field in the absence of
nearfields, the force component is equal to the moment component. Noiseux’s proposal for
measurements in plane wave fields was to measure only the moment component and assume
that IxF = IxM . However, the moment component contains the angular velocity term, ∂2 η/∂x∂y,
which is obtained by using four accelerometers (Pavić, 1976). To avoid using more than two
′
accelerometers, Noiseux introduced a modified moment component, IxM
, which, using the
notation in Chapter 2 is
′
IxM ≈ IxM
=
9
Mxy − Myx
ωy
1+ν
:
t
= −Bp
9
∂2 η
∂2 η
+
∂x 2
∂y 2
∂2 η
∂t∂x
:
(3.227)
t
The term in brackets can be simplified for plane waves in the free-field, where
∂2 η
∂2 η
+ 2 = kB2 η
2
∂x
∂y
(3.228)
This gives a displacement, whereas we normally measure acceleration. So, if we assume that
the signal is a sinusoid of frequency, ω, or a narrow band of noise with a centre frequency, ω,
then,
η=
1 ∂2 η
(iω)2 ∂t 2
(3.229)
Hence the modified moment component in Eq. 3.227 can be written in the form:
′
IxM
=
9
:
Bp kB2 ∂2 η ∂2 η
ω2 ∂t 2 ∂t∂x t
(3.230)
This requires measurement of the time-averaged value of the product of the lateral acceleration
and the angular velocity. Noiseux proposed use of two accelerometers: one to measure the
lateral acceleration and the other to measure the angular acceleration. Since this time, single
transducers have become available that can output both lateral and angular acceleration; these
can work well in practice (e.g. see Bauman, 1994). However, the modified moment component
also allows a two-accelerometer probe to be used where the accelerometers are positioned
side-by-side, and separated by a distance, d between their centres (see Figs 3.100 and 3.101).
Using this approach allows the angular velocity to be determined using a finite-difference
approximation. This has similarities to sound intensity measurement with a p–p probe where a
finite-difference approximation is used to determine the particle velocity; hence we will refer to
an a–a probe for structural intensity. This approach conveniently allows use of a two-channel
analyser that is capable of sound intensity measurement (Rasmussen and Rasmussen, 1983).
It also allows analysis of the measurement errors in a similar way to sound intensity.
387
S o u n d
I n s u l a t i o n
y
Propagation direction
η1
θ
d
η2
x
d
Figure 3.101
Orientation of two a–a probes in a propagating plane wave field to measure structural intensity in the x- and y-directions.
3.12.3.1.1 a–a structural intensity probe
An a–a probe comprises two accelerometers: No.1 and No.2, for which the associated lateral
displacements are η1 and η2 respectively (Fig. 3.101). For positive intensity, a plane bending
wave propagates in the x-direction from accelerometer No.1 to No.2. The accelerometers
are equally spaced about the point for which the intensity estimate is made, so the lateral
acceleration at the mid-point between the accelerometers is
∂2 η
1 ∂2 η1
∂2 η2
=
+
(3.231)
∂t 2
2 ∂t 2
∂t 2
and the angular velocity is estimated using a finite-difference approximation,
2
∂2 η
1 ∂η2
∂η1
∂2 η1
1
∂ η2
=
−
−
dt
=
∂t∂x
d ∂t
∂t
d
∂t 2
∂t 2
Hence, the time-averaged structural intensity in the x-direction is
9
2
:
2Bp kB2 1 ∂2 η1
∂2 η2 1
∂ 2 η1
∂ η2
′
+ 2
− 2 dt
Ix = 2IxM =
ω2
2 ∂t 2
∂t
d
∂t 2
∂t
t
(3.232)
(3.233)
The probe measures the intensity component along its axis; hence Iy is measured by rotating
the probe by 90◦ .
For sound intensity measurements made with a p–p probe, the sound pressure signals are
processed according to Eq. 3.170, in either one-third-octave or octave-bands. The similarity
between Eqs 3.170 and 3.233 means that structural intensity can be measured using a dualchannel analyser designed for sound intensity (Rasmussen and Rasmussen, 1983). Each
channel needs to be calibrated to measure acceleration in decibels re 20 × 10−6 m/s2 , and the
388
Chapter 3
spacing, d, must be set on the analyser to the accelerometer spacing. This gives the structural
intensity level in decibels re 10−12 W/m when the intensity level from the analyser is corrected
by adding:
ρ0 Bp ρs
for plates
(3.234)
10 lg
πf
and
10 lg
3.12.3.1.2
ρ 0 B b ρl
for beams.
πf
(3.235)
Structural power measurement
Measurement of net structural power with an a–a probe is defined in a similar way to sound
power, but using one-dimension rather than two. Instead of using a surface integral, a line integral is used to calculate the net structural power. A measurement line is chosen that encloses
the source that is of interest; we are usually interested in the net structural power transmitted
across a junction of plates and/or beams. Examples of measurement lines on the receiving
plate are shown in Fig. 3.102.
The axis of the a–a probe must be perpendicular to the measurement line so that the intensity
component, In , that is being measured is normal to this line, and is defined by Eq. 3.173. For N
equally spaced measurement points along a measurement line of length, L, the net structural
Power
input
Power
input
Power
input
Figure 3.102
Examples of measurement lines used to determine the net structural power.
389
S o u n d
I n s u l a t i o n
power, Wnet , is
N
Wnet =
L
In,i
N
(3.236)
i=1
3.12.3.1.3 Error analysis
Structural intensity is usually used to determine the power flowing across a line (i.e. measuring
Ix or Iy ) or to determine the vector intensity (i.e. measuring Ix and Iy ). In a similar way to
sound intensity measured with a p–p probe, there will be errors due to the finite-difference
approximation and phase-mismatch. We will assume a plane bending wave propagating at an
angle, θ, to the x-axis as previously shown in Fig. 3.101.
The normalized errors, eFD (Ix ) and eFD (Iy ), due to the finite-difference approximation are:
1
eFD (Ix ) ≈ − (kB d cos θ)2
6
and
1
eFD (Iy ) ≈ − (kB d sin θ)2
6
(3.237)
The normalized errors, ePM (Ix ) and ePM (Iy ), due to phase-mismatch are:
ePM (Ix ) ≈
±φPM
kB d cos θ
and
ePM (Iy ) ≈
±φPM
kB d sin θ
(3.238)
The combined normalized error due to the finite-difference error and phase-mismatch is calculated in the same way as described in the section on sound intensity. This combined error
can be used to define an accelerometer spacing on a plate or beam of specific material and
thickness by choosing a tolerable error (e.g. ± 5%). The magnitude of the structural intensity
will only be valid for thin plates, i.e. below the thin plate limit, fB(thin) . This means that in contrast
to the p–p sound intensity probe, all valid frequency bands in the building acoustics frequency
range can often be covered by using a single spacing. For a probe with |φPM | < 0.4◦ and
measurements taken on masonry/concrete walls or floors, an appropriate spacing is usually
between 50 and 100 mm.
Unlike sound intensity, there is no well-established set of indicators that can be used to validate a structural intensity measurement. For sound intensity measurements the reactivity of
the sound field can be reduced by adding absorbent material into the room. In buildings there
is little or no possibility of reducing the reactivity of the vibration field by introducing additional
absorption at the boundaries of walls, floors, or columns. Hence structural intensity measurements are always taken on structures that are reverberant to various degrees. On very
reverberant structures, the phase difference of the active intensity component may be small,
and phase-mismatch may be problematic.
For a stationary vibration signal, phase-mismatch can potentially be removed by using probeswitching. This involves a repeat measurement at each position when the accelerometer
positions have been rotated by 180◦ , i.e. switched over. The improved estimate of the intensity
is then calculated using
Ix,0◦ − Ix,180◦
(3.239)
2
are the measured intensity values before and after probe-switching
Ix =
where Ix,0◦ and Ix,180◦
respectively.
If the phase-mismatch is negligible, probe-switching should give intensity of the same magnitude but opposite sign. If the phase-mismatch is greater than the actual phase difference in
390
Chapter 3
the vibration field the intensity will not change sign after probe-switching. (In principle this can
be done with a sound intensity probe, but it is often easier to ensure accurate repositioning in
two dimensions rather than three.) When switching the accelerometers it should be ensured
that the fixing strength is the same (particularly when using bees wax), and that the accelerometer’s axis of minimum transverse sensitivity is not aligned at a different angle relative to the
probe axis.
Commercial sound intensity probes come with matched pairs of microphones to minimize
phase-mismatch errors. In contrast, individual accelerometers for structural intensity probes
are usually taken off the shelf. Problematic phase-mismatch (typically |φPM | > 0.4◦ ) can often
be avoided by using accelerometers with consecutive serial numbers. However, the phase
difference between two accelerometers should always be checked, even if probe-switching is
to be used. It can be measured by rigidly fixing one accelerometer on top of the other, and
then fixing the pair of accelerometers on top of an electrodynamic shaker so they are exposed
to the same vibration signal. Dual-channel FFT analysis can then be used to determine the
phase from the Frequency Response Function.
Measurement using an a–a probe is only valid for plane bending waves in the absence of
nearfields. However, it is difficult to quantify the error and offer definitive guidance on appropriate distances from junctions or structural discontinuities. On masonry/concrete elements a
distance of 0.3 to 0.6 m from a junction is usually adequate for the building acoustics frequency
range (Craik et al., 1995; Kruppa, 1986). On lightweight elements such as plasterboard walls,
the probe will often be close to stud connections. However, it is difficult to generalize due to
the frequency-dependent effects of different screw spacings and different types of studs (e.g.
heavy thick timber or light thin steel).
Above the thin plate limit (but below the fundamental quasi-longitudinal mode) the calculated
error in the measured magnitude of the structural intensity is an overestimate of the actual
intensity by more than 20% (Maysenhölder, 1990). However, when it is only the direction of
the intensity vector that is of interest on homogeneous isotropic plates, it should still be possible
to gain reasonable estimates above the thin plate limit.
When measuring two-dimensional intensity vectors, the finite-difference and phase-mismatch
errors combine to give an absolute error, ε(θ), in radians for the propagation angle, θ, where
⎞⎤
⎡
⎛
±φPM
1
sin θ 1 +
− (kB d sin θ)2
⎢
⎜
⎟⎥
kB d sin θ
6
⎜
⎥
⎟
ε(θ) = ⎢
(3.240)
⎣arctan ⎝
⎠⎦ − θ
±φPM
1
− (kB d cos θ)2
cos θ 1 +
kB d cos θ
6
3.12.3.1.4 Visualizing net energy flow
Structural intensity measurements can be used to visualize net energy flow across the surface
of a plate. Typically, a square grid is drawn over the plate surface so that the measurement
points are defined by the intersections of the grid lines. This results in a grid of active intensity
vectors over the surface.
Interpreting a grid of measured vectors is not always intuitive, even when the measurement
errors are insignificant. At a specific point, or along a line on the surface, there can be an
outgoing or incoming flow of vibrational energy; referred to as a source or a sink respectively.
Sources and sinks are usually the features of interest. However, individual bending wave
391
S o u n d
I n s u l a t i o n
Circulating vectors
Source
Sink
Circulation point
Figure 3.103
Idealized examples of a source, a sink, and circulating vectors.
modes on a plate give rise to a circulatory flow of active intensity vectors that sometimes make
it difficult to identify these features (Cuschieri, 1991; Noiseux, 1970). Fortunately, strongly
circulating vector fields are not usually evident when the frequency band is sufficiently wide,
the response is multi-modal, and the plate has sufficiently high damping. One-third-octaveband measurements in the low-frequency range on masonry/concrete plates may show some
circulation (see Kruppa, 1986). Note that circulatory active intensity due to individual modes
also occurs in two- or three-dimensional sound fields (Fahy, 1989). When discussing vector
fields it is useful to introduce another term, circulation point; this is defined as a grid point at
which the energy flow is simply passing through (Maysenhölder and Schneider, 1989). These
points are of no specific interest, but it is important to be aware of them as they may cause
confusion in the identification of a source or a sink. Idealized examples of a source, sink, and
circulating vectors are shown in Fig. 3.103. To identify these features it is necessary to make
sure that the grid has a sufficient number of points; if the grid spacing is too coarse, some
features may be interpreted incorrectly.
The sound insulation problem under investigation may involve an airborne or a structure-borne
sound source. In general, the latter is more likely to result in useful measurements of structural
intensity. Grid measurements on walls or floors that face into a source room excited by an
airborne source tend not to be particularly informative (Kruppa, 1986). This is primarily due to
difficulty in measuring active intensity in a highly reactive vibration field.
An example is now used to look at some of the features that can be encountered in practice
with structural intensity measurements on a masonry construction (Hopkins, 2000). The test
construction used for the measurements is shown in Fig. 3.104. In this set-up the separating
wall is excited directly by a structure-borne sound source that transmits vibration to the flanking
wall across the corner junction. Measurements were taken before and after the introduction
of a window opening into the flanking wall. Two different grids were used on the flanking wall;
a relatively coarse grid of 13 × 7 positions (entire flanking wall) and a more detailed grid of
12 × 15 positions over the area around the opening. One-third-octave-band measurements
were taken with probe switching to remove the phase error, and a probe spacing of 50 mm.
The structure-borne sound source was a plastic headed hammer. This source was used to
392
Chapter 3
Separating wall
LJ
Flanking wall used
for structural
intensity
measurements
Figure 3.104
Sketch of the test construction used for structural intensity measurements. Flanking wall: 100 mm masonry wall (Lx = 4 m,
Ly = 2.4 m, ρs = 70 kg/m2 , cL = 2370 m/s, solid aircrete blocks with mortar on each side, 13 mm plaster finish). Window
opening: 0.9 × 1.2 m, LJ = 0.25 m.
excite the separating wall at many different positions during a 20 s period with ≈ 5 hammer hits
per second. A selection of the structural intensity plots are shown in Fig. 3.105.
An example of disordered vectors is seen in the 50 Hz band for the wall with an opening; note
that this could potentially occur on walls or floors without openings. This seemingly random
array of vectors is shown to emphasize the importance of using a sufficiently large number of
measurement positions when trying to identify the position of a source.
The flanking wall is far from being an ideal plate. The lintel forms a discontinuity, and with
a window opening there are only narrow strips of wall near the edges of the opening. When
measurements are used to quantify the net structural power transmitted from the separating
wall to the flanking wall, the measurement line is usually 0.3 to 0.6 m from a junction. This
gives a limited number of measurement positions underneath and above the opening. The
modified moment component and the assumption that the force component equals the moment
component is not valid near the free edges of an opening. However, the magnitude of the
intensity in the direction perpendicular to the junction is not significantly in error (Hopkins,
1997). The measured vectors generally point in a direction parallel to the free boundaries of
the opening as would be expected for the true intensity vectors.
Examples of well-ordered vector flow can be seen for the wall with and without an opening.
The vectors either head towards the upper wall boundary or towards the left wall boundary.
The discontinuity formed by the lintel limits the length of the upper boundary over which the
vectors can head towards the in-line junction that leads to the flanking wall on the first floor.
393
S o u n d
I n s u l a t i o n
(a) Flanking wall with and without a window opening.
50 Hz one-third-octave-band
80 Hz one-third-octave-band
100 Hz one-third-octave-band
250 Hz one-third-octave-band
400 Hz one-third-octave-band
1000 Hz one-third-octave-band
Figure 3.105
Measured structural intensity vectors on the flanking wall shown in Fig. 3.104 with and without a window opening (vector
magnitude in decibels). Measured data are reproduced with permission from Hopkins (2000).
394
Chapter 3
(b) Flanking wall with a window opening (detailed grid)
250 Hz one-third-octave-band
400 Hz one-third-octave-band
Figure 3.105
(Continued)
3.12.3.1.5 Identifying construction defects
High levels of sound insulation are often achieved by isolating two parts of a construction.
For example a floating floor is isolated from the floor base by a resilient layer, and the leaves
of a cavity wall are sometimes isolated from each other over their surface (i.e. no wall ties,
lintels, etc.) and only connected at the foundations. Defects in the construction can remove
this isolation and significantly reduce the sound insulation. A common example is the bridging
of a floating screed floor via a gap or hole in the resilient layer; screed fills the hole and makes
a rigid connection between the screed and the floor base. Vibration levels by themselves often
give little or no clue in detecting the connection point on reverberant plates. Similarly, placing
a pressure microphone or intensity probe at successive positions on a grid near the surface of
the plate is unlikely to identify the connection point. Structural intensity offers the possibility of
tracking down structural connections between plates.
A procedure to identify the location of a point source, such as a rigid connection on an isotropic,
homogeneous plate is described by Maysenhölder and Schneider (1989). The first step is to
divide the rectangular plate into four equal areas and measure the intensity vector at the centre
of each of these areas. This gives a rough indication of the region containing the source. Subsequent measurements can then be used to more accurately identify its position. An example
to identify the position of a single rigid connection between the leaves of a masonry cavity wall
is shown in Fig. 3.106 (Maysenhölder and Schneider, 1989).
Structural intensity measurements have also been used to track down a bridged screed in
a factory where there was excessive transmission of vibration from a structure-borne sound
source (Sorainen and Rytkönen, 1989). However, an airborne and a structure-borne sound
source may give similar patterns of vectors (Rasmussen and Rønnedal, 1992). This makes
it necessary to look at structural intensity measurements alongside other measurements
(or predictions of the sound transmission paths) before coming to a conclusion. As with any
measurement of structural intensity on reverberant plates, it is important to check that the
phase error is low and the results are not affected by the residual intensity (Craik et al., 1996).
395
S o u n d
I n s u l a t i o n
Figure 3.106
Structural intensity measurements used to identify the position of a single sound bridge (indicated by ) between two leaves
of a masonry cavity wall. Shaker excitation was used on one leaf with structural intensity measurements on the other leaf
to determine the frequency-average intensity (500 to 2000 Hz). Intensity vectors are shown with arrows and vibration levels
are shown with circles for which the highest vibration levels have the largest diameter. Measured data are reproduced with
permission from Maysenhölder and Schneider (1989).
Success in identifying a connection partly depends on the width of the frequency band(s) or
frequency-average that is used to determine the structural intensity vectors. This is usually
determined by trial and error because the frequency range in which there is significant transmission via the connection is not usually known a priori. However, one-third-octave-bands
may be too narrow for this purpose. For point connected masonry/concrete plates, a bandwidth between 500 and 2000 Hz wide can be beneficial in identifying sources (Maysenhölder
and Schneider, 1989; Craik et al., 1996).
3.12.3.2
Velocity level difference
The velocity level difference is used to calculate vibration transmission between two elements
(e.g. walls, floors, columns) across a junction or some other type of connection. For the purpose of calculating the vibration reduction index, its measurement is described in the relevant
Standard (ISO 10848 Part 1).
One element is treated as the source element, which is excited by an airborne or a structureborne sound source; the other element is treated as the receiving element (see Fig. 3.107).
For connected walls and floors between a source and receiving room, there are many possible
permutations of source and receiving element. For this reason, the velocity level difference,
Dv,ij , is defined as the difference between the velocity level on source element, i, and receiving
element, j. Each element can therefore be numbered to simplify their identification.
At the outset it is important to establish the aim of the measurement. Usually it is to quantify
the difference between the bending wave vibration on elements, i and j; hence it is the velocity
perpendicular to the surface of each element that is required. Excitation of element, i, is
arranged to ensure that only bending waves are excited; this is most conveniently done with
a structure-borne sound source. Only element, i, can be the source element, so when using
airborne excitation via the sound field in a room, it is not possible for elements i and j to
396
Chapter 3
Element, j
Element, i
Excitation
Figure 3.107
Measurement of velocity level difference between two walls connected across an L-junction.
face into the same room. When the bending waves on the source or receiving plate impinge
upon the element boundaries and the junction, there is often some conversion to in-plane
waves. In this measurement, transmission via the junction is effectively treated as a black box;
bending waves are measured on the source and receiving elements, yet transmission across
the junction involving wave conversion is treated as an unknown process.
If the intention is to determine a spatial average velocity level difference over the surface of
each element with a structure-borne sound source, then it is important to ensure that there
is no significant decrease in vibration with distance across the element (Section 2.7.7). This
is particularly important for lightweight walls and floors because the measurements may be
specific to a particular layout of the frame and the sheets/boards; as well as specific to particular
excitation and measurement points.
Bending waves are usually of most interest, and it is rarely necessary to measure the velocity
level difference for in-plane wave motion on the source and/or receiving element. This is fortunate because interpretation of such measurements can be complex. The accelerometer does
not distinguish between transverse shear and quasi-longitudinal motion, and the presence of
bending waves can affect the measurement of in-plane motion (Craik, 1998).
Measurements are usually taken in one-third-octave or octave-bands over the building acoustics frequency range.
3.12.3.2.1 Stationary excitation signal and fixed power input
For structure-borne excitation of a wall or floor, a random noise signal can be sent to an
electrodynamic shaker that is pushed up against the element; this is only possible if the power
input is the same each time the shaker is moved to a different source position. In practice,
it is awkward and time-consuming to measure the power input; although it can be done with
structural intensity or by using a force transducer between the shaker and the element. It is
simpler for a floor, because the ISO tapping machine can be considered as a structure-borne
sound source that provides a fixed power input.
397
S o u n d
I n s u l a t i o n
For a stationary signal from either a structure-borne or airborne sound source, Dv,ij is calculated
from
(3.241)
Dv,ij = Lv,i − Lv, j
where Lv,i and Lv,j are the temporal and spatial average velocity levels determined from different
source positions and different accelerometer positions for each source position.
The temporal and spatial average value, standard deviation and 95% confidence interval can
be calculated from Eqs 3.18, 3.19, and 3.20 respectively. Corrections for background noise
can be made in the same way as for sound pressure measurements using Eq. 3.75.
3.12.3.2.2 Impulse excitation
An impulse can be generated using a hammer weighing ≈0.5 kg with a hard plastic head;
this provides a reasonably flat spectrum over the building acoustics frequency range, and
avoids excessive damage to the surface. One approach is to repeatedly hit the wall during a
defined measurement period. Multiple hits have the advantage of incorporating many different
source positions (Craik, 1982b). The decay of the signal in each filter band depends on the
reverberation time of the filter as well as the element. Masonry/concrete elements tend to have
short reverberation times, so to maintain a steady signal it is usually necessary to use at least
two hits per second. An alternative approach is to use a single impulse from a hammer hit.
A single hit is more suitable than multiple hits when the measurer is likely to cause vibration
whilst moving around the floor and cause an increased level of vibration on the receiving
element. An additional consideration for lightweight floors is that the mass of the measurer and
the measurement equipment on the floor must not affect the velocity level, or induce a static
deflection that alters vibration transmission via the junction under test.
A dual-channel analyser and a pair of accelerometers are needed to measure the vibration on
elements, i and j simultaneously. This is because of the variation in the energy spectral density
between each individual hammer hit. The term ‘excitation’ will now be used to refer to either
a single hit at different positions, or multiple hits over the surface. To account for the spatial
variation in vibration and ensure that different modes are excited, M different excitations are
applied with N different pairs of accelerometer positions for each excitation. For excitation, m,
and pair of accelerometer positions, n, the velocity level difference (Dv,ij )m,n is determined using
(Dv,ij )m,n
⎛
Tint
⎞
2
v
(t)dt
i
⎟
⎜ 0
⎟
= 10 lg ⎜
⎠
⎝ Tint
vj2 (t)dt
0
(3.242)
m,n
where Tint is the integration time for linear averaging.
Level differences from different excitations and different accelerometer positions are arithmetically averaged to give Dv,ij according to:
Dv,ij =
N
M
1
(Dv,ij )m,n
MN
(3.243)
m=1 n=1
The excitation is applied after t = 0, and the resulting signal must have decayed to a negligible
level (e.g. decreased by 60 dB) before t = Tint . This requirement needs to be balanced against
the need to avoid excessive averaging of background noise after the excitation has ceased;
398
Chapter 3
so Tint must not occur too long after the end of the excitation. With a single hit on highly
damped elements, Tint may only need to be 2 or 3 s. If so, it is better to automate the process.
A computer can be used to start and stop the analyser, and an audible tone can be emitted
to let the measurer know when the integration is starting and finishing. The frequency of this
tone must be well-above the highest frequency band being measured.
Although Dv,ij for an impulsive source is defined in terms of a ratio of velocities, it is usually
more accurate to use the acceleration signal from the charge amplifiers. This avoids potential
distortion of the impulse due to phase non-linearities in the charge amplifiers. There is no need
to convert each acceleration signal to velocity because Dv,ij is formed from a ratio.
In the light of non-linearities relating to structural reverberation times (Section 3.11.3.3), it is
prudent to pay attention to the strength of the hammer hits and the resulting vibration levels.
This issue can be avoided by measuring (Dv,ij )m,n using MLS shaker excitation at low levels.
On masonry/concrete elements, the MLS signal may need a red noise spectrum to give a
relatively flat acceleration spectrum on the source element. MLS is also advantageous when
the background level and/or the velocity level difference is high.
In Section 2.7 it was seen that there can be significant spatial variation of vibration over the
surface of a plate. Examples of standard deviations for Dv,ij between solid masonry walls
measured with impulse excitation are shown in Fig. 3.108. A generalized curve is included
on the figure because the curve shape can usually be described using just four points. The
assumption of a normal (Gaussian) probability distribution for the velocity level difference is
reasonable for solid masonry walls when there is no significant decrease in vibration level with
distance (Hopkins, 2000). The 95% confidence interval (Eq. 3.20) can therefore be calculated
from the standard deviation in decibels.
12
11
Measured – Average (N ⫽ 88)
10
Measured – Minimum and maximum
Standard deviation (dB)
9
Generalized curve shape
8
7
6
5
4
3
2
1
0
50
80
125
200
315
500
800 1250 2000 3150 5000
One-third-octave-band centre frequency (Hz)
Figure 3.108
Measured standard deviations for the velocity level difference, Dv,ij (hammer or MLS shaker excitation). Average, minimum
and maximum values calculated from 88 measurements on different combinations of solid masonry walls with surface areas
of ≈ 10m2 , with and without plaster, and with mass per unit areas in the range 70 to 430 kg/m2 . Measured data from Hopkins
are reproduced with permission from ODPM and BRE.
399
S o u n d
I n s u l a t i o n
3.12.3.2.3 Excitation and accelerometer positions
Accelerometers need to be used at a number of random positions over the surface of each
element to account for the spatial variation in vibration. However, at these positions, the direct
field of the excitation and any nearfields close to the boundaries must be negligible (Sections
2.7.5 and 2.7.2). Suitable distances from the excitation point or from the boundaries will vary for
each element as well as with frequency. In practice it is awkward to prescribe different distances
for different elements at different frequencies. It is therefore necessary to compromise and
choose distances based upon typical elements as well as the practicalities of the measurement
(Pedersen, 1993). In addition, the excitation positions must be chosen to ensure excitation of
the different modes and so that the direct field is negligible at the junction or other connection
under test. The values quoted in ISO 10848 Part 1 are representative of distances in common
use. The distance between the accelerometer and the excitation point is at least 1 m, the
distance between the accelerometer and the boundaries is at least 0.25 m, and the distance
between accelerometer positions is at least 0.5 m. The distance between the excitation position
and the junction is at least 1 m, and is at least 0.5 m from the other boundaries.
As with spatial sampling of sound pressure in rooms, the inclusion of correlated samples on a
plate should also be avoided. As a rule-of-thumb for a two-dimensional diffuse bending wave
field it can be assumed that accelerometer positions should be separated from each other by a
distance, d ≥ λB /2. However a wide variety of plate materials and thicknesses are encountered
in practice, the majority of which do not have diffuse bending wave fields over the entire building
acoustics frequency range. A minimum distance between accelerometer positions of 0.5 m is
a pragmatic choice because the direct field near the excitation point limits the available area
in which accelerometers can be fixed.
3.12.3.3
Coupling Loss Factor, ηij
Two methods are particularly convenient to determine the coupling loss factor: the first uses
the velocity level difference and the total loss factor (Craik, 1982b), the second uses structural
intensity (Craik, 1995). In this section it is more appropriate to refer to SEA subsystems rather
than building elements.
The method using the velocity level difference is the most basic form of experimental SEA. We
restrict our attention to two connected subsystems, i and j, for which we want to determine the
coupling loss factor, ηij . A thorough introduction to the two-subsystem SEA model shown in
Fig. 3.109 is given in Section 4.2. This model is based on a blinkered view of the building
because these two subsystems will rarely be isolated from other subsystems. As an example,
Wij
Win(i)
Subsystem
i
Subsystem
j
Wji
Wd(i)
Figure 3.109
Two-subsystem SEA model.
400
Wd( j)
Chapter 3
consider a floor connected to a wall; they will inevitably be connected to other walls or columns
for structural support. However, we can account for these other subsystems by considering the
coupling losses from our two subsystems of interest to these other subsystems. Effectively,
these other subsystems are treated purely as places of energy dissipation. This is reasonable
when these other subsystems do not form flanking transmission paths between the two subsystems of interest. This can often be arranged in a laboratory, but field measurements will
usually be affected by other flanking paths.
From Fig. 3.109, a power input, Win(i) , is applied to subsystem i from which some power
is dissipated through internal losses and coupling losses to other subsystems, and some is
transmitted to subsystem j. For subsystem i the dissipated and transmitted powers are Wd(i)
and Wij respectively. The transmitted power from subsystem i to subsystem j is
Wij = ωηij Ei
(3.244)
This transmitted power can also be calculated from the perspective of subsystem j, because
the power transmitted from i to j must be the same as the total power dissipated by subsystem j.
This is described using the total loss factor of subsystem j,
Wij = ωηj Ej
(3.245)
Assuming that there is negligible power flow back from subsystem j to i, the coupling loss factor
is calculated by equating Eqs 3.244 and 3.245 to give
ηij =
Ej
ηj
Ei
(3.246)
Hence the coupling loss factor in decibels can be calculated from the velocity level difference,
Dv,ij , and the total loss factor of subsystem j using
mi
Lηij = Lηj − Dv,ij + 10 lg
(3.247)
mj
where mi and mj are the mass of subsystems i and j respectively.
An alternative method of determining the coupling loss factor is to use structural intensity to
quantify the net power, Wnet,ij transmitted from i to j. From the two-subsystem SEA model,
Wnet,ij = Wij − Wji = ωηij Ei − ωηji Ej
(3.248)
hence the coupling loss factor is
ηij =
Wnet,ij + ωηji Ej
ωEi
(3.249)
In practice, the power returning from subsystem j to i is usually negligible, hence
ηij ≈
Wnet,ij
ωEi
and the coupling loss factor in decibels can be calculated using
Wnet,ij
Lηij = 10 lg
− [Lv,i + 10 lg (ωmi )]
10−12
(3.250)
(3.251)
401
S o u n d
I n s u l a t i o n
An advantage with structural intensity is that it is possible to make use of the probe directivity to distinguish between different transmission paths, thus avoiding problems with flanking
transmission via other subsystems. As an example, it can be used to quantify the coupling loss
factor between two leaves of a cavity wall via the foundations when there is another connection
between the leaves, such as a lintel above a window.
Measured coupling loss factors between plates are not only specific to the type of junction,
but also the properties of both plates, including the area of the source plate and the junction
length. This can be seen by referring back to the coupling loss factor definition in Eq. 2.154.
Hence care needs to be taken in the comparison of measured coupling loss factors from
different laboratories/buildings and when using them in SEA models of other constructions.
Measurements on beams are restricted in a similar way.
3.12.3.4 Vibration Reduction Index, Kij
The vibration reduction index, Kij , is defined as (EN 12354 Part 1)
Lij
Kij = Dv,ij + 10 lg √
ai aj
(3.252)
where Lij is the junction length between elements i and j, a is the absorption length calculated
from the measured structural reverberation time (Eq. 2.148) and the direction-averaged velocity
level difference, Dv,ij , is
Dv,ij + Dv,ji
(3.253)
2
The reasons behind this definition of the vibration reduction index and the reason for using
a direction-averaged velocity level difference are discussed in Section 5.4 along with the
SEA-based model in which they are used.
Dv,ij =
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408
Chapter 4
Direct sound transmission
4.1 Introduction
D
irect sound transmission across a single building element occurs where the element
is excited by an airborne or a structure-borne sound source on one side, and radiates
sound from the other side without any flanking transmission.
This chapter looks at predicting and interpreting features that describe the airborne and impact
sound insulation of various building elements. Splitting up a building into its component parts
(e.g. floating floor, concrete base floor, ceiling finish, flanking walls, wall linings) is a convenient
way to approach calculation of the sound insulation; it also suits the way that designers and
manufacturers approach the design of a building. In many cases this approach is well-suited
to prediction using Statistical Energy Analysis (SEA); hence this is discussed at the beginning
of the chapter. For some types of building elements this approach allows insight into sound
transmission mechanisms from which decisions can be made on ways to improve the sound
insulation. However, it is not suited to all types. The inability of prediction models to deal with
every single type of building element indicates why laboratory measurements are so important
in providing information at the design stage. At the same time we find that prediction models
illustrate the inherent limitations of many laboratory measurements; in some cases the most
useful information lies somewhere between the two.
4.2 Statistical energy analysis
SEA is a framework of analysis for predicting the transmission of sound and vibration in builtup structures by using a statistical approach with energy as the primary variable. SEA was
introduced in the 1960s and is a well-established engineering tool used in construction, ship,
automobile, and aerospace industries (Lyon and DeJong, 1995). It is introduced in this chapter
because if the concepts are grasped in the context of direct sound transmission, it is easier
to apply them to flanking transmission in the next chapter. It is important to note that SEA
provides a framework for the analysis of complex systems; the classical theories of sound and
vibration transmission can be, and usually are, incorporated within this framework.
The origins of SEA reside in a linear system comprised of two ‘weakly’ coupled oscillators
excited by independent broadband random noise (Lyon and Maidanik, 1962; Scharton and
Lyon, 1968). Considering the temporal average energy for each oscillator, E1 and E2 , the net
energy flow between oscillators is proportional to the difference in the uncoupled energies of
the oscillators. Net power transfer takes place from the oscillator with higher energy to the
oscillator with lower energy, and can be expressed using K1 and K2 as (temporarily undefined)
coupling terms:
Wnet,12 = K1 (E1 − K2 E2 )
(4.1)
409
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I n s u l a t i o n
This approach is then extended from two oscillators, to two sets of oscillators that are coupled
together. It is now appropriate to refer to the oscillators as modes and to introduce the term
‘subsystem’ to represent one set of oscillators. A subsystem comprises a similar group of modes
each with similar modal energy for the form of excitation, where the subsystem response is
determined in frequency bands. Between the modes of a subsystem it is assumed that there
is no correlation and no significant transfer of energy. For this to hold true, the subsystems
must be ‘weakly’ coupled with power input from statistically independent excitation forces. With
this more general approach for subsystems 1 and 2, we have K1 = ωη12 (which introduces a
coupling loss factor, η12 ) and K2 = n1 /n2 (which introduces a ratio of modal densities). This
gives the following relationship between two coupled subsystems; commonly referred to as
the consistency relationship,
η12
η21
=
n2
n1
(4.2)
Power flowing from 1 to 2 can now be written as
W12 = ωη12 E1
(4.3)
where ω is the angular frequency.
From Eqs 4.1–4.3 the net power flow can now be described using only coupling loss factors
as
Wnet,12 = W12 − W21 = ωη12 E1 − ωη21 E2
(4.4)
For a frequency band with a bandwidth, ω, and a centre frequency, ω, the modal energy of
a subsystem is defined as the total subsystem energy divided by the mode count, N, in that
band. Equation 4.4 can therefore be written in terms of the modal energies as
E2
E2
E1
E1
−
−
= ωη12 N1
(4.5)
Wnet,12 = ωη12 n1 (ω)ω
n1 (ω)ω
n2 (ω)ω
N1
N2
This shows that the net power transferred between two coupled subsystems is proportional to
the difference in their modal energies. The thermal analogy is a useful way of describing this
process (Lyon and DeJong, 1995). For each subsystem, the modal energy can be considered
as acoustic temperature, such that there will be heat (energy) flow from subsystems of high
temperature (high modal energy) to those of low temperature (low modal energy). This provides the basis upon which systems with more than two subsystems can be studied when the
following assumptions are satisfied (Hodges and Woodhouse, 1986; Lyon and DeJong, 1995):
1. Statistically independent excitation forces.
2. Equal probability of modes occurring in a certain frequency range.
3. Equipartition of modal energy in a subsystem, and incoherent modal response
between modes in the coupled subsystems.
4. ‘Weak’ (or ‘light’) coupling between subsystems.
These four points are discussed below.
The requirement for statistically independent excitation forces is due to the SEA assumption
that when individual subsystems are coupled together, the modal vibrations must be uncorrelated so that a linear relationship can be used between net power transfer and modal energies
410
Chapter 4
using local modes. If the excitation causes the modal response to be coherent then the requirement for equipartition of modal energy will not be satisfied. Statistically independent excitation
forces can be realized using rain-on-the-roof excitation (also called delta-correlated excitation); this is defined as unity magnitude, random phase, multi-point excitation over the entire
subsystem. In measurements it is common to use point excitation for reasons of practicality.
When point excitation is applied at the anti-node of two or more modes that are sufficiently
close in frequency, correlation exists between the modal responses, thus violating one of the
assumptions in SEA (Fahy, 1970). However, physical experiments on structures indicate that
using point excitation and averaging the response from a number of randomly chosen excitation points can be used to approximate statistical independence between the modes (Bies
and Hamid, 1980). Hence sound and vibration measurements using point sources are almost
always averaged from a number of different source positions.
For real spaces and structures, mode frequencies and mode shapes are rarely the precise entities that analytic solutions and deterministic models (such as finite element methods) would
lead us to believe. There will be uncertainty in describing the modes of real spaces or structures due to uncertainty in the dimensions, material properties, and particularly for building
structures, the quality of workmanship. Hence it is reasonable to assume that there is equal
probability of the mode frequencies falling within a certain frequency range. Embracing the
issue of uncertainty when describing modes is a liberating step as it allows use of the statistical
modal density. We no longer have to be concerned by the fact that for a given room, wall, or
floor, there are large numbers of modes to deal with, for which we are unable to accurately
quantify the mode frequency or mode shape. It is often assumed that uncertainty only occurs
at ‘high frequencies’. However, from the viewpoint of a structure-borne sound wave, buildings
tend to be relatively imprecise, highly variable structures and we can often assume that there
is uncertainty in the modal description over most of the building acoustics frequency range.
Equipartition of modal energy in a subsystem means that every mode has equal energy, which,
as we rarely know otherwise, is often assumed; partly because we have already assumed
broadband statistically independent excitation forces. This form of broadband excitation also
tends to satisfy the requirement for incoherent modal response between modes in the coupled
subsystems.
For predictive SEA, ‘weak’ coupling can (to some extent) be considered as occurring when
the local mode behaviour of an uncoupled subsystem is hardly changed when it is coupled
to the other subsystems such that energy flow can be related to the local modal energies.
However, ‘weak’ coupling between subsystems is an awkward criterion about which there has
been much debate. A review of the literature by James and Fahy (1994) suggests that there is
confusion between the validation of the fundamental SEA equation (Eq. 4.1), the use of SEA
with wave theory calculation of the coupling loss factor, and the requirements necessary to use
experimental SEA. This has led to different definitions of ‘weak’ coupling depending upon the
model under consideration. James and Fahy (1994) concluded that ‘weak’ coupling definitions
that were created to assess the validity of Eq. 4.1 are of little or no use.
4.2.1 Subsystem definition
We can now start to define SEA subsystems that form parts of a building. These are either
space subsystems (e.g. rooms, cavities) or structural subsystems (e.g. plates, beams).
411
S o u n d
I n s u l a t i o n
Subsystems are defined by their ability to store modal energy. Therefore, the boundaries of
a subsystem must cause reflections so that the sound or vibration field is reverberant for the
specific wave type considered in the subsystem. Reflections occur when there is an impedance
change at a boundary. Hence for space subsystems where there is only one wave type, the
surfaces that define a room or cavity usually define the subsystem. For structural subsystems
it can be slightly more complex. Although bending waves are of primary importance for sound
radiation, in-plane waves can be important for structure-borne sound transmission. As these
waves have different modal energies, they need to be represented as separate subsystems.
For example, a plate can be represented by three subsystems using a separate subsystem
for bending, transverse shear, and quasi-longitudinal waves. Conversion between these wave
types at a junction can therefore be included in an SEA model using coupling loss factors from
one subsystem to another. The subsystem boundaries may vary depending upon the wave
type under consideration. This is relevant to structural subsystems where junction impedances
are very different for bending waves compared to in-plane waves (Fahy, 1974). Fortunately for
rigidly connected masonry/concrete plates, the visible boundaries of a wall or floor that face into
a room usually give a reasonable demarcation of a subsystem for either bending, transverse
shear, or quasi-longitudinal waves (Craik, 1998). A plate or beam can be represented by one
subsystem for each wave type, although there are many situations where it is only necessary
to consider bending waves and a single subsystem will be sufficient. This is generally the case
for direct transmission described in this chapter. In the next chapter we will consider separate
subsystems for each of the three wave types on a plate.
In Section 2.7.7 examples were given of non-reverberant plates that had a significant decrease
in vibration with distance. This does not automatically mean that SEA cannot be used; it may
still be suitable when the excitation is distributed over the plate surface (i.e. airborne excitation)
although it does indicate that other models should be considered. If this decrease is only due
to high internal damping for bending wave motion, a rule of thumb for the maximum dimension
of a plate subsystem, Lmax , is (Lyon and DeJong, 1995)
Lmax <
cg,B
ωηint
(4.6)
When there is a significant decrease in vibration with distance, use of SEA will depend on the
type of excitation and the most important transmission mechanisms. Structural excitation at a
single point or along a line on a plate may indicate a significant decrease in vibration across
plates that are highly damped, non-homogeneous or spatially periodic. For airborne or multipoint structural excitation (i.e. rain-on-the-roof) over the plate surface it may still be possible to
include the plate as a single subsystem in an SEA model. Alternatively it may be necessary to
model a single plate as more than one subsystem to account for these losses over distance.
4.2.2 Subsystem response
SEA gives the temporal and spatial average response of a subsystem in terms of its energy.
By assuming a statistical description for each subsystem, the subsystem response represents
the ensemble average of ‘similar’ subsystems with physical parameters drawn from statistical
distributions. Hence it is not necessary to know the exact geometry of a room when representing
it as a space subsystem, just the volume. This approach makes SEA an attractive form of
analysis from an engineering viewpoint. It also simplifies interpretation of the results because
the predicted subsystem energies usually have a smooth variation with frequency. However,
412
Chapter 4
this has implications for the comparison of SEA predictions against measured data. To validate
SEA models we ideally need to compare the prediction against the average value calculated
from measurements on a number of ‘similar’ constructions because SEA does not predict the
response of an individual system with specific modal features.
Ideally we would like to calculate the variance as well as the average response. Unfortunately,
calculation of the variance for SEA systems is only possible to a fairly limited extent (Lyon
and DeJong, 1995) although research continues to address this issue (e.g. see Langley and
Cotoni, 2004). For airborne sound insulation the variance is mainly determined by the variation
in the coupling loss factors and the spatial variation of the sound pressure level. For impact
sound insulation, the power input also varies due to variation in the driving-point impedance
over the plate surface. Whilst we cannot predict the variance, there are some ideal conditions
that give reasonable agreement between an SEA prediction and the measured result from
a single construction. To minimize spatial variance of the response and to ensure that resonant transmission occurs under damping control of the modes we require ‘high’ mode counts
and ‘high’ modal overlap in the frequency band of interest. Quantifying suitably ‘high’ values
depends on the system under analysis.
For plates and/or beams that are coupled together at a junction, a wave approach can be used
to calculate the transmission coefficients; these are needed for calculation of the coupling loss
factors (Section 5.2). Assuming that the junction has been correctly modelled, it is useful to
know the requirements on the subsystems such that these apply to the ensemble with a low
variance. Computational and physical experiments indicate that the wave transmission coefficients only give accurate estimates when the geometric mean of the modal overlap factors,
Mav , is at least equal to unity (Fahy and Mohammed, 1992; also see Clarkson and Ranky, 1984;
Davies and Wahab, 1981). The geometric mean of the modal overlap factors for subsystems
i and j is given by
(4.7)
Mav = Mi Mj
For coupled plates, an additional condition for each plate subsystem is that there should be at
least five modes in the frequency band, Ns ≥ 5 (Fahy and Mohammed, 1992). In the low- and
mid-frequency ranges, many plate and beam subsystems in buildings have zero, fractional,
or low mode counts in one-third-octave-bands. Uncertainty in predicting the mode frequencies
gives rise to uncertainty as to which bands are under damped modal control, and which (if
any) are not. For this reason it is more meaningful to use the statistical mode count. We will
adopt the condition Ns ≥ 5 as a quantitative definition of the term ‘multi-modal’. There are many
plates in buildings (particularly masonry/concrete walls and floors) that are not multi-modal for
bending modes in the low- and mid-frequency ranges. For structural coupling between plates
that is calculated using a wave approach, SEA can still be used if the empirical conditions,
Mav ≥ 1 and Ns ≥ 5 are not met for each plate subsystem, but the levels of uncertainty may
be large. These conditions tend to be overly restrictive in the application of SEA to buildings
where masonry/concrete plates are not multi-modal, but have high total loss factors (Craik et al.,
1991). Section 5.2.3 contains examples and further discussion relating to these conditions for
vibration transmission between plates.
In contrast to many plates in buildings, rooms often form multi-modal subsystems in the lowand mid-frequency range. Therefore coupling between a room and a plate may involve resonant
transmission between a multi-modal space subsystem and a plate subsystem with a fractional
mode count. In addition the space subsystem may have high modal overlap compared to the
413
S o u n d
I n s u l a t i o n
plate subsystem with a value well-below unity. In such situations, reasonable estimates can
still be achieved when Ns ≥ 1 for the plate subsystem and Mav ≥ 1; examples are given in
Section 4.3.1.
4.2.3 General matrix solution
SEA requires knowledge of the dissipative subsystem losses, the coupling losses between
subsystems and the actual or nominal power input into the subsystem(s). The losses are
described using loss factors; these give the fraction of energy transferred per radian cycle.
Three loss factors are defined: internal (dissipative) subsystem losses (ηii ), coupling losses
between subsystems (ηij ), and total subsystem losses (ηi ). These have already been introduced
as the internal loss factor, the coupling loss factor, and the total loss factor in Chapters 1 and 2.
The internal loss factor accounts for energy that is converted to heat; it can also be used to
account for energy that is transferred to parts of the system that are not included in the SEA
model. The coupling loss factor accounts for energy transferred to another subsystem. The
total loss factor of a subsystem is the sum of its internal loss factor and all the coupling loss
factors from that subsystem (Eq. 1.106).
A two-subsystem model with a single power input illustrates the principles of energy flow
between subsystems (see Fig. 4.1). This is the simplest situation of relevance to sound insulation. It could represent power input from a loudspeaker in a source room, with the source and
receiving rooms as subsystems 1 and 2 respectively; the sound could be transmitted between
the rooms via an aperture, ventilator, or porous material. Alternatively it could represent power
input from the ISO tapping machine into a plate (subsystem 1) representing a floor that radiates
sound into the receiving room (subsystem 2). From conservation of energy, the power balance
equations for subsystems 1 and 2 are
Win(1) + ωη21 E2 = ωη11 E1 + ωη12 E1
(4.8)
ωη12 E1 = ωη22 E2 + ωη21 E2
(4.9)
When the power input and the loss factors are known, Eqs 4.8 and 4.9 can be solved to find
the subsystem energies. However, as we almost always deal with more than two subsystems,
Win(1)
W12
Subsystem
1
Subsystem
2
W21
Wd(1)
Wd(2)
Figure 4.1
Two-subsystem SEA model. A power input, Win(1) , is applied to subsystem 1. Transmitted power between subsystems is
denoted by W12 and W21 . Dissipated power (through internal losses and coupling losses to other subsystems) is denoted by
Wd(1) and Wd(2) .
414
Chapter 4
the power balance equations are generalized into a matrix solution for N subsystems to give
the general SEA matrix solution,
⎡
N
η1n −η21
⎢n=1
⎢
N
⎢
⎢ −η12 η2n
⎢
⎢
n=1
⎢
⎢
⎢ −η13 −η23
⎢
⎢
⎢ ..
⎢ .
⎢
⎣
−η1N
⎤
⎤
⎡W
in(1)
−η31 · · · −ηN1 ⎥ ⎡ E1 ⎤
⎥⎢ ⎥ ⎢ ω ⎥
⎥⎢ ⎥ ⎢
⎥
⎥ ⎢ E ⎥ ⎢ Win(2) ⎥
−η32
⎥⎢ 2⎥ ⎢
⎥
⎥⎢ ⎥ ⎢ ω ⎥
⎥
⎢
N
⎥ ⎢ Win(3) ⎥
⎥
⎥⎢
=
E3 ⎥
η3n
⎥
⎥⎢
⎢
⎥ ⎢
⎥
⎢
n=1
⎥ ⎢ ω ⎥
⎥
⎥⎢
⎥
⎢
.
.
⎥⎢ . ⎥ ⎢ . ⎥
..
⎥⎣ . ⎦ ⎢ . ⎥
.
⎥
⎦
⎣
Win(N)
⎦
N
E
N
ηNn
ω
n=1
(4.10)
where ηij is the coupling loss factor from subsystem i to j, and ηii is the internal loss factor for
subsystem i.
For general discussions it is convenient to simplify Eq. 4.10 into the form
*
+
Win
[η]{E} =
ω
where [η] is the square matrix of loss factors, {E} is the column matrix for energy, and
the column matrix for power input terms.
(4.11)
;
Win
ω
<
is
The energy matrix contains the unknown subsystem energies in which we are interested. We
are able to fill the loss factor matrix and the power input matrix with predicted (or measured)
values. It is unlikely that every subsystem will be connected to every other subsystem, so some
of the coupling loss factors in the loss factor matrix will be zero. The power input matrix will
only usually have one row that is non-zero because with sound insulation measurements we
only usually have one power input into a subsystem; e.g. a loudspeaker in the source room,
or the ISO tapping machine on a floor.
The subsystem energies are determined using
{E} = [η−1 ]
*
Win
ω
+
(4.12)
where [η−1 ] is the inverse of the loss factor matrix.
In some situations there will be more than one coupling loss factor between two subsystems;
the individual coupling loss factors can then be added together to give a single value for use
in the matrix solution. An example of this could occur with two rooms separated by a solid
plate containing an aperture and a small area of porous material. In this case there will be
three non-resonant coupling loss factors between the rooms; one for non-resonant (mass
law) transmission across the plate, one for transmission through the aperture, and one for
transmission through the porous material.
4.2.4 Converting energy to sound pressures and velocities
Solving the general SEA matrix gives subsystem energies, which can then be converted into
more practical variables such as sound pressure or velocity.
415
S o u n d
I n s u l a t i o n
For space subsystems, energy is converted to a mean-square pressure using Eq. 1.154. In
Section 1.2.8.1 we discussed use of the Waterhouse correction to account for higher energy
density near the boundaries of reverberant rooms. Unlike real spaces, there is no spatial
variation of energy within an SEA space subsystem; it simply represents the spatial average
reverberant energy. To compare measured and predicted sound pressure levels in a room,
the Waterhouse correction can either be accounted for in the measured level or in the SEA
predicted level. In practice, the correction is rarely used because it is only significant in the
low-frequency range where other uncertainties in the SEA model are usually much larger.
Typically, we are interested in the sound pressure level difference, D, between rooms which
can be calculated from the energy ratios between space subsystems, i and j, using
D = 10 lg
pi2
pj2
= 10 lg
Ei
Ej
+ 10 lg
Vj
Vi
(4.13)
Equation 4.13 allows subsequent calculation of all the descriptors commonly used to describe
airborne sound insulation such as R, Dn , DnT .
For structural subsystems, energy is converted to a mean-square velocity using Eq. 2.237.
Energy ratios between structural subsystems, i and j, can be converted to a velocity level
difference, Dv,ij , using
Dv,ij = 10 lg
vi2
vj2
= 10 lg
Ei
Ej
+ 10 lg
mj
mi
(4.14)
Note that Dv,ij is usually defined with i as the source subsystem, hence when i is not the source
subsystem this should be stated for clarity.
When validating an SEA model with measurements on space and structural subsystems it is
sometimes necessary to look at the ratio of sound pressure to vibration; in this situation it is
simplest to convert all the measured variables to energy, and to work purely in terms of the
energy level difference.
4.2.5
Path analysis
Whilst sound pressure or velocity are usually the variables of interest for each subsystem,
they do not give an immediate insight into how sound and vibration is transmitted from one
subsystem to another along the various transmission paths. With SEA it is possible to use
path analysis to assess the relative importance of one transmission path compared to another
(Craik, 1996). The ability to carry out path analysis alongside the matrix solution makes SEA
a powerful tool with which to make design decisions.
We start by considering a simple, ordered SEA system as shown in Fig. 4.2a, focusing on only
the first three subsystems. The power balance equations can be written in such a way that we
exclude the power input term; this allows us to use the energy in the source subsystem as the
reference point in the path analysis,
416
ωη12 E1 = ωη2 E2
(4.15)
ωη23 E2 = ωη3 E3
(4.16)
Chapter 4
(a)
Win(1)
W12
Subsystem
1
W21
W23
Subsystem
2
Wd(1)
...
Subsystem
3
W32
Wd(2)
Subsystem
N
Wd(N)
Wd(3)
(b)
Win(1)
W12
Subsystem
1
Wd(1)
W51
Wd(5)
Subsystem
5
W21
W15
W41
W45
W54
Wd(4)
Subsystem
2
W14
W32
Wd(2)
W23
Subsystem
3
Wd(3)
W43
W34
Subsystem
4
...
Subsystem
N
...
Wd(N)
Figure 4.2
Example SEA systems: (a) Simple, ordered system and (b) more realistic/complex system.
Equations 4.15 and 4.16 can now be combined to find the energy ratio, E1 /E3 , due to energy
flowing along the transmission path, 1 → 2 → 3,
E1
η2 η3
=
E3
η12 η23
(4.17)
Hence for any system with power injected into subsystem 1, the energy ratio between
subsystem 1 and subsystem N for transmission along the chain of subsystems,
1 → 2 → 3 → · · · → N, is
E1
η2 η3 . . . ηN
=
EN
η12 η23 . . . η(N−1)N
(4.18)
Note that Eqs 4.13 and 4.14 can be used to convert the energy ratio for each path to sound
pressure or velocity ratios.
The path with the lowest energy ratio is the strongest path, and the path with the highest energy
ratio is the weakest path. It is simplest to convert the energy ratio into an energy level difference
in decibels by taking 10 times the logarithm (base 10).
417
S o u n d
I n s u l a t i o n
Equation 4.18 may give the impression that path analysis only applies to a very simple ordered
SEA system (Fig. 4.2a), but it equally applies to real systems with more complex connections
between subsystems. An example is given in Fig. 4.2b where there are many transmission
paths of potential interest. For example, to assess the strength of different transmission paths
between subsystems 1 and 4 we could start by comparing E1 /E4 for the following three paths:
(1) 1 → 4, (2) 1 → 2 → 3 → 4, and (3) 1 → 5 → 4. It is also useful to combine different paths to
give E1 /E4 for a specific combination of paths. As we usually look at energy level differences in
decibels, the energy level difference due to transmission between subsystem 1 and subsystem
N along P different paths can be calculated from
⎛
⎞
−1
P
E1
E
1
⎠
10 lg
= −10 lg ⎝
(4.19)
EN Due to
EN p
P paths
p=1
There are other permutations that we have not yet considered; although paths cannot re-enter
the subsystem that contains the source, they can revisit other subsystems. Using the system in
Fig. 4.2b, two examples for E1 /E4 would be 1 → 5 → 4 → 5 → 4 and 1 → 2 → 3 → 2 → 3 → 4.
In practice, these paths are often insignificant compared to the paths that visit each subsystem
only once.
In a complete building there are many transmission paths that determine the overall sound
insulation and so the matrix solution is used to determine the distribution of energy between
the subsystems. However, when we want to find ways of increasing the sound insulation it
is useful to know whether there is a dominant transmission path. A dominant path is defined
here as giving nominally the same energy level difference as the matrix solution; therefore
the combination of all the other paths is relatively unimportant. We can then test out various
changes to the dominant path that might lead to an increase in the sound insulation. Changes
to one path will change the relative importance of the other paths; therefore the overall sound
insulation then needs to be checked by re-calculating the sound insulation using the matrix
solution.
4.3 Airborne sound insulation
This section starts by looking at the airborne sound insulation of a solid homogeneous isotropic
plate as a basis from which to look at the wide range of building elements that are encountered
in practice.
4.3.1 Solid homogeneous isotropic plates
A solid homogeneous isotropic plate represents the simplest type of wall or floor that is found
in a building. Many wall and floor constructions are far more complex than this idealized form;
yet an understanding of sound transmission via this simple plate is of fundamental importance
as it is often used as a benchmark for comparison with more complex constructions. The
main features of sound transmission can also be explained by considering a plate of finite
thickness but infinite size. In practice, many problems in sound insulation design, prediction,
and measurement revolve around the finite size of plates and their connections to other plates.
By starting with an SEA model for a solid, homogeneous, finite size plate, it is easier to grasp
the various concepts, and to extend the calculations to more complicated constructions that
418
Chapter 4
W13
Non-resonant transmission
W31
Win(1)
Subsystem
1
W12
W21
Subsystem
2
Wd(1)
Wd(2)
Source room
(1)
Plate
(2)
W23
W32
Subsystem
3
Wd(3)
Receiving room
(3)
Figure 4.3
Three-subsystem SEA model for airborne sound transmission between a source room (subsystem 1) and a receiving room
(subsystem 3) via a solid homogeneous plate (subsystem 2). Arrows with dashed lines represent non-resonant transmission,
and arrows with solid lines represent resonant transmission.
are found in practice. However, we will make use of the infinite plate model to gain insights
into angle-dependent sound transmission and to note the important links between finite and
infinite plate models for airborne sound transmission.
For airborne sound transmission from a source room to a receiving room across a solid homogeneous plate, a three-subsystem SEA model is required as shown in Fig. 4.3. The plate
subsystem only needs to represent bending waves because radiation from in-plane modes
is insignificant. Note that by limiting the model to three subsystems we are ignoring flanking
transmission and assuming that vibration can be transmitted from the plate to any connected
structure, but not vice versa.
SEA is based on energy flow between groups of modes at resonance; this is referred to as
resonant transmission. For airborne sound transmission across a plate it is also necessary
to include non-resonant transmission between the space subsystems. In an SEA model this
form of transmission bypasses the plate subsystem (see Fig. 4.3) even though the transmitted
power is determined by the plate properties (Crocker and Price, 1969). We will now look at
resonant and non-resonant transmission separately before combining them to calculate the
overall sound insulation.
In deriving the SEA equations we will often refer to coupling between subsystems i and j rather
than using the specific subsystem numbers (1, 2, and 3). As there are only three subsystems there is little scope for confusion and the generalization will make it easier for the reader
to apply the equations to other SEA systems.
4.3.1.1 Resonant transmission
Resonant transmission concerns coupling between modes with resonance frequencies that fall
within the frequency band of interest. In the three-subsystem model considered here, it occurs
419
S o u n d
I n s u l a t i o n
between room modes and plate bending modes. The SEA framework conveniently allows this
to be described in a few short steps.
The sound power, Wij , radiated from a plate subsystem, i, that is undergoing bending wave
motion, into a room subsystem, j, is
Wij = ωηij Ei
(4.20)
Substituting Eqs 2.237 and 2.198 into Eq. 4.20 gives the coupling loss factor from a plate to a
room in terms of the frequency-average radiation efficiency (Section 2.9.4) as
ηij =
ρ0 c0 σ
ωρs
(4.21)
The consistency relationship (Eq. 4.2) can now be used to calculate the coupling loss factor
from a room to a plate, ηji .
We can use SEA path analysis to calculate a resonant sound reduction index, RR , for the resonant transmission path 1 → 2 → 3. This is only the resonant component of the sound reduction
index that we determine using standard sound insulation measurements in a transmission suite.
Assuming that ρ0 and c0 are the same in both rooms, Eqs 4.17 and 4.13 give the resonant
sound reduction index as
η2 η3 V3
S
RR = 10 lg
(4.22)
+ 10 lg
η12 η23 V1
A
Using the consistency relationship with Eqs 1.59 and 2.139 for the room and plate modal
densities respectively, Eq. 4.22 can be rewritten in terms of the plate properties as
2π2 hcL ρs2 f 3 η
RR = 10 lg
(4.23)
√ 2 4
3ρ0 c0 σ 2
where η is the total loss factor of the plate. Note that RR only increases at 9 dB/octave if η and σ
are independent of frequency.
As an example we consider a plate of 6 mm glass. We will assume that the glass is installed
in a filler wall in a transmission suite and that the edges of the glass pane are embedded
in putty held within a wooden frame. Resonant transmission is under damping control of the
plate so we need to determine the total loss factor of the plate from its individual loss factors.
The internal loss factor for glass is usually between 0.003 and 0.006; this is a property of the
glass and does not include any dissipative loss mechanisms related to the putty. The coupling
losses are due to sound radiation and structural coupling. The coupling loss factor for sound
radiation is calculated using Eq. 4.21; note that because the plate radiates into both the source
room and the receiving room, the sum of the coupling losses due to radiation will be twice this
value. We assume that the plate boundaries can be considered as being simply supported
and that the filler wall represents an infinite rigid baffle; the radiation efficiency under these
assumptions is calculated using method no.1 (refer back to 6 mm glass in Fig. 2.65a). At this
point we hit a complication because the structural coupling losses for this particular junction
are difficult to calculate accurately. This is partly because the properties of the putty change
as it dries out, and partly because the junction detail is not simple to model even when the
properties of the putty are fixed. Hence, we accept that whilst we can estimate some parts of
420
Chapter 4
50
SEA: Resonant transmission (thin plate)
Sound reduction index, RR (dB)
45
40
35
30
25
f11 ⫽ 15 Hz
20
fc ⫽ 2079 Hz
15
fB(thin) ⬎ 20 kHz
10
5
50
80
125
200
315
500
800
1250 2000
One-third-octave-band centre frequency (Hz)
3150
5000
Figure 4.4
Resonant transmission across a plate of 6 mm glass. Plate properties: Lx = 1.5 m, Ly = 1.25 m, h = 0.006 m, ρs = 15 kg/m2 ,
cL = 5200 m/s, ν = 0.24, total loss factor η = 0.024.
the total loss factor, we cannot estimate all of them. In such cases the total loss factor can be
calculated from measurements of the structural reverberation time. Measured total loss factors
are usually frequency-dependent, but to simplify matters for this example it is reasonable to
use a single frequency-average value.
The sound reduction index for resonant transmission is shown in Fig. 4.4. We can only consider
resonant transmission when the fundamental mode is below the frequency range of interest.
The thin plate limit is well-above 20 kHz, which conveniently means that thin plate theory can
be used across the entire building acoustics frequency range. As the total loss factor depends
upon the specific installation (and is usually frequency-dependent) it is only appropriate here
to make a brief qualitative assessment of the general trends. Starting at the lowest frequency
band, the sound insulation increases with frequency, then begins to level off as it approaches
the critical frequency. A minimum value is reached very close to the critical frequency, after
which the sound insulation begins to increase with frequency once more.
4.3.1.2 Non-resonant transmission (mass law)
Non-resonant transmission between two spaces separated by a plate is quantified using a
non-resonant transmission coefficient, τNR . This gives a non-resonant sound reduction index,
RNR , for the non-resonant transmission path 1 → 3. We will shortly look at quantifying the
transmission coefficient using infinite and finite plate models, but the first step is to determine
the coupling loss factor.
In a three-dimensional space with a diffuse sound field, the sound power incident upon a plate
of surface area, S, is calculated from Eq. 3.34, giving
W = SI = E
c0 S
Ec0 S
=
dmfp ST
4V
(4.24)
421
S o u n d
I n s u l a t i o n
Reflected wave
Transmitted wave
u
Incident wave
x
z
Figure 4.5
Plane wave incident upon an infinite plate along with the reflected and transmitted waves.
Therefore the sound power, Wij , transmitted between two space subsystems i and j, across a
plate subsystem with a non-resonant transmission coefficient, τNR , is
Wij =
Ei c0 S
τNR = ωηij Ei
4Vi
(4.25)
c0 S
τNR
4ωVi
(4.26)
which gives the coupling loss factor,
ηij =
The consistency relationship (Eq. 4.2) can be used to calculate the coupling loss factor, ηji , in
the reverse direction.
4.3.1.2.1
Infinite plate theory
For an infinite plate, non-resonant transmission is based on the assumption that it simply acts
as a limp mass with no stiffness. For this reason it is also referred to as mass law transmission,
or forced transmission. It is assumed that the infinite plate lies in the xy plane and that a plane
wave is incident upon the plate at an angle, θ, as shown in Fig. 4.5. For this infinite plate we
only need to focus on the x- and z-dimensions, so the incident sound wave can be described by
pi (x, z) = p̂i exp(−ikx sin θ) exp(ikz cos θ)
(4.27)
where the time dependence, exp(iωt), has been excluded for brevity.
Therefore the reflected wave is
pr (x, z) = p̂r exp(−ikx sin θ) exp(−ikz cos θ)
422
(4.28)
Chapter 4
and the transmitted wave is
pt (x, z) = p̂t exp(−ikx sin θ) exp(ikz cos θ)
(4.29)
Continuity of sound pressure at z = 0 gives the total pressure acting on the plate as
(4.30)
p̂ = p̂i + p̂r − p̂t
The plate acts as a limp mass, so from Newton’s third law the total pressure equals the inertial
reaction from the plate,
(4.31)
p̂ = iωρs v̂
where ρs is the mass per unit area of the plate.
At z = 0, continuity requires that the lateral plate velocity equals the z-component of the particle
velocity on each side of the plate,
(4.32)
v̂ = ûi + ûr = ût
Relating the particle velocities to the sound pressures (Eq. 1.16) therefore gives
v̂ = (p̂i − p̂r )
cos θ
cos θ
= p̂t
ρ0 c 0
ρ0 c0
(4.33)
From Eqs 4.30, 4.31, and 4.33, the angle-dependent transmission coefficient, τNR,θ , equals the
ratio of the transmitted sound power to the incident sound power,
τNR,θ =
2
p̂t
Wt
= =
Wi
p̂i
1
1+
ωρs cos θ
2ρ0 c0
2
(4.34)
Hence non-resonant transmission is lowest at normal incidence (0◦ ) and increases with increasing angle of incidence. Equation 4.34 can be used to calculate a diffuse field transmission
coefficient, τd , for an infinite plate. When one side of the plate is exposed to a diffuse sound
field, all angles of incidence are equally probable and the plane waves that are incident upon
the plate will have equal intensity, I. We therefore consider the incident sound as radiating
from a hemisphere that encloses, and is centred around a unit area on the plate. The incident
and transmitted intensities can then be determined from the component of the incident plane
wave intensity that is normal to the plate surface, I cos θ, by integrating over the element of the
solid angle, d = sin θ dθ dφ. Hence the total transmitted intensity, It , is determined from
It =
τθ I cos θ d =
0
2π
π/2
τθ I cos θ sin θ dθ dφ
(4.35)
0
The total incident intensity can be found in the same way. By considering a unit area on the plate,
the diffuse field transmission coefficient, τNR,d , is defined as the ratio of the total transmitted
intensity to the total incident intensity. The incident plane wave intensity is a constant, and τθ
does not vary with φ, hence the integration simplifies to give,
π/2
τNR,d = 0
τθ cos θ sin θ dθ
π/2
cos θ sin θ dθ
=
π/2
τθ sin 2θ dθ
(4.36)
0
0
423
S o u n d
I n s u l a t i o n
The sound field in the low-frequency range is not usually diffuse for typical rooms and will not
contain all angles of incidence. In addition, with transmission suite measurements it is common
to mount a plate within a niche (Section 3.5.1.3.3); therefore the plate may be shielded from
some angles of incidence near 90◦ . An empirical adjustment is often quoted which changes the
upper integration limit in the numerator and denominator of Eq. 4.36 from 90◦ to 78◦ (Vér and
Holmer, 1988). Other empirical values that are sometimes quoted are 75◦ and 80◦ . Here we
will also use 78◦ to define field incidence, and the field incidence transmission coefficient, τNR,f .
Whilst the transmission coefficient is convenient for SEA calculations, it is simpler to compare
the transmission loss in decibels for different types of non-resonant transmission using
1
RNR = 10 lg
(4.37)
τNR
Therefore the transmission loss for non-resonant transmission is
ωρs cos θ 2
RNR,θ◦ = 10 lg 1 +
2ρ0 c0
(4.38)
for a single angle of incidence, which gives
RNR,0◦ = 10 lg 1 +
ωρs
2ρ0 c0
2
(4.39)
for normal incidence, and
RNR,d = RNR,0◦ − 10 lg (0.23RNR,0◦ )
(4.40)
for diffuse incidence (Vér and Holmer, 1988) assuming RNR,0◦ > 15 dB, otherwise use numerical
integration of Eq. 4.36, and
RNR,f = RNR,0◦ − 5 dB
(4.41)
for field incidence (Vér and Holmer, 1988) assuming RNR,0◦ > 15 dB, otherwise use numerical
integration of Eq. 4.36.
Figure 4.6 shows non-resonant sound reduction indices for an infinite plate of 6 mm glass.
These vary significantly with the angle of incidence. A single angle of incidence is relevant to
façade sound insulation measurements when a loudspeaker is directed towards the façade. In
practice there will usually be a range of angles of incidence, whether it is from a reverberant
sound field within a room, or an external environmental noise source. For this reason we are
almost always interested in an angular average value, such as for diffuse or field incidence.
As field incidence excludes angles close to grazing which have very low transmission loss, the
field incidence values are higher than for diffuse incidence.
4.3.1.2.2 Finite plate theory
For airborne sound transmission across finite plates, non-resonant transmission describes
transmission due to bending modes that have their resonance frequencies outside the frequency band of interest (Sewell, 1970). Looking back at Fig. 2.63 we find that individual modes
with resonance frequencies below the critical frequency will have a higher radiation efficiency
at frequencies above their resonance frequency, than actually at their resonance frequency. So
when the frequency band of interest is below the critical frequency, radiation from modes with
424
Chapter 4
55
Infinite plate: θ°
Sound reduction index, RNR (dB)
50
45
Infinite plate: Field incidence
40
Infinite plate: Diffuse incidence
35
0°
45°
30
60°
25
78°
20
15
10
5
89°
fc ⫽ 2079 Hz
0
50
80
125
200
315
500
800
1250
2000
3150
5000
One-third-octave-band centre frequency (Hz)
Figure 4.6
Non-resonant transmission across an infinite plate of 6 mm glass (ρs = 15 kg/m2 , cL = 5200 m/s).
resonance frequencies within the band can be lower than radiation from modes with resonance
frequencies outside the band that have been excited ‘off-resonance’. For the former modes,
the modal response is under damping control; for the latter modes, it is predominantly under
mass control. Non-resonant transmission can therefore be considered as being unaffected by
the plate damping.
In contrast to infinite plate theory, finite plate theory considers the plate to have both mass
and stiffness. Whilst the mass per unit area is still important in quantifying the non-resonant
transmission, the role of bending modes means that it also depends on the plate dimensions
and the critical frequency. For a finite plate, the non-resonant transmission coefficient below
the critical frequency is given by (Leppington et al., 1987)
√
⎫
⎧
ln (k S) + 0.16 − U(Lx /Ly )
⎪
⎪
⎪
⎪
⎤
⎡
⎪
⎪
⎪
2 ⎪
⎬
⎨
(2μ2 − 1)(μ2 + 1)2 ln (μ2 − 1)
2ρ0
⎥
⎢
τNR =
(4.42)
1
⎢+ (2μ2 + 1)(μ2 − 1)2 ln (μ2 + 1)⎥⎪
ρs k(1 − μ−4 ) ⎪
+
⎪
⎪
⎦
⎣
⎪
⎪
6
⎪
⎪
⎭
⎩ 4μ
− 4μ2 − 8μ6 ln μ
where
μ=
fc
f
and U(Lx /Ly ) is a function of the rectangular shape of the plate
&
2 '
Ly
Ly
Lx
Lx
1 Lx
U
+
=U
=
ln 1 +
Ly
Lx
2π Ly
Lx
Ly
2 1 arctan t
Lx
ln 2
Lx
−
dt
ln
−
− 0.5 +
πLy
Ly
π
π Lx /Ly
t
(4.43)
(4.44)
425
S o u n d
I n s u l a t i o n
50
Finite plate: 1.5 ⫻ 1.25 m
45
Sound reduction index, RNR (dB)
Finite plate: 4 ⫻ 2.5 m
40
Infinite plate: Field incidence
35
Infinite plate: Diffuse incidence
30
25
20
15
10
fc ⫽ 2079 Hz
5
50
80
125
200
315
500
800
1250
2000
3150
5000
One-third-octave-band centre frequency (Hz)
Figure 4.7
Non-resonant transmission across finite and infinite plates of 6 mm glass (ρs = 15 kg/m2 , cL = 5200 m/s).
Plates that form windows, doors, walls, and floors usually have aspect ratios in the range,
1/3 ≤ Lx /Ly ≤ 3, and satisfactory estimates for the transmission coefficient can be found by
ignoring the U(Lx /Ly ) term in Eq. 4.42.
Using 6 mm glass as an example we can now compare the non-resonant sound reduction index
for finite plates with infinite plates. This is done using SEA path analysis for the non-resonant
transmission path 1 → 3. For a finite glass plate, a reasonable size to consider is 1.5 × 1.25 m,
but to illustrate the effect of plate size, we will also use rather unrealistic dimensions of 4 × 2.5 m.
The non-resonant sound reduction indices are shown in Fig. 4.7. For the finite plates, the
smaller plate has higher values than the larger plate, and both finite plates have significantly
higher values than an infinite plate assuming diffuse incidence. As the frequency approaches
the critical frequency, the non-resonant sound reduction index for the finite plates starts to
decrease. Above the critical frequency, the non-resonant transmission coefficient is undefined
for the finite plates.
The sound reduction index for field incidence increases by 6 dB per doubling of frequency. This
is steeper than the average slope for each of the finite plates although at some frequencies the
finite plate curves have similar values to the field incidence curve. Field incidence assumes
that angles of incidence are restricted to being equally probable between 0◦ and 78◦ . Its name
unfortunately suggests that this assumption is always valid in the field (i.e. in situ); which it is not.
The assumption simply gives fortuitous agreement with measured data for particular plate sizes
with particular critical frequencies, usually with particular mounting conditions in a niche and
usually where non-resonant transmission dominates in the low- and mid-frequency ranges and
the sound field in the rooms is far from being diffuse. Whilst the infinite plate formulae for diffuse
and field incidence are very useful for quick calculations and illustrative purposes, they do not
describe all the features of non-resonant transmission that relate to finite size plates (Leppington et al., 1987). Unless stated, all calculations of non-resonant transmission in SEA models
will use finite plate theory to calculate the non-resonant transmission from this point onwards.
426
Chapter 4
Note that when modelling more complex plates (e.g. plates that are non-homogeneous,
periodic, or profiled) using an infinite plate approach it is necessary to link the models to
measurements that will be made on finite plates in practice. For convenience this is often done
by restricting the range for the angles of incidence in the same way as with field incidence.
Unlike with finite plate theory, non-resonant transmission across infinite plates is defined above
the critical frequency. This implies that if the plate damping is sufficiently high such that resonant
transmission is negligible at frequencies well-above the critical frequency, then non-resonant
transmission will dominate at these high frequencies. It is unusual to find homogeneous plates
in buildings that are this highly damped in the building acoustics frequency range, and using
infinite plate theory for non-resonant transmission above the critical frequency is not usually
appropriate.
4.3.1.3 Examples
We will now compare predictions of the sound reduction index with transmission suite measurements for different solid plates. For these particular laboratory measurements it is reasonable to
assume that the test elements are surrounded by a rigid infinite baffle, and that the plate boundaries are simply supported. Glass and plasterboard plates are used to illustrate non-resonant
transmission and the dip in the sound reduction index at the critical frequency. Masonry walls
are used to show the importance of resonant transmission across the entire building acoustics
frequency range. These have a plaster finish to remove any transmission via air paths through
the blocks and/or the mortar joints. In contrast to the plates of glass and plasterboard, the
masonry walls have low mode counts and low modal overlap. To ensure accurate prediction
of the resonant transmission, the measured total loss factors are used in the SEA models.
By using the measured total loss factor we are accounting for all the coupling losses from the
plate to the laboratory structure, but not including the laboratory structure in the model. For the
masonry walls the measured total loss factor is approximately described by η = 0.01 + 0.3f −0.5 .
For the plasterboard and masonry walls the measurements were made in a transmission suite
with reverberation times between 1 and 2 s, and source and receiving room volumes of 130
and 115 m3 respectively. The modal overlap factors in such rooms (refer back to Fig. 1.24) will
be much larger than for most masonry/concrete plates. It is therefore reasonable to assume
that it is the geometric mean of the modal overlap factors for the plate and the source room (or
receiving room) that is relevant when assessing the SEA prediction of resonant transmission.
Note that the source room is used for these calculations of Mav (Eq. 4.7) because the two
rooms are similar and the small difference is not important here.
SEA path analysis is used to predict the sound reduction index for the resonant path 1 → 2 → 3,
and the non-resonant path 1 → 3. The matrix solution is used to give the overall transmission
due to resonant and non-resonant transmission; however, in these particular examples, combining the resonant and non-resonant transmission paths using Eq. 4.19 gives the same result.
As we are interested in a wide range of plate thicknesses and materials in buildings, the important plate parameters are included on the figures: statistical mode count, modal overlap factor,
fundamental mode frequency, critical frequency, and the thin plate limit (bending waves).
4.3.1.3.1 Glass
A sheet of 6 mm glass has already been used as an example in this chapter (refer back to
Figs 4.4 and 4.7). Figure 4.8a shows the predicted sound reduction index for the resonant and
427
S o u n d
I n s u l a t i o n
(a)
50
f11 ⫽ 15 Hz
45
fc ⫽ 2079 Hz
Sound reduction index (dB)
40
fB(thin) ⬎ 20 kHz
35
30
25
20
SEA: Non-resonant transmission
15
SEA: Resonant transmission (thin plate)
10
SEA: Resonant and non-resonant transmission
5
50
80
125
200
315
500
800
1250
2000
3150
5000
One-third-octave-band centre frequency (Hz)
(b)
Ns = 1.2 1.5 1.9 2.4 3.0 3.8 4.8 6.0 7.5 9.6 12 15 19 24 30 38 48 60 75 96 120
M = 0.1 0.2 0.2 0.2 0.3 0.4 0.5 0.6 0.8 1.0 1.2 1.6 2.0 2.5 3.1 4.0 5.0 6.2 7.9 10 12
50
f11 ⫽ 15 Hz
45
fc ⫽ 2079 Hz
Sound reduction index (dB)
40
fB(thin) ⬎ 20 kHz
35
30
25
20
SEA: Resonant and non-resonant transmission
15
Measured (after Yoshimura)
10
Measured (after Cops and Soubrier)
5
50
80
125
200
315
500
800
1250
2000
3150
5000
One-third-octave-band centre frequency (Hz)
Figure 4.8
Airborne sound insulation of 6 mm glass: (a) predicted non-resonant and resonant transmission using SEA and (b) comparison
of measurements and SEA. Upper x-axis labels show the predicted statistical mode count and modal overlap factor for the
plate in each frequency band. (Note: Mav is not shown here because measurements were taken in two different laboratories
with different volumes and reverberation times.) Measurements according to ISO 140 Part 3 and niche detail according to
ISO 140 Parts 1 and 3. Plate properties: Lx = 1.5 m, Ly = 1.25 m, h = 0.006 m, ρs = 15 kg/m2 , cL = 5200 m/s, ν = 0.24, total
loss factor η = 0.024 (frequency-average value from structural reverberation time measurements on similar test elements).
Measured data are reproduced with permission from Cops and Soubrier (1988) and Yoshimura (2006).
428
Chapter 4
non-resonant paths separately as well as in combination. Below the critical frequency, nonresonant transmission generally dominates over resonant transmission and the overall sound
insulation is mainly determined by non-resonant transmission. At and above the critical
frequency there is only resonant transmission.
In the vicinity of the critical frequency it is notoriously difficult to accurately predict the radiation
efficiency (even when Ns ≥ 30 and M ≥ 3). This can partly be overcome by calculating lower and
upper limits for the radiation efficiency using method no. 2 (Section 2.9.4.2). SEA calculations
are therefore carried out twice; once with the lower limit and once with the upper limit. Measured
data usually lies within the shaded area between the two limits. Comparison of the combination
of resonant and non-resonant transmission with measurements is shown in Fig. 4.8b (Cops
and Soubrier, 1988; Yoshimura, 2006). Close agreement below the critical frequency confirms
that non-resonant transmission for a finite plate gives a better estimate than diffuse or field
incidence for an infinite plate (refer back to Fig. 4.7).
4.3.1.3.2
Plasterboard
A partition formed from 12.5 mm plasterboard without a frame is used to provide an example
of a plate with a high statistical mode count over the entire frequency range. The structural
coupling losses from the plasterboard to the laboratory structure are negligible; hence the total
loss factor equals the sum of the internal loss factor plus the two radiation coupling loss factors
(i.e. both sides).
Figure 4.9a allows an assessment of the resonant transmission path. Below the critical frequency, there is a large difference between the SEA prediction for resonant transmission and
the measured values. This indicates that there is a much more important sound transmission
mechanism than resonant transmission at these frequencies. As we have previously seen
with the 6 mm glass plate, this is non-resonant transmission. At and above the critical frequency where there is only resonant transmission, the measured values lie within the range
predicted for the radiation efficiency using method no. 2 (Section 2.9.4.2). In Fig. 4.9b the agreement between measurements and the three different predictions for non-resonant transmission
below the critical frequency confirms that non-resonant transmission is the dominant mechanism. The SEA prediction uses finite plate theory to determine the non-resonant transmission
coefficient. Infinite plate theory for a diffuse field tends to overestimate the sound transmission
measured in practice; field incidence fortuitously gives a reasonable estimate, but there is
no firm basis on which it can be applied to plates of all sizes. Combining resonant and nonresonant transmission (finite plate theory) gives good agreement between measurements and
the SEA prediction over the entire building acoustics frequency range as shown in Fig. 4.9c.
4.3.1.3.3
Masonry wall (A)
Masonry wall (A) is a 115 mm aircrete wall with a 13 mm plaster finish on one side (see
Fig. 4.10). As ρ and cL for plaster and aircrete are similar this allows the prediction to assume a
128 mm thick solid homogeneous plate. Measurements on the plastered wall are used to give
cL and the total loss factor. In Section 4.3.8.1 we will discuss the effect of a bonded surface
finish such as plaster in more detail.
Below the critical frequency, resonant transmission dominates over non-resonant transmission.
This can be seen by the fact that the curve for resonant transmission is very similar to the curve
for the combination of resonant and non-resonant transmission. Due to the low mode count, the
429
S o u n d
I n s u l a t i o n
60
Sound reduction index (dB)
(a)
Measured
SEA: Resonant transmission
(thin plate)
50
40
30
20
f11 ⫽ 2 Hz
fc ⫽ 3483 Hz
10
fB(thin) ⫽ 5891 Hz
0
50
80
125
200
315
500
800
1250 2000 3150 5000
One-third-octave-band centre frequency (Hz)
60
(b)
Sound reduction index (dB)
Measured
SEA: Non-resonant transmission
50
Infinite plate: Field incidence
40
Infinite plate: Diffuse incidence
30
20
10
0
50
80
125
200
315
500
800
1250 2000 3150 5000
One-third-octave-band centre frequency (Hz)
Ns = 10 13 16 20 25 32 40 50 63 79 99 125 159 199 248 318 397 496 626 794 993
M = 0.6 0.8 1.0 1.2 1.5 2.0 2.5 3.1 3.9 4.9 6.1 7.7 9.8 12 15 20 25 31 49 59 71
Mav = 0.5 0.6 0.8 1.1 1.4 2.0 2.7 3.7 5.2 7.3 10 14 20 28 38 55 77 108 170 238 326
60
Measured
Sound reduction index (dB)
(c)
SEA: Resonant and non-resonant transmission
50
40
30
20
10
0
50
80
125
200
315
500
800
1250 2000 3150 5000
One-third-octave-band centre frequency (Hz)
Figure 4.9
Measured and predicted airborne sound insulation of 12.5 mm plasterboard. Upper x-axis labels show the predicted statistical
mode count and modal overlap factor for the plate and Mav (plate and room). Measurements according to ISO 15186 Part 3
(50–100 Hz) and ISO 15186 Part 1 (125–5000 Hz). Plate properties: Lx = 3.53 m, Ly = 2.63 m, h = 0.0125 m, ρs = 10.8 kg/m2 ,
cL = 1490 m/s, ν = 0.3, ηint = 0.0141. Measured data from Hopkins are reproduced with permission from ODPM and BRE.
430
Chapter 4
Ns ⫽ 0.8 1.0 1.3 1.6 2.0 2.5 3.2 4.0 5.0 6.4 7.9 10 13 16 20 25 32 40 50 64 79
M ⫽ 0.1 0.1 0.1 0.2 0.2 0.3 0.4 0.5 0.6 0.9 1.1 1.3 1.4 2.0 2.3 2.7 2.4 3.4 3.9 4.4 4.5
Mav ⫽ 0.2 0.2 0.3 0.4 0.6 0.7 1.1 1.5 2.1 3.1 4.3 5.6 7.6 11 15 21 24 36 48 65 82
80
Measured
SEA: Non-resonant transmission
Sound reduction index (dB)
70
SEA: Resonant transmission (thin plate)
SEA: Resonant and non-resonant transmission
60
Plateau for thick plates above 4fB(thin)
50
40
f11 ⫽ 24 Hz
30
fB(thin) ⫽ 778 Hz
fc ⫽ 278 Hz
20
50
80
125
200
315
500
800
1250 2000
One-third-octave-band centre frequency (Hz)
3150
5000
Figure 4.10
Measured and predicted airborne sound insulation of a 115 mm masonry wall (solid aircrete blocks) with a 13 mm lightweight
plaster finish (one side). Upper x-axis labels show the predicted statistical mode count and modal overlap factors for the plate
and Mav (plate and room). Measurements according to ISO 15186 Part 3 (50–100 Hz) and ISO 15186 Part 1 (125–5000 Hz).
Plate properties: Lx = 3.53 m, Ly = 2.63 m, h = 0.128 m, ρs = 71 kg/m2 , cL = 1820 m/s, ν = 0.2, measured total loss factor.
The plateau is calculated using material properties corresponding to the plate thickness: cL = 1920 m/s and an internal loss
factor ηint = 0.0125. Measured data from Hopkins are reproduced with permission from ODPM and BRE.
radiation efficiency has been calculated using method no. 4 (Section 2.9.4.4). For 0.2 ≤ Mav < 1
the measured and predicted values show close agreement; however if many similar walls were
measured one would expect to see a wide range of values due to the low mode counts and
low modal overlap. Hence there will often be individual bands where the differences are large.
In this particular example, the largest difference occurs in the 50 Hz band.
At and above the critical frequency where Ns ≥ 4 and Mav ≥ 1 there is close agreement between
measurements and predictions up to a frequency of 4fB(thin) . Above 4fB(thin) the measured values
start to level-off to form a plateau; the prediction of this plateau is discussed in Section 4.3.1.4.
4.3.1.3.4
Masonry wall (B)
Masonry wall (B) is a 100 mm dense aggregate masonry wall with a 13 mm plaster finish on
one side (see Fig. 4.11). Compared to the plaster, the dense aggregate wall has a high mass
per unit area and is much stiffer; hence the plaster can be ignored in calculating the plate
thickness and density for the SEA model.
Below the critical frequency the mode count Ns < 3, so the radiation efficiency has been calculated using method no. 4 (refer back to Section 2.9.4.4 and Fig. 2.65d). If we assume that
non-resonant theory for finite plates still gives a better estimate than the infinite plate theory, we find that resonant transmission dominates over non-resonant transmission. Below the
critical frequency where 0.2 ≤ Mav < 1 the agreement between measurement and prediction
431
S o u n d
I n s u l a t i o n
Ns ⫽ 0.6 0.7 0.9 1.2 1.4 1.8 2.3 2.9 3.6 4.6 5.8 7.3 9.2 12 14 18 23 29 36 46 58
M ⫽ 0.1 0.1 0.1 0.1 0.2 0.2 0.3 0.5 0.6 0.8 0.9 0.9 1.1 1.2 1.4 1.6 1.9 1.9 2.4 3.3 3.9
Mav ⫽ 0.2 0.2 0.2 0.4 0.5 0.7 1.0 1.4 2.0 2.9 3.8 4.8 6.5 8.7 12 16 22 26 38 56 77
80
Measured
SEA: Non-resonant transmission
Sound reduction index (dB)
70
SEA: Resonant transmission (thin plate)
60
SEA: Resonant and non-resonant transmission
50
f11 = 33 Hz
40
fc = 203 Hz
fB(thin) = 1751 Hz
30
20
50
80
125
200
315
500
800
1250
2000
3150
5000
One-third-octave-band centre frequency (Hz)
Figure 4.11
Measured and predicted airborne sound insulation of a 100 mm masonry wall (solid dense aggregate blocks) with a 13 mm
lightweight plaster finish (one side). Upper x-axis labels show the predicted statistical mode count and modal overlap factors
for the plate and Mav (plate and room). Measurements according to ISO 15186 Part 3 (50–100 Hz) and ISO 15186 Part 1
(125–5000 Hz). Plate properties: Lx = 3.53 m, Ly = 2.63 m, h = 0.1 m, ρs = 200 kg/m2 , cL = 3200 m/s, ν = 0.2, measured
total loss factor. Measured data from Hopkins are reproduced with permission from ODPM and BRE.
is generally similar to when Mav ≥ 1. To make a thorough assessment of the uncertainty we
cannot rely on a single measurement, we would need to measure a number of similar walls.
Above the critical frequency where Ns ≥ 1 and Mav ≥ 1 there is close agreement between measurement and prediction. The thin plate limit lies within the high-frequency range; this provides
more evidence that it is reasonable to use thin plate theory up to 4fB(thin) .
4.3.1.3.5
Masonry wall (C)
For masonry wall (C) we look at a thicker version of masonry wall (B). This is a 215 mm dense
aggregate masonry wall with a 13 mm plaster finish on each side (see Fig. 4.12); as before,
the plaster is ignored in calculating the plate thickness and density.
Assuming simply supported boundaries, the fundamental mode frequency is 70 Hz. This is
above the lowest one-third-octave-band considered in the building acoustics frequency range,
and close to the critical frequency. For this reason, only resonant transmission is predicted and
shown above the fundamental mode.
In the low-frequency range where 0.1 ≤ Mav < 1, the agreement between measurement and
prediction is not significantly worse than in the mid- and high-frequency ranges where Ns ≥ 1
and Mav ≥ 1. Although errors are incurred by using thin plate theory between fB(thin) ≤ f < 4fB(thin)
they are not large enough to warrant changing to thick plate theory. Above 4fB(thin) the curve
reaches a plateau due to thickness resonances; these are discussed in the following section.
432
Chapter 4
Ns ⫽ 0.3 0.3 0.4 0.5 0.7 0.9 1.1 1.3 1.7 2.2 2.7 3.4 4.3 5.4 6.7 8.6 11 13 17 22 27
M ⫽ 0.04 0.03 0.03 0.1 0.1 0.1 0.2 0.2 0.3 0.4 0.3 0.3 0.4 0.4 0.5 0.5 0.7 0.9 1.1 1.2 1.4
Mav ⫽ 0.1 0.1 0.1 0.3 0.4 0.5 0.7 1.0 1.3 2.0 2.2 2.6 3.8 5.2 6.6 8.9 13 18 26 34 46
80
Measured
Sound reduction index (dB)
70
SEA: Resonant transmission (thin plate)
Plateau for thick plates above 4fB(thin)
60
50
f11 = 70 Hz
40
fc = 94 Hz
fB(thin) = 814 Hz
30
20
50
80
125
200
315
500
800
1250
2000
3150
5000
One-third-octave-band centre frequency (Hz)
Figure 4.12
Measured and predicted airborne sound insulation of a 215 mm masonry wall (solid dense aggregate blocks) with a 13 mm
lightweight plaster finish (each side). Upper x-axis labels show the predicted statistical mode count and modal overlap factor
for the plate and Mav (plate and room). Measurements according to ISO 15186 Part 3 (50–100 Hz) and ISO 15186 Part 1
(125–5000 Hz). Plate properties: Lx = 3.53 m, Ly = 2.63 m, h = 0.215 m, ρs = 430 kg/m2 , cL = 3200 m/s, ν = 0.2, measured
total loss factor. The plateau is calculated using material properties corresponding to the plate thickness: cL = 4000 m/s and
an internal loss factor ηint = 0.01. Measured data from Hopkins are reproduced with permission from ODPM and BRE.
4.3.1.4 Thin/thick plates and thickness resonances
In the low-frequency range, thick masonry/concrete plates sometimes have their fundamental bending mode above the critical frequency. In such cases it is not suitable to calculate
non-resonant transmission below the critical frequency using finite plate theory. However, resonant transmission can still be calculated at and above the fundamental mode using thin plate
bending wave theory.
For most solid masonry/concrete walls the thin plate limit for bending waves falls in the midor high-frequency range. In Section 2.3.3.1 we discussed this thin plate limit in terms of a 10%
difference in the phase velocity between pure bending waves on thin plates and bending waves
on thick plates. Whilst it is referred to as a ‘limit’, thin plate theory does not instantly break down
at a specific frequency. For direct airborne sound insulation across a solid homogeneous plate,
thin plate theory can often be used up to a frequency of 4fB(thin) (Ljunggren, 1991). We will now
treat 4fB(thin) as a limit whilst acknowledging that it is not quite so clear-cut in practice. Errors
from using thin plate theory in the range fB(thin) ≤ f < 4fB(thin) are usually less than 3 dB. This
is often tolerable due to the uncertainty in predicting the total loss factor. Above 4fB(thin) the
airborne sound insulation effectively stops increasing with frequency and reaches a plateau
433
S o u n d
I n s u l a t i o n
Figure 4.13
Dilatational wave motion corresponding to thickness resonance on a plate.
with dips due to thickness resonances across the plate (Ljunggren, 1991). It can be useful
to visualize a longitudinal wave in air impinging upon a thick infinite plate at normal incidence
(Vér, 1992); and this wave exciting longitudinal waves across the plate thickness that efficiently
radiate sound into the air on the other side of the plate. In practice, we are interested in an
incident sound field that is diffuse and where the thickness modes are described by dilatational
waves (see Fig. 4.13). The mode frequencies for these thickness modes are calculated using,
rcD
(4.45)
fr =
2h
where r takes positive integer values 1, 2, 3, etc.
The phase velocity for dilatational waves, cD , is (Timoshenko and Goodier, 1970)
λ + 2μ
cD =
ρ
(4.46)
where Lamé’s constants, μ and λ, are
E
2(1 + ν)
νE
λ=
(1 + ν)(1 − 2ν)
μ=G=
(4.47)
(4.48)
and E and ν correspond to the material properties of the plate for longitudinal wave motion in
the thickness direction.
This sound transmission mechanism is not included within the SEA framework as there are
rarely more than two thickness modes in the high-frequency range for masonry/concrete plates.
In practice, the thickness resonances don’t always appear as distinct dips in the sound reduction
index. A plateau therefore provides a reasonable estimate for the sound reduction index and
can be estimated according to (Ljunggren, 1991)
ηint
ρcD
Rplateau = 20 lg
(4.49)
+ 10 lg
4ρ0 c0
0.02
If 4fB(thin) falls within the frequency range of interest, a smooth transition into the plateau region
can be achieved using R calculated from thin plate theory where R < Rplateau , and changing
over to Rplateau in frequency bands where R ≥ Rplateau . Due to increasing attenuation of the
dilatational waves with increasing frequency it is likely that the plateau will only extend a
few octaves above 4fB(thin) before starting to increase with frequency once more (Ljunggren,
1991). It is unusual to need to predict the sound insulation above 5000 Hz so for typical solid
masonry/concrete walls there is no need to consider the region beyond the plateau.
Using this approach with masonry/concrete plates gives measured values of R that are typically
within ±3 dB of the plateau region. Most thin surface finishes (such as plaster or render) that
are bonded to a thick plate will have negligible effect on the plateau region. Note that if the
434
Chapter 4
80
100 mm
200 mm
300 mm
Sound reduction index (dB)
70
60
f11
19
38
58
fc
341
171
114
fB(thin)
1040
520
347
f1
f2
f3
(Hz)
9812 19623 29435
4906 9812 14717
3271 6541 9812
50
40
30
20
50
80
125
200
315
500
800
1250
2000
3150
5000
One-third-octave-band centre frequency (Hz)
Figure 4.14
SEA predictions above the critical frequency (resonant transmission only) using thin plate theory combined with the plateau
region. Plate properties: Lx = 4 m, Ly = 2.5 m, h = 0.1/0.2/0.3 m, ρs = 80/160/240 kg/m2 , cL = 1900 m/s, ν = 0.2, internal
loss factor ηint = 0.0125, total loss factor η = 0.0125 + f−0.5 .
solid plate is orthotropic, and the thin plate theory uses an effective bending stiffness, then
the effective Young’s modulus will be different to the Young’s modulus that applies to the
dilatational modes in the thickness direction.
We have already seen an example of a thick plate in Fig. 4.12 for a 215 mm dense aggregate
masonry wall. The first two thickness modes of this wall are f1 = 9607 Hz and f2 = 19 215 Hz.
Although f1 is well above the building acoustics frequency range it still affects the top of the
high-frequency range.
For thermal purposes, some low-density walls are used in a wide range of thicknesses. Figure 4.14 shows an example of how the plateau region potentially forms an upper limit for the
airborne sound insulation in the high-frequency range. For the 200 and 300 mm plates the first
thickness resonance, f1 , lies within the building acoustics frequency range.
4.3.1.5
Infinite plates
So far we have mainly focused on finite plates. It might seem that further analysis of infinite
plates would be of little benefit; this is not the case. We will soon see that above the critical
frequency, the infinite plate formulae yield the same sound reduction index as for finite plates.
In addition, very large walls and floors that face into large spaces can be modelled as infinite
plates. Infinite plates also provide a convenient way of assessing the effect of different angles
of incidence. Whilst we usually need to know the sound reduction index for a diffuse incidence
sound field, there are some occasions where the incident sound field inside or outside a building
is highly directional.
We have already looked at non-resonant transmission across an infinite plate acting as a limp
mass. To complete the analysis we need to calculate the transmission coefficient when this
infinite plate has mass, bending stiffness, and damping.
435
S o u n d
I n s u l a t i o n
As in Section 4.3.1.2.1 we assume that the infinite plate lies in the xy plane and that a plane
wave is incident upon the plate at an angle, θ (refer back to Fig. 4.5). Therefore the sound
pressure for the incident, reflected, and transmitted waves can be taken from Eqs 4.27–4.29.
At this point it is necessary to make a brief return to the equation of motion for bending waves on
an isotropic plate. This is to account for the force per unit area applied by the sound pressure,
p(x, y, t), that drives the plate into motion. The equation of motion (Eq. 2.87) therefore becomes
−
∂Qy
∂Qx
∂2 η
dx dy −
dy dx + p(x, y, t) = ρs dx dy 2
∂x
∂y
∂t
which gives the wave equation for bending waves on a thin homogeneous plate as
4
∂4 η
∂4 η
∂2 η
∂ η
Bp
+
2
+
+ ρs 2 = p(x, y, t)
4
2
2
4
∂x
∂x ∂y
∂y
∂t
(4.50)
(4.51)
Note that to be consistent with Chapter 2 we will temporarily use η as the lateral displacement
of the plate. Once we have derived the transmission coefficient we will continue to use it to
represent loss factors again.
The sound pressure that drives the plate is
p(x, y, t) = p̂ exp(−ikx x) exp(−iky y) exp(iωt)
(4.52)
and the plate displacement must take the same form, hence
η(x, y, t) = η̂ exp(−ikx x) exp(−iky y) exp(iωt)
(4.53)
Substituting Eq. 4.53 into the wave equation (Eq. 4.51) gives
Bp [(kx2 + ky2 )2 − kB4 ] =
p
η
(4.54)
It is now convenient to define a surface impedance for the plate, Zp , as the ratio of the complex
sound pressure that drives the plate to the complex plate velocity,
Zp =
p
p
=
v
iωη
(4.55)
Hence Eqs 4.54 and 4.55 give the surface impedance as
Zp =
Bp 2
[(k + ky2 )2 − kB4 ]
iω x
(4.56)
Returning to the derivation of sound transmission, we are restricting our attention to the bending
wave that propagates in the positive x-direction, so we can set ky = 0. In the x-dimension, the
sound pressure and plate velocity must have the same spatial dependence, hence kx = k sin θ
and the surface impedance simplifies to
Zp =
Bp 4
(k sin4 θ − kB4 )
iω
(4.57)
Ignoring time dependence, Eq. 4.57 can now be used to describe the plate velocity in terms of
the sound pressure acting on the plate at z = 0,
v = v̂ exp(−ikx sin θ) =
436
(p̂i + p̂r − p̂t )
exp(−ikx sin θ)
Zp
(4.58)
Chapter 4
At z = 0 there must be continuity for the z-component of the particle velocity on both sides of the
plate (Eq. 4.32). This gives the relationship between the plate velocity and the sound pressures
described by Eq. 4.33. The angle-dependent transmission coefficient for an infinite isotropic
plate with mass and stiffness, τ∞,θ , can now be determined from Eqs 4.33 and 4.58, yielding
2
p̂t
1
Wt
= =
(4.59)
τ∞,θ =
2
Wi
p̂i
Z
1 + p cos θ
2ρ c
0 0
To include the effect of damping in the calculation of τ∞,θ we use the loss factor, η, and calculate
the surface impedance using Bp (1 + iη) instead of Bp . An infinite wall or floor clearly has no
boundaries at which it can lose energy via coupling losses to other walls and floors, so the loss
factor represents the internal loss factor for the plate material.
An example of the angle-dependent sound reduction index is shown in Fig. 4.15a for a 150 mm
thick infinite plate with the properties of cast in situ concrete. From Eq. 4.59 we see that when
θ = 90◦ , the sound reduction index will be 0 dB at all frequencies; this is due to trace matching
between the incident sound wave and the bending wave (Section 2.9.2). Whilst this angle can
potentially occur, it is of more interest to look at an angle very close to 90◦ to see how much
the sound reduction index differs from that for 90◦ ; for θ = 89◦ we see that there is a dip very
close to the critical frequency, above which the level climbs steeply. At any frequency above
the critical frequency there will always be an angle of incidence at which trace matching (also
called coincidence) occurs. For each angle of incidence (θ > 0◦ ) this results in a coincidence
dip in the sound reduction index; such dips can be seen in Fig. 4.15a where f > fc . The depth of
the coincidence dip is determined by the damping. At frequencies well above the coincidence
dip, each curve tends towards a slope that increases by 18 dB per doubling of frequency.
When θ = 0◦ there is no trace matching and the sound reduction index is the same as for an
infinite plate acting as a limp mass (Eq. 4.39) at all frequencies. When θ > 0◦ for f < fc the surface
impedance (Eq. 4.57) is primarily determined by the mass per unit area because k < kB , hence
Bp 4
(4.60)
k = iωρs
Zp ≈
iω B
Using this surface impedance to determine the angle-dependent transmission coefficient
(Eq. 4.59) gives the same equation as when the infinite plate acts as a limp mass (Eq. 4.34).
Well-below the critical frequency, we therefore find that sound transmission is only determined
by the mass per unit area.
To determine the sound reduction index for diffuse incidence, integration of the transmission
coefficient (Eq. 4.59) is carried out in the same way as for non-resonant transmission (Eq. 4.36).
This gives a smooth curve up to frequencies just above the critical frequency. At higher frequencies, the curve is full of peaks and troughs due to the coincidence dips. To give a smooth
curve above the critical frequency, the sound reduction index for diffuse incidence is calculated
using (Cremer, 1942)
f
R∞ = RNR,0◦ + 10 lg
(4.61)
− 1 + 10 lg η − 2 dB for f > fc
fc
The sound reduction index for diffuse incidence is shown in Fig. 4.15b for the same 150 mm
thick infinite plate. Above the critical frequency, the grey parts of the lowest three curves
indicate where curve fitting was used to connect the smooth curve given by Eq. 4.61 to the
value calculated by integration of the angle-dependent transmission coefficient.
437
S o u n d
(a)
I n s u l a t i o n
80
0°
Sound reduction index (dB)
70
30°
60
45°
50
60°
89°
40
30
20
10
fc ⫽ 114 Hz
0
0.1
(b)
1
f / fc
10
80
Infinite plate acting as a limp mass (diffuse incidence)
Sound reduction index (dB)
70
Infinite plate with mass, stiffness, and damping (diffuse incidence)
η
0.08
0.04
0.02
0.01
0.005
60
50
40
30
20
10
fc ⫽ 114 Hz
0
0.1
1
f /fc
10
Figure 4.15
Predicted airborne sound insulation for a 150 mm thick infinite plate: (a) angle-dependent sound reduction index for an
infinite plate (mass, stiffness, and damping) with the properties of cast in situ concrete and (b) sound reduction index (diffuse
incidence) for an infinite plate with different loss factors. Plate properties: h = 0.15 m, ρs = 330 kg/m2 , cL = 3800 m/s, ν = 0.2,
ηint = 0.005.
Figure 4.15b shows that well-below the critical frequency the infinite plate acts as a limp mass
and that the plate stiffness and damping have no effect on the sound reduction index. As the
frequency approaches the critical frequency, the effect of stiffness in the surface impedance
becomes more important and causes the curve to sag down below the diffuse incidence mass
law. This sagging curve leads to a dip at the critical frequency; the lowest frequency at which
coincidence occurs. At and above the critical frequency, the sound reduction index for diffuse incidence is under damping control. Referring back to the single angles of incidence in
438
Chapter 4
Fig. 4.15a it is clear that angular weighting carried out by integration of the angle-dependent
transmission coefficient gives a curve that is primarily determined by the coincidence dips. It
is these dips that are under damping control. At frequencies well-above the critical frequency
the sound reduction index for diffuse incidence increases by 3 dB per doubling of the loss factor. For a frequency-independent loss factor, the curves well-above the critical frequency tend
towards a slope that increases by 9 dB per doubling of frequency.
We can now make a link between the sound reduction index for a finite and an infinite plate.
For sound radiation from bending modes on finite plates, modes are classified as corner, edge,
or surface radiators depending on the frequency (Section 2.9.3). Below the critical frequency
the radiation from these modes depends on the plate boundaries; this is because most of the
sound power is radiated by corners or edges. Above the critical frequency, the plate modes
are all surface radiators and the perimeter length of the plate does not affect the radiation
efficiency. So although the existence of plate boundaries defines the modes, these boundaries
do not effect the sound radiation by these modes above the critical frequency. For this reason,
the sound insulation of finite plates above the critical frequency can be modelled using a plate
without boundaries, an infinite plate. The link that remains to be made is how the damping
affects the mean-square velocity of a finite plate and the equivalent infinite plate. For infinite
plates the only losses are material losses; hence the loss factor that is used to incorporate
damping is the internal loss factor. For finite plates we must also consider the coupling losses
because the plate is usually coupled to other parts of the structure (e.g. to other plates). So
to model a finite plate above the critical frequency using the infinite plate equation, we need
to replace the internal loss factor with the total loss factor of the finite plate. Above 4fB(thin) the
plateau region for thick plates is applicable.
An example to illustrate the equivalence of a finite and an infinite plate above the critical frequency is shown in Fig. 4.16 that compares measurement and prediction for a 150 mm concrete
slab. SEA is used to calculate resonant transmission for the finite plate above the critical frequency assuming a radiation efficiency of unity. The finite and the infinite plate predictions are
calculated for two different damping scenarios: one where the damping equals the internal
loss factor and the other where the damping equals the total loss factor of the finite plate that
was measured. Figure 4.16 shows that above the critical frequency in each scenario, the finite
plate SEA prediction tends towards the infinite plate prediction. The measurements only agree
with the finite plate SEA prediction and the infinite plate prediction that use the total loss factor, not the internal loss factor. Note that the measured data does not have a deep dip at the
critical frequency like the infinite plate theory. A shallow or unidentifiable dip is a feature of
masonry/concrete plates that leads to the use of radiation efficiencies of unity at and above
the critical frequency.
Unlike the internal loss factor, the total loss factor for masonry/concrete plates that are connected to other plates is usually frequency-dependent (Section 2.6.5). Therefore the slope
of the sound reduction index well-above the critical frequency will often differ from 9 dB per
doubling of frequency.
There is no equivalence between finite and infinite plates below the critical frequency. When
modelling finite plates using SEA we need to consider both non-resonant and resonant transmission. The relative importance of these paths to each other has previously been illustrated
with a number of examples in Section 4.3.1.3. These show that resonant or non-resonant
(mass law) transmission may dominate below the critical frequency. This is in contrast to
439
S o u n d
I n s u l a t i o n
80
Measured
Sound reduction index (dB)
70
SEA: Resonant transmission
60
Infinite plate with mass, stiffness,
and damping (diffuse incidence)
50
40
30
f11 = 46 Hz
20
fB(thin) = 1386 Hz
fc = 114 Hz
10
0
0.1
1
10
100
f /fc
Figure 4.16
Sound reduction index (diffuse incidence) for a 150 mm concrete slab. Comparison of predicted data (finite and infinite
plates) with measurements. For the predicted data, the upper curves are predictions using the measured total loss factor,
η ≈ 1.83f −0.62 , the lower curves are predictions using the internal loss factor, ηint = 0.005. Measurements according to ISO
140 Part 3. Plate properties: Lx = 3.4 m, Ly = 3.3 m, h = 0.15 m, ρs = 330 kg/m2 , cL = 3800 m/s, ν = 0.2. Measured data are
reproduced with permission from ODPM and BRE.
infinite plate theory where the surface impedance is determined by the mass term below the
critical frequency; this implies that there will only be mass-law (non-resonant) transmission.
For infinite plates, the field incidence transmission coefficient for an infinite plate acting as a
limp mass assumes that all angles of incidence are between 0◦ and 78◦ . In Section 4.3.1.2.2
we noted that although this sometimes provides fortuitous agreement with measured data
it is a less robust approach than using finite plate theory for non-resonant transmission. A
reduced range of angles is sometimes assumed in transmission suite measurements due to
the mounting of the test element within a niche. It is therefore of interest to assess the effect
of a reduced range of incident angles on the coincidence dip. To do this it is convenient to use
the ranges 0–90◦ , 0–78◦ , and 0–68◦ for a sheet of 6 mm glass as shown in Fig. 4.17. In this
example the effect of reducing the range of angles of incidence is to shift the coincidence dip
upwards by ≈100 Hz. This serves as a reminder of the uncertainty associated with predictions
in the vicinity of the critical frequency when the incident sound field cannot be considered as
diffuse.
4.3.1.6 Closely connected plates
It is usually clear when part of a wall or floor can be modelled as a single plate. It is not so
obvious when there are two or more plates that are very closely connected together over their
surfaces with an air gap of a millimetre or so. A common example occurs in lightweight walls
and floors when layers of board (e.g. plasterboard, plywood, chipboard) are closely connected
together with nails, screws, or dabs of adhesive. Other examples include sheets of metal that
are screwed or glued to timber doors. Modelling such multi-layer plates is complex and it is
440
Chapter 4
50
Sound reduction index (dB)
45
40
Infinite plate with mass, stiffness, and damping: 0–90˚ (diffuse incidence)
Infinite plate with mass, stiffness, and damping: 0–78˚ (field incidence)
Infinite plate with mass, stiffness, and damping: 0–68˚
Finite plate: Non-resonant transmission
35
30
25
20
fc = 2079 Hz
15
10
5
0.01
0.1
1
10
f /fc
Figure 4.17
Comparison of the sound reduction indices for 6 mm glass with different ranges for the angle of incidence. The frequency of the
coincidence dip is indicated by arrows. Plate properties: Lx = 1.5 m, Ly = 1.25 m, h = 0.006 m, ρs = 15 kg/m2 , cL = 5200 m/s,
ν = 0.24, total loss factor η = 0.024.
not usually possible to describe all of their behaviour by a solid plate, a laminated plate, or
a plate–cavity–plate system. In addition, variation in workmanship and uniformity of materials
means that the width of the air gap between two plates can differ considerably between similar
constructions. This usually means that design decisions concerning substitution of one type of
plate in a wall or floor for another type are reliant upon laboratory measurements.
The existence of a very narrow air gap sometimes causes a mass–spring–mass resonance (see
Section 4.3.5.1). Its presence is apparent with closely connected plates that are used in some,
but not all, lightweight wall and floor constructions (Warnock, 2000). This may partly be due to
non-uniform air gaps and variation in the strength of other sound transmission mechanisms.
For solid homogeneous plates we have already noted the difficulty in predicting excitation and
radiation at the critical frequency. When two layers of board are closely connected together
with screws, nails, or dabs of adhesive, the depth and width of the critical frequency dip often
changes. However, for boards with the same material properties the dip itself does not usually
shift down to a lower frequency as it would for a solid homogeneous plate with the combined
thickness of the two individual boards (Sharp, 1978; Sharp and Beauchamp, 1969). Figure 4.18
shows measurements where the critical frequency dip in the 3150 Hz band does not move to a
different band when two plates are closely connected together. This can be a desirable design
feature if it keeps the critical frequency dip at the top of the building acoustics frequency range.
By spot bonding boards with different critical frequencies it is sometimes possible to remove
any visible dip at either of the critical frequencies (Matsumoto et al., 2006). At frequencies wellbelow the lower critical frequency and well-below any mass–spring–mass resonance, a rough
estimate for the non-resonant transmission can be found by assuming that the closely connected plates act as a solid plate with the combined mass per unit area of the individual plates.
Resonant transmission is highly dependent on the type of connections between the plates.
441
S o u n d
I n s u l a t i o n
40
9.6 mm plywood
30
Two layers of 9.6 mm plywood
bolted together
Sound reduction index (dB)
20
10
12.7 mm plasterboard
40
Two layers of 12.7 mm plasterboard
spot bonded together with adhesive
30
20
10
50
80
125
200
315
500
800
1250 2000
One-third-octave-band centre frequency (Hz)
3150
5000
Figure 4.18
Comparison of measured sound reduction indices for single plates and two plates closely connected together. Measured data
are reproduced with permission from Sharp and Beauchamp (1969) and Sharp (1978).
For thin sheet materials that are closely connected, the internal losses tend to be frequency
dependent and higher than for the individual plates; however, the effect is highly dependent
upon the type of connection, air space, and the plates that are connected together (Trochidis,
1982). For two layers of 12.5 mm plasterboard that are screwed together, an increase in the
internal loss factor of up to 50% can occur below the critical frequency. Internal losses near the
critical frequency are more difficult to measure accurately, but they are still important because
they affect the depth of the critical frequency dip.
4.3.2 Orthotropic plates
Orthotropic plates have two critical frequencies; one in each of the two orthogonal directions.
Many plate materials are orthotropic to some degree. It is therefore important to identify when
a plate can simply be modelled as an isotropic plate, and when it is necessary to take full
account of their orthotropic nature. As a rule of thumb, if the critical frequencies are only a few
one-third-octave-bands apart then the plate can often be treated as an isotropic plate using
the effective bending stiffness; the errors are insignificant compared to the uncertainty in the
radiation efficiency and the critical frequencies themselves. When the critical frequencies are
separated by an octave or more, the effect on the sound reduction index is more significant.
Infinite plate theory is now used as the starting point to investigate the effect of the two critical
frequencies.
442
Chapter 4
4.3.2.1
Infinite plate theory
Infinite plate theory for thin homogeneous isotropic plates with mass, stiffness, and damping
can be extended to orthotropic plates (Heckl, 1960, 1981). The bending wave equation for an
orthotropic plate forced into motion by sound pressure is given by
Bp,x
∂4 η
∂4 η
∂2 η
∂4 η
+
2B
+
B
+
ρ
= p(x, y, t)
p,eff
p,y
s
∂x 4
∂x 2 ∂y 2
∂y 4
∂t 2
(4.62)
assuming that Bp,xy ≈ Bp,eff (Section 2.3.3.2).
Substituting the plate displacement (Eq. 4.53) into the wave equation (Eq. 4.62) gives the
surface impedance for an orthotropic plate as
&
'
Bp,eff kx2 ky2
Bp,y ky4
Bp,x kx4
Zp = iωρs 1 − 2
−2
− 2
(4.63)
ω ρs
ω 2 ρs
ω ρs
For an isotropic plate we focussed purely on the x-direction; this clearly cannot be done for
an orthotropic plate. To determine the constants, kx and ky , we maintain use of the angle, θ,
as the angle of incidence from the plate normal, and introduce φ as the angle between the
x-axis and the projection of the wave number for the incident wave onto the kx ky plane (see
Fig. 4.19). Hence,
kx = k sin θ cos φ
ky = k sin θ sin φ
(4.64)
The surface impedance can now be found in terms of the bending wave number in the x- and
y-directions by substituting Eq. 4.64 into Eq. 4.63,
⎡
⎤
2
2
2
sin
φ
cos
φ
4
4
Zp = iωρs ⎣1 −
+ 2
k sin θ ⎦
(4.65)
2
kB,x
kB,y
Using the same approach as for an isotropic infinite plate, the angle-dependent transmission
coefficient for an infinite orthotropic plate with mass and stiffness can now be found by inserting
the surface impedance (Eq. 4.65) into Eq. 4.59. To determine the diffuse incidence transmission
coefficient, the integration over θ and φ is carried out according to (Heckl, 1960)
2 π/2 1 d( sin2 θ)dφ
(4.66)
τ∞,d =
2
π 0
0
1 + Zp cos θ
2ρ c
0 0
The effect of damping is included in the surface impedance by using Bp,x (1 + iη) and Bp,y (1 + iη)
to calculate the bending wave numbers. The integration in Eq. 4.66 gives a smooth curve up
to frequencies just above the highest critical frequency, depending on the damping. Above this
the curve is full of peaks and troughs due to the coincidence dips. To give a smooth curve above
the highest critical frequency, the sound reduction index can be calculated using Eq. 4.61 for
an infinite isotropic plate by replacing fc with the effective critical frequency, fc,eff .
Figure 4.20 compares an isotropic concrete plate to orthotropic versions of the same plate that
have been created by increasing the stiffness in the x-dimension. The damping used in the
calculation corresponds to a total loss factor that is representative of a finite-sized concrete
slab connected to other masonry/concrete plates. Below the lowest critical frequency, the
infinite plate acts as a limp mass and the prediction is made using the diffuse incidence mass
443
S o u n d
I n s u l a t i o n
kz
k
θ
ky
φ
kx
Figure 4.19
Angles relating the wave number of the incident sound wave to kx , ky , and kz .
60
Sound reduction index (dB)
Infinite plate acting as a limp mass
Infinite isotropic plate with mass, stiffness,
and damping (cL,x ⫽ cL,y)
50
Infinite orthotropic plate with mass,
stiffness, and damping (cL,x ⫽ XcL,y)
40
fc,y ⫽ 114 Hz
30
X⫽
1.5
1.2
1.1
20
0.1
1
f/fc,y
10
Figure 4.20
Sound reduction index (diffuse incidence) for a 150 mm thick infinite plate (mass, stiffness, and damping) modelled as being
isotropic and orthotropic. The isotropic plate is based on the properties of cast in situ concrete. Orthotropic plates are
created by keeping cL,y and the other properties the same as the isotropic plate, but by increasing cL,x to give different
orthotropic plates. Plate properties: h = 0.15 m, ρs = 330 kg/m2 , ν = 0.2, total loss factor η = 0.005 + f −0.5 . Isotropic plate:
cL = cL,x = cL,y = 3800 m/s.
444
Chapter 4
law (Eqs 4.36 and 4.40). For 10% or 20% differences between cL,x and cL,y the coincidence
dip is seen to broaden slightly. Such differences are quite common for masonry/concrete
plates. When predicting the sound insulation it is reasonable to treat these orthotropic plates
as isotropic by using the effective bending stiffness. This is in contrast to a 50% difference
where there are two distinct critical frequency dips at fc,x = 76 Hz and fc,y = 114 Hz; this feature
will not be predicted when modelling the plate with an effective bending stiffness. The extended
critical frequency region is defined by the lowest and highest critical frequencies. At and above
the lowest critical frequency, the sound reduction index is affected by the plate damping. Above
the highest critical frequency, the plate acts as an isotropic plate with a bending stiffness equal
to the equivalent bending stiffness.
Many sheet materials used in lightweight walls and floors have low damping, and the coincidence dip is quite prominent. Whilst some sheet materials are significantly orthotropic, most
sheets (whether isotropic or orthotropic) are fixed to a framework of studs, battens, or joists. It is
often the frame that causes large changes to the bending stiffness so that they act as orthotropic
plates. For the moment we will ignore the complexity of including the frame. As before, we consider an infinite isotropic plate, and then increase the stiffness in the x-dimension to make it
orthotropic. The example in Fig. 4.21 assumes that the isotropic plate has similar properties to
plasterboard. For differences in the stiffness that are larger than 50% there is a characteristic
extended critical frequency region. For relatively small changes, such as 20%, the short slope
in this critical frequency region runs in the opposite direction. It has previously been mentioned
that it is difficult to accurately predict sound insulation at the critical frequency; the assumption
that a plate is perfectly isotropic is one of the reasons.
40
Infinite isotropic plate with mass, stiffness, and damping (cL,x ⫽ cL,y)
Sound reduction index (dB)
Infinite orthotropic plate with mass, stiffness, and damping (cL,x ⫽ XcL,y)
30
20
X⫽
fc,y ⫽ 3483 Hz
2.5
2
1.5
1.2
10
0.1
1
f/fc,y
10
Figure 4.21
Sound reduction index (diffuse incidence) for a 12.5 mm thick infinite plate (mass, stiffness, and damping) modelled as
being isotropic and orthotropic. The isotropic plate is based on the properties of plasterboard. Orthotropic plates are
created by keeping cL,y and the other properties the same as the isotropic plate, but by increasing cL,x to give different
orthotropic plates. Plate properties: h = 0.0125 m, ρs = 10.8 kg/m2 , ν = 0.3, internal loss factor ηint = 0.0141. Isotropic plate:
cL = cL,x = cL,y = 1490 m/s.
445
S o u n d
I n s u l a t i o n
80
f11 ≈ 41 Hz
fc,x⫽180 Hz fc,y⫽108 Hz
Sound reduction index (dB)
70
fB(thin) ≈ 442 Hz
60
50
Measured
40
SEA: Non-resonant transmission
30
SEA: Resonant transmission (thin plate)
Plateau for thick plates above 4fB(thin)
20
50
80
125
200
315
500
800
1250
2000
3150
5000
One-third-octave-band centre frequency (Hz)
Figure 4.22
Sound reduction index for an orthotropic 240 mm masonry wall (solid blocks) with a 5 mm plaster finish (each side).
Measurements according to ISO 140 Part 3. Plate properties: Lx = 3.96 m, Ly = 3.00 m, h = 0.24 m, ρs = 430 kg/m2 ,
cL,x = 1500 m/s, cL,y = 2500 m/s, ν = 0.2, measured total loss factor. The plateau is calculated using material properties
corresponding to the plate thickness: cL = 2500 m/s and an internal loss factor ηint = 0.01. Measured data are reproduced
with permission from Schmitz et al. (1999).
4.3.2.2 Masonry/concrete plates
Masonry walls and concrete floors with embedded reinforcement are usually orthotropic rather
than isotropic. For cast in situ concrete floors and solid block/brick walls (horizontal and vertical
joints mortared), the bending stiffness is not usually more than 20% higher in one direction than
the other. However walls and floors can also be built from beams and/or blocks that are rigidly
bonded together in one direction, but only touch or slot together in the other direction. The
orthotropic nature of these plates is often more pronounced, but for many plates it is possible
to use the effective bending stiffness to predict the sound insulation as if the plate were isotropic.
Examples for solid block walls with all joints mortared have already been seen in Section 4.3.1.3.
Here we look at an example of a highly orthotropic masonry wall from Schmitz et al. (1999). The
blocks are slotted together in the horizontal x-direction and are only rigidly connected with mortar in the vertical y-direction. This gives a bending stiffness that is 178% higher in the y-direction
than in the x-direction. Figure 4.22 shows the measured and predicted data. The predicted critical frequencies are in the 100 and 200 Hz one-third-octave-bands. Non-resonant and resonant
transmission are predicted using the effective bending stiffness; this gives an effective critical
frequency of 139 Hz. Resonant transmission is only shown at frequencies above the effective
critical frequency because of the complexity in predicting the radiation efficiency when there
are two widely spaced critical frequencies, relatively few plate modes, and laboratory surfaces
that form perpendicular baffles to the wall. Instead, the non-resonant transmission is shown;
this gives a rough indication of the maximum sound reduction index that can occur at these
low frequencies. Above the effective critical frequency, the predicted resonant transmission is
in close agreement with the measurements up until 4fB(thin) . Above this frequency the plateau
446
Chapter 4
80
Measured (Rigid connections: A, B, C, and D)
Measured (Rigid connection: A. Resilient connections: B, C, and D)
Sound reduction index (dB)
70
SEA: Resonant transmission (thin plate)
60
Plateau for thick plates
above 4fB(thin)
50
y
C
40
B
30
D
A
x
fc, x ⫽ 180 Hz fc, y ⫽108 Hz
20
50
80
125
200
315
500
800
1250 2000
One-third-octave-band centre frequency (Hz)
3150
5000
Figure 4.23
Orthotropic 240 mm masonry wall with rigid connections (mortar) or resilient connections (mineral fibre and sealant) at the
plate boundaries. Measurements according to ISO 140 Part 3. Plate properties are described in Fig. 4.22. Measured data are
reproduced with permission from Schmitz et al. (1999).
for the thick plate provides a reasonable estimate. The first thickness resonance occurs in the
5000 Hz one-third-octave-band for which the measurement shows a distinct dip.
The general shape of the extended critical frequency region corresponds to infinite plate theory
(diffuse incidence) for orthotropic plates. However, even when the damping is incorporated
using the total loss factor, this theory significantly underestimates the sound reduction index
both in the critical frequency region and below it.
For the orthotropic masonry wall discussed above, all four boundaries were fixed with rigid
mortar to the laboratory structure. Most walls in buildings are rigidly connected on at least
two sides. The effect of reducing the number of rigidly connected boundaries and replacing
them with resilient connections is to reduce the coupling losses from the plate. For isotropic
plates, it is not particularly relevant which rigid connection is replaced by a resilient connection;
however, the effect can be more marked with orthotropic plates. This can be seen in Fig. 4.23 by
comparing the masonry wall with rigid connections to the same wall with resilient connections on
three sides (Schmitz et al., 1999). Compared to the rigidly connected wall, the sound reduction
index is much lower between the highest critical frequency (fc,x ) and 4fB(thin) . This is due to the
reduction in the total loss factor and is confirmed by the agreement between the measurements
and the prediction for resonant transmission using the measured total loss factors. Concerning
the critical frequencies, it is worth considering what would happen if we didn’t know that the wall
was orthotropic. The measured sound reduction index for the wall with resilient connections
would then lead us to believe that there was only one critical frequency in the 200 Hz band;
and we would infer from this that the wall was isotropic. With the resilient connections, the
average coupling loss factor for the two plate boundaries that run parallel to the x-direction is
higher than the average of those that run parallel to the y-direction. In combination with the
low modal density and the low total loss factor, this causes the coincidence dip for fc,x to be
447
S o u n d
I n s u l a t i o n
80
Sound reduction index (dB)
150 mm concrete slab ( ρs ⫽ 320 kg/m2, fc ⫽ 108 Hz)
70
150 mm beam and block floor with 5 mm
levelling compound ( ρs ⫽ 313 kg/m2, fc,eff ≈ 150 Hz)
60
150 mm beam and block floor with 70 mm
screed ( ρs ⫽ 443 kg/m2, fc,eff ≈ 150 Hz)
50
40
30
20
50
80
125
200
315
500
800
1250
2000
3150
5000
One-third-octave-band centre frequency (Hz)
Figure 4.24
Comparison of a solid homogeneous concrete floor slab with an orthotropic beam and block floor (different surface finishes).
Both blocks and beams are solid, and the only rigid material that bonds them together is the surface finish. The measured
total loss factors for these three floors are within 2 dB of each other. Measurements according to ISO 140 Part 3. Measured
data are reproduced with permission from ODPM and BRE.
more prominent than fc,y . This indicates a degree of complexity with orthotropic plates when
applying a laboratory measurement to in situ where the boundary conditions are different.
Some orthotropic plates are built from beams and blocks without any material such as mortar
to rigidly bond them together within the cross-section of the plate. This occurs with some types
of beam and block floor. These floors commonly have a surface finish such as a screed or
levelling compound which means that they are only rigidly bonded together across their surface.
For such floors, the assumption of a homogeneous orthotropic plate to predict the airborne
sound insulation is not always appropriate. An example is shown in Fig. 4.24 comparing measurements on a solid homogenous concrete floor slab to one specific type of beam and block
floor (solid beams and blocks) with different thicknesses of surface finish.
4.3.2.3 Masonry/concrete plates containing hollows
Many blocks, bricks, and slabs are not homogeneous and isotropic due to the hollows inside
them. Hollows reduce the mass per unit area, cause orthotropic or anisotropic behaviour, and
form cavities that may support one-, two-, or three-dimensional sound fields within the plate.
In concrete floor slabs with circular hollows, the sound field in the hollows usually has negligible
effect on the direct transmission of airborne sound across the slab (Vinokur, 1995). Resonant
transmission across heavyweight concrete slabs dominates over most of the building acoustics
frequency range. The main effect of the hollows is to give a different bending stiffness in different
directions compared to a solid slab, and to change the coupling losses to the supporting walls.
Apart from hollows with large volumes inside diaphragm walls, the cavities inside many
masonry bricks and blocks are usually so small that they do not support modes over the
448
Chapter 4
60
Thickness
resonance
Sound reduction index (dB)
380 mm hollow brick wall
(plaster on both sides)
ρs ⫽ 420 kg/m2
55
50
380 mm
45
Hollow brick
40
50
80
125
200
315
500
800
1250
2000
3150
5000
One-third-octave-band centre frequency (Hz)
Figure 4.25
Measured sound reduction index for a hollow brick wall. Measurements according to ISO 140 Part 3. Measured data are
reproduced with permission from Lang (1993).
majority of the building acoustics frequency range. Common examples are perforated bricks
with narrow hollows that are used to achieve high levels of thermal insulation. There are a
wide variety of designs for which the sound insulation is highly variable. This is mainly due to a
deep dip at the lowest thickness resonance that typically occurs in the mid- or high-frequency
range (e.g. see Fringuellino and Smith, 1999; Scholl and Weber, 1998). An example is shown
in Fig. 4.25 (Lang, 1993). Unlike thickness resonances with solid masonry/concrete walls
(Section 4.3.1.4), the frequency of the lowest thickness resonance is more difficult to predict
but can be identified by measurements on single blocks (Weber and Bückle, 1998).
4.3.2.4 Profiled plates
Profiled plates are mainly used to increase the structural strength of thin sheet materials. Profiles make the bending stiffness of the plate much higher in the direction along the raised
profile, than in the perpendicular direction. This results in different critical frequencies in the
two orthogonal directions, sometimes remarkably different. Consequently, it is useful to estimate the critical frequencies of profiled plates from the bending stiffness for each direction. The
bending stiffness calculation is described in Section 2.3.3.2.1. Profiled sheet materials are commonly encountered as metal cladding used for facades and roofing. As an example, consider a
flat sheet of 0.65 mm thick steel with a critical frequency of ≈19 kHz. When this plate is given a
trapezoidal profile commonly used for cladding, the lowest critical frequency may be as low as
150 Hz in the direction of the raised profiles. In the orthogonal direction, the critical frequency
will be higher than for the flat sheet (i.e. >19 kHz). For these orthotropic plates, the extended
critical frequency region covers the majority of the building acoustics frequency range.
When the two critical frequencies are widely spaced (but fall within the building acoustics
frequency range) it is not useful to predict the sound insulation by modelling the profiled plate as
449
S o u n d
I n s u l a t i o n
an isotropic plate using the effective bending stiffness. This would give a single coincidence dip
within the extended critical frequency region. However, this approach can be used to estimate
the sound insulation at frequencies above the highest critical frequency of the profiled plate.
Below the lowest critical frequency, the sound reduction index for profiled plates with a low
mass per unit area (e.g. metal cladding) can be assumed to be dominated by non-resonant
transmission. An infinite plate model assuming diffuse incidence will overestimate the masslaw transmission; hence it is necessary to assume field incidence, or to use finite plate theory
to calculate the non-resonant transmission.
In the extended critical frequency region, the model for an infinite orthotropic plate can be used
to estimate the sound reduction index. However, using Eq. 4.66 underestimates the sound
reduction index in the lower part of the extended critical frequency region. This is because
diffuse incidence for infinite plates does not adequately represent non-resonant transmission
across finite plates. To overcome this, the following approximation based on Eq. 4.66 can be
used in the extended critical frequency region, Min(fc,x , fc,y ) < f < Max(fc,x , fc,y ) (Heckl, 1960)
τ∞,d ≈
2
ρ0 c0 Min(fc,x , fc,y )
4f
ln
πωρs
f
Min(fc,x , fc,y )
(4.67)
This approximation provides a relatively smooth link from the field incidence mass law (infinite
plate) below the lowest critical frequency to the infinite orthotropic plate model. Although the
approximation ignores the effect of damping, this error can be considered as negligible in
the context of other assumptions. More important is the assumption in the bending stiffness
calculations for the profiled plate (Section 2.3.3.2.1) that allows the plate to be represented by
50
Measured
Infinite orthotropic plate (approximation)
Infinite plate: Field incidence (non-resonant transmission)
Sound reduction index (dB)
40
Steel thickness
0.45 mm
0.65 mm
0.90 mm
30
20
b2
10
b1
θ1
θ3
b3
b4
dR
0
50
80
125
200
315
500
800
1250 2000
One-third-octave-band centre frequency (Hz)
3150
5000
Figure 4.26
Measured sound reduction indices for profiled steel cladding of different thicknesses. Measurements according to ISO 140
Part 3. Plate properties: Lx = 2.4 m, Ly = 3 m, ρ = 7800 kg/m3 , cL (steel) = 5270 m/s, ν = 0.28 (Note: x- and y-directions
are defined so that the raised profiles run parallel to the x -direction). Profile: b1 = b3 = 46.1 mm, b2 = b4 = 95 mm,
θ1 = θ3 = 49.4◦ , θ2 = θ4 = 0◦ , dR = 250 mm. For all three thicknesses, fc,x = 216 Hz, fc,y > 15 kHz. Measured data are
reproduced with permission from Lam and Windle (1995).
450
Chapter 4
a flat infinite orthotropic plate. This assumption concerns the bending wavelength being much
larger than the repetition distance, dR , of the profile. When this assumption is not satisfied,
the shape of the profile within the repetition distance will determine the mode shapes, and the
radiation efficiency of these modes.
For profiled metal cladding with trapezoidal profiles, it is often found that λB,y < dR , over most
of the building acoustics frequency range; it is only in the low- and mid-frequency ranges that
λB,eff > dR . The effect of this can be seen in Fig. 4.26. This compares the infinite orthotropic
plate approximation (Eq. 4.67) in the extended critical frequency region with measurements
on different thickness plates with the same profile. Differences between measurements and
predictions in the extended critical frequency region are generally less than 5 dB; even though
it is only in the low- and mid-frequency ranges that λB,eff > dR . However, one should be aware
that the sound insulation of profiled plates is characterized by fluctuations that depend on the
shape of the profile (Cederfeldt, 1974; Lam and Windle, 1995). Figure 4.27 gives an indication
of the variation between different profiles for a single thickness of plate material. The infinite
plate approximation takes no account of the plate modes that are determined by the shape
of the profile. It is resonant transmission via these plate modes that causes significant dips in
the sound reduction index at frequencies above the lowest critical frequency (see Fig. 4.27).
To some extent it is possible to predict these fluctuations by using finite element methods to
calculate the plate modes (Lam and Windle, 1995).
4.3.3
Low-frequency range
The definition of the sound reduction index gives the impression that the measured sound
insulation of any test element can be described in such a way that it is independent of the source
and receiving room, and can be referenced to the element area such that the result can be used
for different sizes of the same test element. In the low-frequency range, simple scaling in terms
of size is not always possible because non-resonant (mass law) transmission for finite plates
depends on the plate dimensions, and resonant transmission can be highly variable when the
rooms and/or the plates have low modal density. The assumption that the sound reduction index
is independent of the source and receiving room tends to break down somewhere in the lowfrequency range; most noticeably below 100 Hz. It is in this range that statistical descriptions
of sound and vibration fields and the coupling between these fields no longer satisfy some of
the assumptions made in SEA. On the basis that these fields can no longer be described in a
statistical sense, it is assumed that the mode shapes and mode frequencies can be calculated
exactly using deterministic models. Analytic descriptions of the modal sound and vibration
fields or finite element methods can then be used to couple together the modes of the source
and receiving rooms via a separating wall or floor (Gagliardini et al., 1991; Kropp et al., 1994;
Maluski and Gibbs, 2000; Osipov et al., 1997; Pietrzyk and Kihlman, 1997). Deterministic
models have mainly been used for a solid rectangular plate between two rectangular rooms to
shed light on some of the factors that cause large variations in low-frequency airborne sound
insulation for nominally identical separating walls, floors, or windows.
Parametric studies with deterministic models have been used to determine the sound reduction
index of a solid homogeneous isotropic plate when placed between different sizes of room;
these cover a range of room dimensions that can typically be found in dwellings. The results
show that the sound reduction index in the low-frequency range is highly dependent on the
source and receiving room dimensions (i.e. their modes) as well as the test element itself
(Kropp et al., 1994; Osipov et al., 1997; Pietrzyk and Kihlman, 1997). This occurs whether
451
S o u n d
I n s u l a t i o n
50
Plate A
Plate B
Sound reduction index (dB)
40
Plate C
Plate D
30
20
b2
10
b1
θ1
θ3
b3
b4
dR
0
50
80
125
200
315
500
800
1250
2000
3150
5000
One-third-octave-band centre frequency (Hz)
Figure 4.27
Measured sound reduction indices for 0.65 mm profiled steel cladding with different profiles. Measurements according to
ISO 140 Part 3. Plate properties: Lx = 2.4 m, Ly = 3 m, ρ = 7800 kg/m3 , cL (steel) = 5270 m/s, ν = 0.28. Profile for plate A:
b1 = b3 = 46.1 mm, b2 = b4 = 95 mm, θ1 = θ3 = 49.4◦ , θ2 = θ4 = 0◦ , dR = 250 mm (fc,x = 216 Hz, fc,y ≈ 20 kHz). Profile for
plate B: b1 = b3 = 39.4 mm, b2 = b4 = 57 mm, θ1 = θ3 = 62.8◦ , θ2 = θ4 = 0◦ , dR = 150 mm (fc,x = 210 Hz, fc,y ≈ 21 kHz).
Profile for plate C: b1 = b3 = 62.7 mm, b2 = b4 = 95 mm, θ1 = θ3 = 61.4◦ , θ2 = θ4 = 0◦ , dR = 250 mm (fc,x = 134 Hz,
fc,y ≈ 21 kHz). Profile for plate D: b1 = b3 = 46.1 mm, b2 = 95 mm, b4 = 20 mm, θ1 = θ3 = 49.4◦ , θ2 = θ4 = 0◦ , dR = 175 mm
(fc,x = 246 Hz, fc,y ≈ 21 kHz). Measured data are reproduced with permission from Lam and Windle (1995).
the dominant transmission mechanism is resonant or non-resonant (mass law) transmission.
A study on a solid plate (30 kg/m2 ) across which there was only non-resonant (mass law)
transmission, gave a standard deviation of 3–5 dB with a range of approximately 15 dB for
the sound reduction index (one-third-octave-bands) in the low-frequency range (Kropp et al.,
1994). Similar studies on resonant transmission across solid concrete plates gave higher
standard deviations with values up to 10 dB (Osipov et al., 1997; Pietrzyk and Kihlman,
1997). A transmission mechanism that was not included in these particular studies was the
mass–spring–mass resonance frequency; this usually occurs in the low-frequency range with
plate–cavity–plate constructions and is also likely to cause high standard deviations.
These parametric studies confirm earlier observations (Heckl and Seifert, 1958) that lowfrequency airborne sound transmission tends to be lower between rooms with identical
dimensions than between rooms with different dimensions. This can be attributed to strong
coupling between identical room modes via the separating wall or floor. It is relatively easy to
visualize this for identical axial modes in each room that are perpendicular to the separating
wall or floor. For identical rooms a quick assessment of potentially adverse dips in the airborne
sound insulation can be made by calculating the mode frequencies for the first few axial modes;
dips may be exacerbated if they coincide with the mass–spring–mass resonance frequency of
the separating wall or floor. However, it is not just identical room dimensions that cause low
sound insulation. There can be other combinations of plate modes and room modes that are
particularly well-coupled and result in dips in the sound insulation (Osipov et al., 1997; Pietrzyk
and Kihlman, 1997).
452
Chapter 4
In the laboratory, accurate low-frequency measurements with good reproducibility can be
achieved by using sound intensity as discussed in Section 3.5.1.1.1. This allows a fairer
comparison of test elements. However, it does not overcome the difficulty in applying such
measurements to estimate the airborne sound insulation in situ when the same element is
installed between rooms with different room dimensions. For simple test elements such as
solid plates there is the potential to predict the airborne sound insulation using deterministic models. However, this is not always straightforward. Using a single deterministic model
assumes that it is possible to accurately determine the eigenfrequencies and eigenfunctions
for a specific pair of rooms and a specific separating element. This takes no account of the
uncertainty in describing their modal features. The modal overlap is not usually high, and local
mode frequencies of the rooms and the plate(s) tend to be widely spaced apart. For this reason,
any deterministic model will output a sound pressure level difference between the source and
the receiving rooms that is characterized by peaks and troughs. Uncertainty in one mode frequency in the model can shift a peak or trough into an adjacent one-third-octave-band. Using
only one deterministic model takes no account of any uncertainty. It assumes that the material
properties of the wall and its boundary conditions are known in sufficient detail to accurately
calculate the mode frequency and mode shape; this is rarely the case. It also assumes that the
impedances of the room boundaries are known exactly, room temperatures are fixed, and that
the source room will only be excited at one or more specific loudspeaker positions; this is rarely,
if ever, the case. Comparisons of single deterministic models with a laboratory measurement
often show good agreement, but there are almost always differences of up to 10 dB in an individual one-third-octave-band (Kropp et al., 1994; Maluski and Gibbs, 2000; Osipov et al., 1997).
To use a deterministic model as a design tool it is necessary to account for uncertainty in the
variables that determine the modal response (e.g. dimensions, damping boundary conditions,
temperature). This can potentially be done using Monte-Carlo simulations, but the computation
time may only be justifiable in critical applications such as the design of recording studios.
With attached dwellings it is quite common to have rooms with nominally identical dimensions
on either side of the separating wall or floor. Theoretically this is the least favourable room layout
from an acoustic point of view whilst it is the most practical from a construction viewpoint. There
are a number of reasons why the adverse effects of identical room dimensions on sound insulation are not always apparent, identifiable, or predictable in the field. Firstly, adverse effects
will not always be detected by field sound insulation measurements. These measurements are
primarily intended for sound fields that can be considered as diffuse. Sound pressure levels
are usually measured in the central zone of the room at certain minimum distances from the
room boundaries. In small rooms with relatively few room modes, these positions tend to be
on, or near the nodal planes of these modes. Therefore it is possible to measure a low spatial
average sound pressure level in the source room, receiving room, or both rooms. Secondly,
there is the question as to which sound reduction index (if any) can be empirically adjusted to
calculate the sound insulation between rooms with identical dimensions in the field. A laboratory measurement on a separating element with low modal density will only apply to the size
of test element and boundary conditions used in the laboratory; this assumes that the effect
of the laboratory receiving room can be removed by using intensity measurements and effectively converting this room into a duct (Section 3.5.1.1.1). Thirdly, occupied rooms in different
dwellings tend to be filled with large items of furniture in different ways; hence the room modes
are unlikely to be identical in practice. The issue of identical axial modes may be more relevant
to buildings such as hotels, where rooms are furnished in exactly the same way in a symmetrical fashion about the separating wall. Fourthly, the temperature is not usually exactly the same
453
S o u n d
I n s u l a t i o n
in each room so the mode frequencies will differ in each room; changing the temperature in
either the source or receiving room by at least a few degrees centigrade can change the sound
pressure level difference in a single one-third-octave-band by at least a few decibels in the
low-frequency range (e.g. see Scholes, 1969). Fifthly, there are many different types of constructions used for separating walls and floors for which there are different sound transmission
mechanisms dominating in the low-frequency range. Empirical corrections to the sound reduction index of the separating element have previously been attempted by splitting constructions
into lightweight and heavyweight categories (Gibbs and Maluski, 2004). To predict the airborne sound insulation in situ where flanking transmission is present there are relatively few
simple, quick, and practical options other than to use SEA or SEA-based models and make an
allowance for uncertainty in the low-frequency range. This allowance can be made on an empirical basis from previous measurements, or with Monte-Carlo simulations using deterministic
models.
4.3.4 Membranes
Roofs in commercial buildings and very large spaces such as arenas are sometimes formed
from membranes using materials such as PTFE or ETFE. Each membrane typically has a
mass per unit area less than 2 kg/m2 and is under tension. For membranes it is not bending
wave motion but wave motion due to tension of the membrane. The phase velocity, cm , is
independent of frequency and is given by (Morse and Ingard, 1968)
T
cm =
(4.68)
ρs
where T is the tension per unit length around the edge of the membrane (N/m).
The modes of a membrane stretched over a rectangular frame are given by (Morse and Ingard,
1968)
2
p 2
cm
q
fp,q =
+
(4.69)
2
Lx
Ly
where p and q take positive integer values 1, 2, 3, etc.
A reasonable estimate for the airborne sound insulation of non-porous single layer membranes
can be calculated using infinite plate theory for an infinite plate acting as a limp mass (Weber and
Mehra, 2002). Practical roof constructions usually comprise two membranes separated by an
air gap; however it is not suitable to predict the performance by assuming a plate–cavity–plate
system.
The airborne sound insulation in the low-frequency range can be increased by loading a membrane with additional weights, and then increasing the tension. To reduce adverse effects in
the mid-frequency range this membrane is then used as part of a double layer with a wide air
gap as shown in Fig. 4.28 (Hashimoto et al., 1996).
4.3.5
Plate–cavity–plate systems
We will now use SEA models to look at airborne sound transmission across a plate–cavity–plate
system such as a cavity wall or floor. As with solid plates we will only consider bending wave
454
Chapter 4
40
0.38 mm single layer PTFE with additional weights, tension 100 kg/m
35
0.38 mm single layer PTFE with additional weights, tension 200 kg/m
Double layer (0.8 mm – 0.1 m air gap – 0.38 mm with weights, tension 100 kg/m)
30
Double layer (0.8 mm – 1 m air gap – 0.38 mm with weights, tension 200 kg/m)
Sound reduction index (dB)
25
20
15
10
5
0.8 mm single layer PTFE ( ρs ⫽ 1.3 kg/m2 )
15
Field incidence
Diffuse incidence
Infinite plate
non-resonant
transmission
10
5
0
100
125
160
200
250
315
400
500
630
800
1000
One-third-octave-band centre frequency (Hz)
Figure 4.28
Measured sound reduction indices for different membrane roof elements. For the single layer membrane, infinite plate theory
(non-resonant transmission) is shown for comparison with the measurement. Measured data are reproduced with permission
from Hashimoto et al. (1996).
motion. The five-subsystem SEA model for a plate–cavity–plate system is shown in Fig. 4.29.
Note that by limiting the model to five subsystems we are ignoring flanking transmission by
assuming that vibration is transmitted from each plate to the rest of the building or laboratory
structure, but not vice versa. Vibration can be transmitted between the two plates via structural
connections as well as via the sound field in the cavity.
Structural connections often form an important transmission path. Examples of structural connections between plates include frameworks of beams, foundations, wall ties, and other plates
that form flanking walls or floors. In some cases the structural coupling can simply be introduced as a coupling loss factor between the plates, in other cases it is necessary to expand
the model by adding other beam and plate subsystems.
The sound field in the cavity is affected by the absorption within it, or around its boundaries.
Cavities will generally be referred to as being empty (this includes cavities with an absorbent
material around the perimeter of the cavity), or as being partly or fully filled with porous material
(partly filled refers to filling across the cavity depth).
455
S o u n d
Win(1)
I n s u l a t i o n
W13
Non-resonant transmission
W31
W23
W12
Subsystem
Subsystem
2
1
W21
W32
Wd(1)
Wd(2)
Subsystem
3
W35
Non-resonant transmission
W53
W45
W34
Subsystem
Subsystem
4
5
W43
W54
Wd(3)
Wd(4)
Wd(5)
W42
Structural coupling
W24
Source room
(1)
Plate
(2)
Cavity
(3)
Plate
(4)
Receiving room
(5)
Figure 4.29
Five-subsystem SEA model for airborne sound transmission between a source room and a receiving room across a plate–
cavity–plate system. Arrows with dashed lines represent non-resonant transmission, and arrows with solid lines represent
resonant transmission.
Plate–cavity–plate systems superficially appear to be simple; they very rarely are. One single
type of prediction model (based on either finite or infinite plates, deterministic or statistical
approaches) is not usually able to accurately predict all important features of sound transmission for even quite simple cavity walls, floors, and windows. For design work, neither
measurement nor prediction on their own will solve all problems. It is usually only in combination that sufficient insight into the transmission mechanisms can be gained. For laboratory
measurements it is useful to have a model that can give some basic insight into the transmission process. This can be used to decide how the plate–cavity–plate system should be
mounted in the laboratory to give a measurement that will be relevant in situ.
The advantage of SEA lies in the fact that structural coupling between the plates and flanking
transmission are often important with cavity walls and floors in situ. Its disadvantages become
more apparent when these two features are absent. SEA is not only useful for quantifying the
airborne sound insulation of plate–cavity–plate systems in the laboratory and the field but also
in qualitative discussions of the transmission mechanisms. Plates and cavities in buildings
usually support modal behaviour over the majority of the building acoustics frequency range;
but not always with high mode counts. Alternative models to SEA that are based on infinite plate
theories or numerical methods can also be used to model a wide variety of plate–cavity–plate
systems; although when compared with measurements the errors are similar to SEA. To be of
practical use, plate–cavity–plate models based on infinite plates usually need to be modified in
some way to deal with finite plate size (Villot et al., 2001). The errors that occur with infinite plate
models can be attributed to modal behaviour that is not considered in the model. In choosing
456
Chapter 4
an appropriate model we sometimes find that the mode counts in frequency bands are too low
or too high to suit one single model over the entire building acoustics frequency range. SEA is
by no means suited to modelling every plate–cavity–plate system over this frequency range;
however, the process of trying to make a system fit an SEA model usually sheds light on the
key features that need to be considered in the design.
This section uses three examples of plate–cavity–plate systems to form a basis from which
the various transmission mechanisms can be discussed. Before we look at these examples it
is necessary to consider another type of non-resonant transmission between two rooms connected via a plate–cavity–plate system. This is modelled by assuming that the gas in the cavity
acts as a spring and the plates act as lump masses. The result is a mass–spring–mass resonance frequency that often falls in the low-frequency range. Plate–cavity–plate models based
on infinite plates very usefully show that below the mass–spring–mass resonance frequency the
two plates effectively act as a single plate with the combined mass per unit area and thickness
of the two individual plates (Beranek and Work, 1949; Fahy, 1985; London, 1950). Hence the
mass–spring–mass resonance frequency can be used to estimate the frequency below which
the cavity wall acts as a single plate, and above which it acts as a plate–cavity–plate system.
4.3.5.1 Mass–spring–mass resonance
In the low-frequency range there is a non-resonant transmission mechanism across a cavity
wall or floor that can be modelled by treating the gas contained within an empty cavity as a
spring and each plate as a lump mass. It is assumed that the plates forming the cavity are nonporous or have a sufficiently high airflow resistivity that the gas does not enter the plates when
it undergoes compression. For any gas in an enclosed cavity with a depth, Lz , the dynamic
stiffness per unit area of the gas, sg′ , is
sg′ =
K
γP
=
Lz
Lz
(4.70)
where K is the bulk compression modulus of the gas.
The mass–spring–mass resonance frequency has already been calculated using an equivalent circuit in Section 3.11.3.2.1. To use Eq. 3.200 with a plate–cavity–plate system it is only
necessary to replace the stiffness with the stiffness per unit area, and each mass with the mass
per unit area. This gives
!
sg′
1 "
"
fmsm =
(4.71)
"
#
ρs1 ρs2
2π
ρs1 + ρs2
where ρs1 and ρs2 are the mass per unit area of plates 1 and 2 respectively.
For air-filled cavities (adiabatic compression, γ = 1.4), the dynamic stiffness per unit area of
the enclosed air, sa′ , is given by
sa′ =
ρ0 c02
1.4P0
=
Lz
Lz
so the mass–spring–mass resonance frequency for air at 20◦ C is
ρs1 + ρs2
fmsm = 60
ρs1 ρs2 Lz
(4.72)
(4.73)
457
S o u n d
I n s u l a t i o n
Mass–spring–mass resonance frequency (Hz)
1000
Cavity
depth
(mm)
5
100
10
20
30
40
50
75
100
200
10
0
10 20 30 40 50 60 70 80 90 100 110 120 130 140 150 160 170 180 190 200
Mass per unit area, ρs1 ⫽ ρs2 (kg/m2)
Figure 4.30
Mass–spring–mass resonance frequencies for a plate–cavity–plate system where the cavity is filled with air, and both plates
have the same mass per unit area.
Figure 4.30 shows calculated resonance frequencies for air-filled cavities (Eq. 4.73) where
both plates have the same mass per unit area.
When a cavity is partly filled across its depth with a porous material that has a porosity close to
unity (e.g. mineral wool), then Eq. 4.73 can also be used to estimate the resonance frequency.
If the porous material has a low porosity, the cavity depth, Lz , can be replaced by an effective
cavity depth, Lz,eff , given by
Lz,eff = (Lz − d) + φd
(4.74)
where d is the thickness of the porous material and φ is the porosity of the porous material.
For fully filled cavities the plates can be strongly coupled together by the porous material and
the plate may need to be modelled as a sandwich panel. If so, the simple lumped element
model will no longer be appropriate. However, in some cases an estimate of the resonance
frequency can be calculated if the dynamic stiffness of the porous material is considered as
the combined stiffness of the enclosed air and the skeletal frame of the material, s′ . This can
be measured in the same way as for resilient materials used under floating floors if the static
load does not significantly alter the dynamic stiffness (Section 3.11.3.1). The mass–spring–
mass resonance frequency can then be calculated using Eq. 4.71 by replacing sg′ with s′ . In
some cases the dynamic stiffness of the skeletal frame may be negligible and the resonance
frequency will only be determined by the air contained within the porous material. This occurs
458
Chapter 4
with very low-density mineral wool used in fully filled cavity walls for which it is necessary to
reconsider the assumption of adiabatic compression for the enclosed air. For cavities filled with
fibrous materials, adiabatic compression is only a reasonable assumption in the high-frequency
range. The mass–spring–mass resonance usually occurs in the low-frequency range where
isothermal compression (γ = 1.0) occurs due to heat conduction by the fibres, hence
ρs1 + ρs2
fmsm = 51
(4.75)
ρs1 ρs2 Lz
Laboratory measurements on simple plate–cavity–plate systems (i.e. no framework other than
around the perimeter of the plate) show that the mass–spring–mass resonance for a cavity
that is fully filled with a fibrous material tends to occur at a lower frequency than with an empty
cavity (e.g. see Nightingale, 1999). For fully filled cavities it is possible to treat the porous
material as an equivalent gas and to calculate the bulk compression modulus for Eq. 4.70
by assuming an appropriate model for the porous material (Allard, 1993). However, there are
limits as to how useful it is to pursue more accurate calculations of the mass–spring–mass
resonance frequency. Many plates and cavities support modal behaviour in the vicinity of the
resonance frequency; therefore the lumped element model assuming simple masses and a
spring becomes less appropriate. Lightweight walls and floors, as well as some lightweight
linings on heavyweight wall and floor bases, require a frame to support the sheet material.
The type of frame, stud spacing, and screw spacing can change the resonance frequency by
constraining and stiffening the sheet material; this means it can no longer be treated as a lump
mass. Measurements on plasterboard cavity walls indicate that the resonance frequency is
often highest for timber studs with closely spaced screws and lowest for light steel studs with
widely spaced screws; reducing the stud spacing tends to increase the frequency at which
resonance occurs (Quirt et al., 1995).
Cavity dimensions differ widely between various plate–cavity–plate systems. When the mass–
spring–mass resonance frequency is below the lowest cavity mode there is usually a marked
dip in the airborne sound insulation due to the resonance. It is hard to generalize when it occurs
above the lowest cavity mode. In this situation the air is not acting as a lump spring and the
mass–spring–mass resonance frequency can be regarded as a transition point from single
plate behaviour to plate–cavity–plate behaviour. However, this may still introduce a significant
dip, a step, or change in slope in the airborne sound insulation curve.
When designing plate–cavity–plate systems, the aim is usually to choose the cavity depth and
the mass per unit area of the plates so that the mass–spring–mass resonance falls well-below
the important frequency range. One good reason to do this is due to a lack of models that can
accurately predict the frequency range over which the dip will occur, and how deep any dip
will be. The risk of including the resonance frequency within the important frequency range is
usually only taken when measurements are available to assess the adverse effects (if any).
Some example laboratory measurements on plate–cavity–plate systems with distinct mass–
spring–mass resonance dips are shown in Fig. 4.31. The dip is typically less than 8 dB when
compared to the level in the adjacent one-third-octave-bands.
Modelling the air in the cavity as a spring can potentially be incorporated into an SEA model
over a wide frequency range by coupling the two plates together using the air stiffness (Brekke,
1981; Craik and Wilson, 1995). The stiffness of the air spring will decrease with increasing
cavity depth which can result in an increase in the airborne sound insulation. This mechanism
can be included as a form of structural coupling that directly couples plate subsystems 2 and
459
S o u n d
I n s u l a t i o n
Masonry wall with plasterboard lining fixed with adhesive dabs
Mass: 12.5 mm plasterboard (fc = 3483 Hz)
Spring: 8–12 mm air gap
(variation is due to workmanship)
Mass: 215 mm masonry wall
(fc ⫽ 95 Hz)
Window with insulating glass unit
Mass: 6 mm glass (fc ⫽ 2079 Hz)
Spring: 12 mm air gap
Mass: 6 mm glass (fc ⫽ 2079 Hz)
Sound reduction index (dB arbitrary values)
fmsm
dip
fmsm
dip
10 dB
Floating floor on concrete slab
Mass: 18 mm chipboard spot
bonded to 19 mm plasterboard
Spring: 30 mm mineral wool
Mass: 150 mm concrete slab
(fc ⫽109 Hz)
fmsm
dip
50
80
125
200
315
500
800
1250
2000
3150
5000
One-third-octave-band centre frequency (Hz)
Figure 4.31
Examples of plate–cavity–plate systems with mass–spring–mass resonance frequencies in the building acoustics frequency
range. Measured data are reproduced with permission from ODPM and BRE.
4 together. Unfortunately it is difficult to identify when an air spring is, or is not appropriate
for a particular plate–cavity–plate system. In addition it relies on a semi-empirical approach
to determine the coupling loss factor (Brekke, 1981; Craik and Wilson, 1995). These factors
make it rather awkward to use, hence it is not considered in the following sections.
4.3.5.1.1
Helmholtz resonators
To reduce the adverse effect of the mass–spring–mass resonance on the sound insulation,
attempts have been made to reduce the depth of the associated dip using arrays of Helmholtz
460
Chapter 4
Enclosed
volume, V
Cross-sectional
area of the neck, S
L
Figure 4.32
Helmholtz resonator.
resonators within the plate–cavity–plate system. These are tuned to resonate and absorb sound
energy at or near the mass–spring–mass resonance frequency. For windows with secondary
glazing, resonators can be placed around the perimeter of the cavity within the supporting
cavity wall to avoid obstructing the line of sight (Enger and Vigran, 1985). For cavity walls or
wall linings the resonators can be incorporated into one or both of the plates on the side that
faces into the cavity (Mason and Fahy, 1988; Warnock, 1993).
The cross-section through a typical Helmholtz resonator is shown in Fig. 4.32. At frequencies
where the wavelength is much larger than the dimensions of the resonator, the slug of air
contained within the neck acts as a lump mass and the larger volume of air in the cavity acts
as a spring. Hence its resonance frequency is given by the mass–spring resonance,
1
f =
2π
k
c0
=
m
2π
S
V (L + L)
(4.76)
where L is an end correction to define the effective mass of the slug of air (Ingard, 1953).
For a slit-shaped neck with rectangular cross-section, L ≈ 0.4 S 0.5 (Junger, 1975).
There are two main reasons why resonators are not widely used. Both theory and practice
show that if the resonators are tuned at or near the mass–spring–mass resonance frequency
it is possible to remove the dip in the sound insulation, or at least reduce its depth. However,
this is usually at the expense of slightly reducing the sound insulation at other frequencies.
There are also practical issues. If workmanship is poor, the use of resonators may introduce
air paths or structural coupling (i.e. bridging) which can negate any improvement in the sound
insulation at the mass–spring–mass resonance.
An alternative approach to try and reduce the mass–spring–mass resonance dip is through
active noise control (e.g. see Jakob and Möser, 2003; Kårekull, 2004). In practice, the cost
and complexity is often prohibitive for use in buildings.
4.3.5.2 Using the five-subsystem SEA model
This section shows how the five-subsystem SEA model can be used to help make design
decisions. Three examples are used to illustrate the basic principles and common assumptions
from which it should be clear how changes to the design of a building element can sometimes be
made on a qualitative basis without the need for calculations. Details concerning the calculation
of structural coupling, and sound transmission into and out of the cavity are included after the
examples so that the initial focus is on the application and interpretation of the SEA model.
461
S o u n d
I n s u l a t i o n
As with the three-subsystem SEA model for solid plates, we can calculate sound reduction
indices for the various transmission paths. There is one purely non-resonant path, 1 → 3 → 5,
for which the sound reduction index is
η 3 η 5 V5
S
R1→3→5 = 10 lg
(4.77)
+ 10 lg
η13 η35 V1
A
Resonant transmission paths include 1 → 2 → 3 → 4 → 5 and 1 → 2 → 4 → 5 for which path
analysis gives
S
η2 η3 η4 η5 V5
+ 10 lg
(4.78)
R1→2→3→4→5 = 10 lg
η12 η23 η34 η45 V1
A
and
R1→2→4→5 = 10 lg
η2 η4 η5 V5
η12 η24 η45 V1
+ 10 lg
S
A
(4.79)
When there is more than one type of structural connection between the two plates that does
not require modelling as a separate subsystem (e.g. wall ties and foundations), then Eq. 4.79
can be used for each connection separately or the individual structural coupling loss factors
can be summed to give η24 ; note that they need to be summed to calculate the general matrix
solution (Eq. 4.10).
The sound reduction index for other paths can be calculated in the same way, such as
those with a combination of resonant and non-resonant transmission, e.g. 1 → 2 → 3 → 5 and
1 → 3 → 4 → 5.
4.3.5.2.1
Windows: secondary glazing
This first example concerns two sheets of glass separated by a cavity depth of at least 50 mm.
The model could form the basis for an internal observation window between two rooms in a
building, or curtain walling, or domestic secondary glazing to improve the façade sound insulation of an external window. Section 4.3.12.3 will discuss the reasons why the five-subsystem
SEA model is not appropriate for modelling insulating glass units, i.e. sealed units with cavity depths typically less than 20 mm. This example uses a rather wide observation window
(3 × 1.25 m) with 6 or 10 mm glass on either side of a 50, 100, or 200 mm cavity; this is chosen
so that the cavity has sufficient local modes in the low-frequency range. Note that for secondary
glazing used with typical sizes of windows in the façades of dwellings, a lack of cavity modes
often restricts the SEA model to the mid- and high-frequency ranges. The mass–spring–mass
resonance is estimated to fall in or below the 100 Hz one-third-octave-band. Hence with only
one cavity mode below the 100 Hz band we will only model at and above the 125 Hz band.
Each pane of glass is usually supported on one leaf of a cavity wall. For specialist applications
such as a studio observation window, the cavity reveal lining will rarely be continuous across
the cavity. It is therefore assumed that transmission via the reveal lining or the supporting
cavity wall is insignificant so that the structural coupling path 1 → 2 → 4 → 5 can be ignored.
In practice it is the air paths around the frame and the structural paths due to bridging across
the reveal that limit the achievable sound insulation.
Figure 4.33 shows path analysis and the predicted sound reduction index using the matrix
solution. In this example the four transmission paths shown on the graph can be summed
using Eq. 4.19 to give approximately the same result as the matrix solution. Below the lower
of the two critical frequencies the non-resonant path 1 → 3 → 5 is generally the strongest, and
462
Chapter 4
90
1 →3→ 5
SEA transmission paths
Dotted lines ⫽ 2D sound field in cavity
Solid lines ⫽ 3D sound field in cavity
80
1 →3→ 4→5
1 →2→ 3→5
1 →2→ 3→4 →5
70
60
Sound reduction index (dB)
50
Subsystem 1: Source room
40
Subsystem 4:
10 mm glass
(fc ⫽ 1247 Hz,
ρs ⫽ 25 kg/m2)
30
80
Subsystem 3: 100 mm cavity
(f1,0,0 ⫽ 57 Hz,
f0,1,0 ⫽ 137 Hz,
f0,0,1 ⫽ 1715 Hz)
Subsystem 2:
6 mm glass
(fc ⫽ 2079 Hz,
ρs ⫽ 15 kg/m2)
Cavity reveal
70
Subsystem 5: Receive room
60
50
SEA matrix solution
2D sound field in cavity
3D sound field in cavity
40
fmsm ⫽ 62 Hz
30
50
80
125
200
315
500
800
1250
2000
3150
5000
One-third-octave-band centre frequency (Hz)
Figure 4.33
SEA model of a wide observation window with a 100 mm deep cavity between panes of 6 and 10 mm glass. Path analysis
is shown in the top part of the figure and the matrix solution in the lower part. Plate properties: Lx = 3 m, Ly = 1.25
m,
cL = 5200 m/s, ν = 0.24, radiation efficiency uses method No. 2 (lower limit only), total loss factor η = 0.024 + ηij ,
α = 0.02 (for the plate surfaces facing into the cavity at and above f0,0,1 in the cavity). Cavity properties: Empty cavity,
Lx = 3 m, Ly = 1.25 m. Reveal properties: ᾱP = 0.05 (125–400 Hz), ᾱP = 0.1 (500–5000 Hz).
the resonant path 1 → 2 → 3 → 4 → 5 is the weakest; but there is no dominant path. Above
the highest critical frequency of the plates there is only the resonant path 1 → 2 → 3 → 4 → 5.
As there is no structural coupling, all paths involve the cavity. The cavity sound field is two
dimensional below f0,0,1 and covers a large part of the building acoustics frequency range;
hence increasing the reveal absorption increases the sound reduction index. This can be seen
by the slight step in the curve that occurs between 400 and 500 Hz where the absorption
coefficient increases. It is not uncommon for reveal materials to have significantly different
463
S o u n d
I n s u l a t i o n
absorption coefficients in adjacent one-third-octave-bands and such steps may be more distinct
in practice. In this example the critical frequency for the 10 mm glass occurs when the cavity
sound field is two dimensional. Therefore increasing the reveal absorption can be used to
try and reduce the depth of the critical frequency dip. The critical frequency for the 6 mm
glass occurs when the sound field is three dimensional; hence the possibilities for significantly
changing the absorption are more limited because it is the two panes of glass that form the
largest surfaces in the cavity. Both critical frequencies are evident as dips in the resulting sound
reduction index.
Figure 4.33 shows that paths involving non-resonant transmission into and out of the cavity
play an important role in determining the overall sound insulation. For such constructions it is
important to note that there is more than one approach to predict the non-resonant coupling
loss factor between rooms and cavities. The calculations used to create Fig. 4.33 are based
on the approach of Price and Crocker (1970). Below the first cross-cavity mode this approach
shows no increase in the sound reduction index with increasing cavity depth; this is at odds
with experimental evidence (e.g. see Brekke, 1981; Mulholland, 1971; Nightingale, 1999). An
alternative SEA model to Price and Crocker for non-resonant transmission into and out of cavities is given by Craik (2003); this takes account of the cavity depth. Both of these approaches
are described in Section 4.3.5.3. Due to the importance of non-resonant transmission for this
particular plate–cavity–plate system we will now look at both of them.
Figure 4.34a uses the approach of Price and Crocker to compare the above result with different
windows that have 6 mm glass on each side of a 50, 100, or 200 mm cavity (i.e. 6–50–6, 6–
100–6, 6–200–6). Below the first cross-cavity mode the sound reduction index is the same for
the three different cavity depths with 6 mm glass on each side. At and above the first crosscavity mode they differ due to the three-dimensional nature of the sound field in the cavity; the
cavity total loss factor changes as well as the coupling loss factors into and out of the cavity.
Identical critical frequencies for both plates result in a deep dip in the sound reduction index at
the critical frequency. Using different glass thicknesses such as 6–100–10 gives two shallower
critical frequency dips. Compared to 6–100–6, the higher mass per unit area of the 10 mm
glass in 6–100–10 reduces the strength of the non-resonant transmission paths 1→3→5 and
1 → 2 → 3 → 5. This results in 6–100–10 having higher sound insulation than 6–100–6 below
the critical frequency dip of the 10 mm glass.
Figure 4.34b uses the approach of Price and Crocker at and above the first cross-cavity
resonance and the approach of Craik below the first cross-cavity resonance. For 6–50–6, 6–
100–6, and 6–200–6 in the low- and mid-frequency ranges (where non-resonant transmission is
dominant) there is a distinct increase in the sound reduction index with increasing cavity depth.
The significant differences between Fig. 4.34a and b below the first cross-cavity resonance
indicate the difficulty in making design decisions with an SEA model when non-resonant transmission is dominant. As yet there is insufficient evidence to choose one or the other approach for
all plate–cavity–plate systems. Fortunately, when there are dominant structural coupling paths
the uncertainty in predicting the non-resonant path becomes less important. The difficulty in
accurately predicting sound radiation at the critical frequency has already been noted with solid
plates. This problem is exacerbated in lightweight plate–cavity–plate systems where the first
cross-cavity resonance and the critical frequencies often occur within a few frequency bands
of each other. At these frequencies, accurately predicting differences due to different cavity
depths is not usually possible and reliance tends to be placed on laboratory measurements.
464
Chapter 4
(a) 80
Sound reduction index (dB)
70
6-50-6
fmsm ⫽ 98 Hz
6-100-6
fmsm ⫽ 69 Hz
6-200-6
fmsm ⫽ 49 Hz
6-100-10 fmsm ⫽ 62 Hz
60
50
40
30
Cavities
f1,0,0 ⫽ 57 Hz, f0,1,0 ⫽ 137 Hz
20
100 mm: f0,0,1 ⫽ 1715 Hz
6 mm glass (fc ⫽ 2079 Hz)
200 mm: f0,0,1 ⫽ 858 Hz
10 mm glass (fc ⫽ 1247 Hz)
50 mm: f0,0,1 ⫽ 3430 Hz
10
50
80
125
200
315
500
800
1250 2000
One-third-octave-band centre frequency (Hz)
3150
5000
(b) 80
Sound reduction index (dB)
70
fmsm ⫽ 98 Hz
6-100-6
fmsm ⫽ 69 Hz
6-200-6
fmsm ⫽ 49 Hz
6-100-10 fmsm ⫽ 62 Hz
60
50
6-50-6
Cavities
f1,0,0 = 57 Hz, f0,1,0 ⫽ 137 Hz
50 mm: f0,0,1 ⫽ 3430 Hz
100 mm: f0,0,1 ⫽ 1715 Hz
40
200 mm: f0,0,1 ⫽ 858 Hz
30
20
6 mm glass (fc ⫽ 2079 Hz)
10 mm glass (fc ⫽ 1247 Hz)
10
50
80
125
200
315
500
800
1250
2000
3150
5000
One-third-octave-band centre frequency (Hz)
Figure 4.34
SEA model (matrix solution) for different combinations of 6 and 10 mm glass and different cavity depths (50, 100, and 200 mm):
(a) using the approach of Price and Crocker to model non-resonant transmission into and out of the cavity and (b) using
the approach of Price and Crocker at and above the first cross-cavity resonance, and the approach of Craik below the first
cross-cavity resonance. Plate and cavity properties are given in Fig. 4.33.
465
S o u n d
I n s u l a t i o n
Because there are different ways of modelling non-resonant transmission, it is useful to briefly
discuss alternative models for plate–cavity–plate systems.
Alternative models (infinite plates): An alternative model for the plate–cavity–plate system
uses an impedance approach that accounts for the above-mentioned effect of cavity depth.
This is based on finite thickness plates and a finite depth cavity for a system of infinite extent
(Beranek and Work, 1949). The model works well for sound at normal incidence when the
cavity is partly or fully filled with a porous material. However, for sound transmission between
rooms we are more interested in diffuse sound fields in the source room. The impedance
model has been adapted to diffuse incidence for plate–cavity–plate systems with cavities that
are either empty, partly filled, or fully filled with a porous material (Fahy, 1985; Novak, 1992;
Ookura and Saito, 1978). Impedance models are ideally suited to cavities that are fully filled
with absorbent material; it is then reasonable to assume that any sound waves propagating
in the cavity (parallel to the plates) are so highly attenuated that there is no need to consider
cavity modes. For empty cavities at frequencies above the lowest cavity mode (which often
occurs in the low-frequency range) the errors tend to be larger than with fully filled cavities
because the cavity modes are ignored.
Most plate–cavity–plate systems in buildings have cavity depths between 50 and 300 mm; for
thinner cavities the plates are often tightly coupled together by the air in the cavity. For cavity
depths between 50 and 300 mm we now consider existing laboratory measurements on systems with empty cavities and negligible structural coupling. These indicate that in any individual
one-third-octave-band between the mass–spring–mass resonance and the first cross-cavity
mode, there is the potential to increase the sound insulation by up to 10 dB by at least doubling
the cavity depth as long as these cavity depths are within the range 50–300 mm (e.g. see
Brekke, 1981; Mulholland, 1971; Nightingale, 1999; Quirt, 1982; Utley and Mulholland, 1968).
For design purposes it is useful to identify such improvements. However, we need to consider
the accuracy of impedance models (diffuse incidence sound field) at frequencies between the
mass–spring–mass resonance and the first cross-cavity mode. Comparisons with laboratory
measurements indicate that differences in individual one-third-octave-bands between predictions and measurements are typically up to 10 dB for empty cavities or cavities partly filled with
fibrous materials, and up to 5 dB for cavities that are almost fully filled with fibrous materials
(Nightingale, 1999; Novak, 1992; Ookura and Saito, 1978). Hence although impedance models indicate the correct trend of increasing sound insulation with increasing cavity depth, the
error in predicting absolute values is similar to the potential improvement that is being sought.
Similarly we should consider the accuracy of SEA models below the first cross-cavity mode.
The original work by Price and Crocker (1970) was on a plate–cavity–plate system formed
from aluminium plates and an empty cavity without structural coupling; the difference between
one-third-octave-band SEA predictions and measurements of the sound reduction index was
generally less than 2 dB. Almost all validation of the theory on building elements tends to be on
plate–cavity–plate systems with structural coupling because most constructions require structural connections for stability. However, an indication of the accuracy can be found by looking
at the prediction of sound transmission from a room into an empty cavity below the first crosscavity mode where the cavity sound field is two dimensional. In such situations the differences
between predictions and measurements for the cavity sound pressure level are typically up
to 5 dB (Craik and Smith, 2000a; Hopkins, 1997; Wilson, 1992). The existing evidence does
not seem to indicate that one model (SEA or impedance approach) will always be significantly
‘better’ than the other.
466
Chapter 4
Neither SEA nor impedance models are perfectly suited to predicting the change in sound
insulation due to increasing the cavity depth below the first cross-cavity mode. It is fortunate that
structurally isolated plates forming cavity walls and floors are not common in buildings. For most
plate–cavity–plate systems there will be structural coupling as well as flanking transmission,
therefore errors in the non-resonant path of an SEA model often become less important due to
the existence of many other transmission paths. To accurately quantify the effect of changing
the cavity depth it is often necessary to use carefully designed laboratory measurements that
avoid confounding factors such as niche effects, unwanted structural coupling, and changes
in the total loss factor of the plates at different positions in the test aperture.
4.3.5.2.2 Masonry cavity wall
In this example a masonry cavity wall is modelled in three different scenarios. Although it is
possible to create an SEA model for this wall in a specific laboratory or in a specific building
we will deliberately chose an example that is a step removed from the reality in situ and in
the laboratory. This allows us to look at features that are important to both. (Note that a
comparison of SEA with measurements on a masonry cavity wall construction will be shown
in Section 5.3.2.2 for the combination of direct and flanking transmission.)
Scenario (A) has no structural connections between the plates so that all sound transmission
occurs via the cavity, scenario (B) has wall ties connecting the plates, and scenario (C) has wall
ties and a foundation connecting the plates. As with SEA modelling of solid plates in Section
4.3.1.3, we use the total loss factor to model each plate as if it were connected to other walls
and floors without actually including them in the model. This is done by treating the sum of these
coupling loss factors (estimated to be 0.3f −0.5 ) as an internal loss factor. When modelling a
separating wall in situ where the connected flanking walls and floors are included in the model
there is no need to make this adjustment. Non-resonant transmission into and out of the cavity
is modelled using the approach of Price and Crocker.
The predicted sound reduction indices using the SEA matrix solution are shown in Fig. 4.35a
for scenarios A, B, and C. Path analysis is shown for scenarios A and C to help assess the
strength of different paths by comparing them with the matrix solution in Fig. 4.35b.
Scenario A has a mass–spring–mass resonance frequency that is well-below 50 Hz and outside
of the building acoustics frequency range. It is common to design masonry cavity walls to have
resonance frequencies below the frequency range of interest. This is done to reap the benefits
of higher sound insulation due to the separation between the wall leaves; we recall that below
the resonance frequency a cavity wall effectively acts as a single solid wall. As all transmission
paths involve the cavity, the sound reduction index will change when the cavity total loss factor
is changed; hence the addition of absorption in the cavity will increase the sound reduction
index (and vice versa). Below the first cross-cavity mode the sound field in the cavity is two
dimensional so absorption could either be placed around the perimeter of the cavity, or the
cavity could be partially or fully filled with absorbent material across its depth. In a transmission
suite the absorption around the cavity perimeter may vary between laboratories unless specific
material is introduced around the perimeter as part of the test element.
Scenario A gives an unrealistically high sound reduction index because there are no structural
connections or flanking transmission. Note that above the critical frequency, the matrix solution
gives the same sound reduction index as path 1 → 2 → 3 → 4 → 5.
467
S o u n d
I n s u l a t i o n
(a)
120
A
B
fmsm ⫽ 26 Hz
fmsm ⫽ 50 Hz
C
110
Sound reduction index (dB)
100
90
fmsm ⫽ 50 Hz
80
70
First cross-cavity
mode
f0,0,1 ⫽ 2287 Hz
60
Plates
f11 ⫽ 22 Hz
fc ⫽ 295 Hz
fB(thin) ⫽ 1204 Hz
50
40
30
50
80
125
200
315
500
800
1250 2000
One-third-octave-band centre frequency (Hz)
3150
5000
(b)
120
110
100
Sound reduction index (dB)
120
1→3→5
1→2→3→5
1→3→4→5
1→2→3→4→5
Matrix solution
100
90
90
80
80
70
1→3→5
1→2→3→5
1→3→4→5
1→2→4→5 (Wall ties)
1→2→4→5 (Foundation)
1→2→3→4→5
Matrix solution
110
A
60
60
50
50
40
fmsm ⫽ 26 Hz
30
C
70
40
fmsm ⫽ 50 Hz
30
50
80
125 200 315 500 800 1250 2000 3150 5000
One-third-octave-band centre frequency (Hz)
50
80
125 200 315 500 800 1250 2000 3150 5000
One-third-octave-band centre frequency (Hz)
Figure 4.35
SEA model of a masonry cavity wall: (a) matrix solution and (b) path analysis for comparison with the matrix solution. Plate
properties: Lx = 4 m, Ly = 2.5 m, h = 0.1
m, ρs = 140 kg/m2 , cL = 2200 m/s, ν = 0.2, radiation efficiency uses method No. 3,
total loss factor η = 0.01 + 0.3 f −0.5 + ηij (Note that 0.3f −0.5 is being used here to represent the sum of the structural
coupling losses to connected walls and floors that are not included in the model.). Cavity properties: Empty cavity, Lx = 4 m,
Ly = 2.5 m, Lz = 0.075 m, T = 0.3 s. Wall ties: 2.5 ties/m2 , k = s75 mm = 2 × 106 N/m. Foundation: d1 = 0.25 m, d2 = 0.6 m,
ρ = 2000 kg/m3 , s′soil = 1.96 × 109 N/m3 , ηsoil = 0.96.
468
Chapter 4
The addition of wall ties gives scenario B where the sound reduction index is now significantly
reduced in the low-frequency range. The combined mass–spring–mass resonance frequency
of the air spring and the wall ties is 50 Hz so in practice there could be a slight dip in the sound
insulation at this frequency. This mass–spring–mass transmission mechanism is not included
in the SEA model; if important, it is easiest to include it as an empirical adjustment.
Scenario C is created by adding a common foundation (concrete) where the ground is stiff
clay with stones. This reduces the sound reduction index across the entire building acoustics frequency range. Note that this foundation detail has deliberately been chosen to give
strong foundation coupling so that scenarios A and C are indicative of the extremes. Different
foundation details will be discussed in Section 4.3.5.4.3 and the wave theory used in these
calculations will be given in Section 5.2.4.
Measurements on masonry cavity walls in the laboratory are not easy to interpret in terms of
their performance in situ. We recall that with solid homogeneous walls we could convert a result
from one laboratory to another laboratory by using the measured total loss factor (refer back
to Section 3.5.1.3.2). With plate–cavity–plate systems there is more than one path involving
resonant transmission; hence there are no simple conversions. The engineer is caught between
a rock and a hard place; laboratory measurements are important because ‘perfect’ theoretical
models do not exist, yet it is difficult to apply the laboratory measurement without the aid of
a theoretical model. Various tactics have been used to try and overcome this difficulty. One
possibility is to establish rule-of-thumb conversions from a specific mounting condition in the
laboratory to a specific situation in buildings. This may be possible when masonry cavity walls
are not rigidly connected to the foundations; they are sometimes built off resilient materials
so the transmission path via the foundations may not be as important. In addition, structural
coupling via the foundations may dominate on the ground floor of a multi-storey building but
not several floors above it. In some transmission suites it is possible to build foundations below
the aperture. Otherwise, if the aperture is sufficiently high a foundation can be built within
the aperture and shielded with linings (Parmanen et al., 1988). Another approach is to use
a flanking laboratory to test the combination of the separating cavity wall and some of the
flanking walls and floors to try and simulate the actual building. Note that transmission via
the foundations can be affected by the underlying soil; hence there are limitations to building
representative foundations in a transmission suite or a flanking laboratory.
4.3.5.2.3 Timber joist floor
This example looks at sound transmission through a basic timber joist floor. This commonly
forms an internal floor within a dwelling, and its sound insulation often requires upgrading in a
building that is being converted into flats. In this case an SEA model is used to form a basis
upon which decisions can be made on how to improve the sound insulation.
Figure 4.36 shows the timber joist floor along with path analysis and the predicted sound
reduction index using the matrix solution. The five-subsystem model has been used with a
single subsystem to model transmission via all the floor cavities (subsystem 3) and an extra
subsystem has been used to model structural coupling via all the joists (subsystem 6). The
structural coupling between the chipboard/plasterboard and the joists is modelled by assuming
point connections at screws/nails. Note that when modelling structure-borne sound excitation
of lightweight walls and floors it may be necessary to model every cavity and every joist as an
individual subsystem.
469
S o u n d
I n s u l a t i o n
90
SEA transmission paths
Dotted lines ⫽ 1D sound field in cavity
Solid lines ⫽ 2D/3D sound field in cavity
80
1→3→5
ⴙ 1→3→4→5
ⴛ 1→2→3→5
70
1→2→3→4→5
1→2→6→4→5
60
50
40
Sound reduction index (dB)
30
20
Subsystem 1: Source room
10
Subsystem 2: Chipboard
Subsystem 6:
Joist
Subsystem 3:
Cavity
0
Subsystem 4: Plasterboard
60
Subsystem 5: Receive room
50
40
30
Measured
20
SEA matrix solution
(1D sound field in cavity)
10
SEA matrix solution
(2D/3D sound field in cavity)
0
50
80
125
200
315
500
800
1250 2000
One-third-octave-band centre frequency (Hz)
3150
5000
Figure 4.36
SEA model of a timber joist floor (matrix solution and path analysis). Plate properties: Lx = 5.45 m, Ly = 4.15 m, radiation
efficiency uses method No. 2 (lower limit only). Chipboard: h = 0.025 m, ρs = 19 kg/m2 , cL = 2210 m/s,ν = 0.3. Plasterboard:
h = 0.0125 m, ρs = 9.7 kg/m2 , cL = 1910 m/s, ν = 0.3. Beam properties: 225 mm deep timber joists spaced at 450 mm
centres with screw/nail fixings at ≈ 500 mm centres (modelled as point connections). Lx = 4.15 m, hy = 0.05 m, hz = 0.225 m.
Fundamental beam mode, f1 = 31 Hz. Cavity properties: Empty cavity, Lx = 4.15 m, Ly = 0.4 m, Lz = 0.225 m. Cavity modes
that demarcate one-, two-, and three-dimensional sound fields are f1,0,0 = 41 Hz, f0,1,0 = 429 Hz, f0,0,1 = 762 Hz. Mass–spring–
mass resonance frequency: 50 Hz. Total loss factors are predicted from the sum of the coupling loss factors and the internal
loss factor. Measured data from Hopkins are reproduced with permission from ODPM and BRE.
470
Chapter 4
The mass–spring–mass resonance frequency is 50 Hz and the lowest cavity mode is 41 Hz,
hence the plate–cavity–plate system is used across the entire building acoustics frequency
range. Above 41 Hz the cavity has a one-dimensional sound field which changes to a twodimensional field at 429 Hz, and to a three-dimensional field at 762 Hz. Although each element
of the floor can support local modes, one-dimensional subsystems such as the cavity (below
the 400 Hz band) and the beam have quite low mode counts; we can refer back to the examples
for a similar cavity and joist in Figs 1.55 and 2.26 respectively. The lower part of Fig. 4.36 shows
measured data for comparison with the SEA matrix solution. For the one-dimensional sound
field, non-resonant transmission into and out of the cavity is predicted using the approach of
Craik. For the two- and three-dimensional sound fields it is modelled using the approach of Price
and Crocker. These approaches are described in Section 4.3.5.3. An abrupt transition between
models is not usually appropriate, so the overlap between them is shown. The agreement
between the measured and predicted sound insulation indicates that it is now appropriate to
draw conclusions from the path analysis shown in the upper part of Fig. 4.36.
In the low- and mid-frequency range, non-resonant transmission via the path 1 → 3 → 5 tends
to dominate. Therefore increasing the airborne sound insulation would require increasing the
mass per unit area of the chipboard or the plasterboard, and/or putting highly absorptive material in the cavity. However, the cavity dimensions also play a role because the sound field is one
dimensional and there are relatively few modes. For this reason the low-frequency sound insulation in individual frequency bands can change significantly with different joist depth, spacing,
and length; these vary depending on the structural requirements. The model is too crude to
predict these changes accurately, but they can be observed in laboratory measurements on
timber floors. Note that timber frame walls usually have much shallower cavities than these
floors and the mass–spring–mass resonance often plays an important role in the low-frequency
range.
In the high-frequency range, paths 1 → 2 → 3 → 5, 1 → 2 → 6 → 4 → 5 and 1 → 2 → 3 → 4 → 5
mainly determine the sound insulation. Sound transmission via paths that involve the sound
field in the cavity could be reduced with absorption in the cavity. However, structural coupling
via the joists along path 1 → 2 → 6 → 4 → 5 will limit the maximum achievable sound insulation.
Hence separating timber floors between dwellings also tend to require resilient devices such as
hangers, bars, or channels to isolate the plasterboard from the joists thus reducing transmission
via this structural path.
4.3.5.3 Sound transmission into and out of cavities
Resonant and non-resonant coupling loss factors for plates and rooms were introduced in
Section 4.3.1. For sound radiation from a plate into a space, use of the radiation efficiency
equations assumes that the plate is radiating into a free field. This free-field assumption also
applies to non-resonant transmission between two spaces. For rooms these assumptions are
appropriate. However, in comparison with rooms, cavity depths are very shallow. It is therefore questionable whether the free-field assumption is appropriate for cavities. For resonant
transmission, comparisons of measured and predicted data indicate that the assumption is
reasonable for all types of plates (e.g. see Craik and Smith, 2000a; Price and Crocker, 1970;
Wilson, 1992). Some experiments also indicate that it is reasonable for non-resonant transmission across a plate into a cavity when the other plate that forms the cavity is sufficiently rigid
that there is negligible interaction with the sound field in the cavity (Craik and Smith, 2000a).
Examples of this include cavities between plasterboard and non-porous masonry/concrete
471
S o u n d
I n s u l a t i o n
plates. Note that for plate–cavity–plate systems where the resonant transmission path
1 → 2 → structural coupling → 4 → 5 is dominant, accurate modelling of transmission into and
out of the cavity becomes less important. For lightweight cavity walls and floors, where neither
plate can be considered rigid, an alternative approach can be used to predict non-resonant
transmission from the cavity to the room and vice versa (Craik, 2003).
The approach that was originally used by Price and Crocker (1970) is described now because
it is still a useful starting point for many models. For resonant transmission, the coupling loss
factor from the plate to the cavity is calculated using Eq. 4.21. The coupling loss factor from
the cavity to the plate can then be calculated with the consistency relationship (Eq. 4.2) by
using the modal density of the cavity. For non-resonant transmission, the coupling loss factor
is calculated from the room to the cavity using Eq. 4.26, and the consistency relationship is
used for the reverse direction.
Cavities often have one- or two-dimensional sound fields in the low- and mid-frequency ranges
where all wave motion is parallel to the plate surfaces. From Craik (2003), an alternative model
to that of Price and Crocker (1970) can be used to predict non-resonant transmission via a plate
that connects a cavity (i) to a room (j) below the first cross-cavity mode. This model assumes
that the sound pressure associated with waves in the cavity (propagating parallel to the plate)
cause the plate to move with a lateral velocity that is only determined by its mass per unit area,
hence
v̂ 2 =
p̂i2
ω2 ρs2
(4.80)
The phase velocity of the structural wave induced in the plate must follow that of the sound
wave in the cavity, so from Eq. 2.198 the radiated power from the plate can be given as
3 4
Wij = σfc Sρ0 c0 v 2 t,s
(4.81)
where S is the surface area of the plate that faces into the cavity and σfc is the radiation
efficiency of the plate at the critical frequency (i.e. the frequency at which the phase velocities
for sound in the cavity and the structural wave in the plate are equal).
Equations 4.3, 4.80, and 4.81 can now be used to determine the coupling loss factor from the
cavity to the room using
Wij = ωηij Ei =
σfc Sρ0 c0 3 2 4
pi t,s
ω2 ρs2
(4.82)
hence, the coupling loss factor at frequencies below the first cross-cavity mode is given by
(Craik, 2003)
ηij =
ρ02 c03 Sσfc
ω3 Vi ρs2
(4.83)
Note: This assumes that ρ0 and c0 are the same in the cavity and the room.
The coupling loss factor for non-resonant transmission from the room to the cavity can be
calculated from the consistency relationship. To cover the entire building acoustics frequency
range it will sometimes be necessary to switch over to the Price and Crocker approach in the
vicinity of the first cross-cavity mode.
472
Chapter 4
4.3.5.4
Structural coupling
There is always more than one way to model structural coupling between two parts of a structure; and in the absence of a model there is usually more than one way to measure it. In any
assessment of a plate–cavity–plate structure it is useful to consider two aspects of structural
coupling separately. The first aspect concerns connections between two plates over their surface such as a framework, resilient mounts, or wall ties that connect the two plates together
(i.e. connections that would exist whether the plate was modelled as being of finite or infinite
size). The second aspect concerns the connections around the boundaries of a finite plate;
these may be different in the laboratory and in situ.
4.3.5.4.1 Point connections between plates and/or beams
Point connections between parallel plates, between plates and beams, and between beams
are very common. Plates can be connected together by wall ties, dabs of adhesive, resilient
mounts, or bridged screeds. Plates and beams are often connected with screws, nails, or bolts.
We start by modelling a single point connection. There is usually more than one connection so
we will need to look at the assumptions made when using this model for many point connections.
When point connections are closely spaced along a line then the appropriateness of the point
connection model depends on the wavelength of the structure-borne sound waves.
Simple models to quantify the power flow across a single point connection make use of the
driving-point mobility (or impedance) of the plates or beams, and the mobility (or impedance)
of the point connection (Cremer et al., 1973; Lyon and DeJong, 1995). These are conveniently
calculated using equivalent circuits based on impedance or mobility (e.g. see Harris and Crede,
1976). The following approach is general and has been used for resilient mounts under floating
floors (Vér, 1971), wall ties between masonry cavity walls (Craik and Wilson, 1995; Narang,
1994) and screw/nail connections between plasterboard and a timber frame (Craik and Smith,
2000b).
We assume that there is a propagating bending wave on a beam or plate (denoted as subsystem i) as shown in Fig. 4.37. This is coupled to another beam or plate (subsystem j) by a
single point connection. The propagating bending wave velocity far from the point connection
on plate i is v0 , but the presence of the point connection changes the velocity directly above
this connection to vi . This results in a velocity, vj , directly above the connection on the other
beam or plate. The axial force, F, on the connection is proportional to vi − vj . To determine
the power transmitted by the point connection, the equivalent circuit in Fig. 4.37 yields the
following relationships for F and vj ,
F=
v0
Yi + Yj + Yc
vj = FY j =
v0 Yj
Yi + Yj + Yc
(4.84)
where Yi and Yj are the driving-point mobilities and Yc is the mobility of the point connector.
The transmitted power is then given by
Wij =
< v2
;
Re Yj
1
Re Fvj∗ = 0
2
2 Yi + Yj + Yc 2
(4.85)
473
S o u n d
I n s u l a t i o n
Yj
Yi
F
Yi
vi
Yc
vi
vj
v0
Yc
vj
Yj
v0
Figure 4.37
Point connection between two subsystems i and j and the equivalent circuit. Each subsystem could be a plate or a beam. The
illustration shows a distinct gap between the two structures; this would apply to wall ties in a cavity wall, or a resilient mount
between a floating floor and a base floor. However the model can also be used for screw, nail, or bolt connections between
plates and/or beams for which this gap will be very small.
which can be written in terms of the coupling loss factor as
Wij = ωηij Ei = ωηij mi
v02
2
(4.86)
where mi is the mass of subsystem i.
For N identical point connections, equating Eqs 4.85 and 4.86 gives the coupling loss factor as
Re Yj
N
ηij =
(4.87)
ωmi Yi + Yj + Yc 2
To calculate the coupling loss factor, the beams and plates can be modelled as being of infinite
extent; this allows the impedances in Section 2.8.3 to be used to calculate the driving-point
mobilities.
This approach assumes that the power transmitted across the connection is only due to forces
and that any power transmitted by moments is negligible. It also assumes that each connection
acts independently of the others; hence there must be no correlation in the vibration field above
the connection points. Correlation could occur with an evenly spaced array of point connections.
However, even if the intention is to construct a wall or floor with a perfectly periodic array of
point connections, very few walls and floors are built so exactly and have spatially uniform
material properties; therefore the assumption of uncorrelated points is often reasonable.
For a resilient point connection such as a resilient mount or a wall tie, it may be appropriate to
model it as a simple linear spring for which
Yc =
iω
k
(4.88)
where k is the dynamic stiffness of the point connection acting as a spring (N/m). Note that for
wall ties, k is identical to sX mm as described in Section 3.11.3.2.
474
Chapter 4
For rigid point connections such as screws, nails, or bolts, the stiffness can be assumed to be
infinite and Yc = 0.
For plates connected by springs across an air-filled cavity (such as cavity walls with ties, or floating floors on resilient mounts) a mass–spring–mass resonance frequency can be calculated
for the air stiffness and the spring stiffness acting in parallel,
!
"
1 " sa′ + NS k
"
fmsm =
(4.89)
ρs1 ρs2
2π #
ρs1 + ρs2
where S is the plate area.
Below this mass–spring–mass resonance frequency the two plates effectively act as a single
plate with the combined mass per unit area and thickness of the two individual plates.
Real point connections are not always simple: springs may be non-linear, structures can be
offset from each other at the connection point, the effective contact area of point connections varies depending on the structural wavelength, and the connection may support wave
motion across its length at high frequencies. These complexities become apparent in laboratory
studies on isolated structures and result in more complex models or empirical solutions for specific connectors. In the absence of such information, the simplicity of the impedance/mobility
approach is ideal for initial estimates.
Full-size walls and floors that are only rigidly or resiliently connected along their edges at
equally spaced points to form L-, T-, and X-junctions are less common in buildings. For these
junctions, other models are available that can be used to determine the coupling loss factors
(Bosmans, 1998; Bosmans and Vermeir, 1997).
For a beam that is connected to a plate using screws, nails, or bolts the point connection model
is appropriate when the structural wavelength is much smaller than the distance between the
connections. When the wavelength is much larger than this distance, the junction between the
beam and plate acts as a line connection. As a rule of thumb for bending waves on lightweight
timber frame walls and floors, the transition from a line connection to individual point connections can be assumed to start when the screw/nail spacing is approximately equal to half a
bending wavelength (Craik and Smith, 2000b). Hence the line connection model is sometimes
needed in the low-frequency range with a changeover to the point connection model at higher
frequencies. This simplification does not deal with the effective contact length of point connections which varies with frequency and makes it difficult to identify a transition frequency
without using more complex models (Bosmans and Nightingale, 2001). When there is uncertainty in whether to assume line or point connections, it is simplest to model them both to give
an indication of the potential range.
4.3.5.4.2
Line connections
Line connections occur between plates, and between plates and beams. Examples include
lightweight walls and floors with a framework, and masonry/concrete cavity walls where there
is a return at a wall boundary. Line junctions occur with lightweight cavity walls and floors
where the plates on either side of the wall are connected to the same beam (see Fig. 4.38).
For a line of closely spaced point connections (e.g. screws, bolts), the line connection model
is appropriate when the spacing is much smaller than the bending wavelength (as discussed
in Section 4.3.5.4.1).
475
S o u n d
I n s u l a t i o n
Figure 4.38
Examples of line connections between plates in a plate–cavity–plate system.
(a)
(b)
d1
d2
Figure 4.39
Examples of different foundation details for masonry/concrete cavity walls. The ground floor is not shown; this could be built
in to each cavity leaf and either lie on the ground or be suspended above it. (a) Concrete deep trench fill: Trench is filled to
within a few tens of centimetres of the ground surface. The trench may be up to 3.5 m below ground level. (b) Strip footing:
Wall extends below the damp proof course to rest upon a strip concrete footing (e.g. d1 ≈ 0.25 m, d2 ≈ 0.6 m). Below ground
level there are a number of different options: thicker or different blocks, wall ties, poured concrete infill, and thermal insulation
within the cavity.
For line connections between parallel plates it is also possible to use an equivalent circuit
approach in the same way as with point connections (Sharp, 1978). However it is usually
necessary to adopt a wave approach. For line connections such as those shown in Fig. 4.38, a
comprehensive overview for bending wave transmission along with tabulated values are given
in the book by Craik (1996). It is sometimes necessary to consider both bending and in-plane
waves for which other models are available (Craik and Smith, 2000b; Langley and Heron,
1990). Vibration transmission between connected plates is discussed in Section 5.2.
4.3.5.4.3
Masonry/concrete walls: foundations
Structural coupling via the foundations is often an important transmission path at the boundary
of a masonry/concrete cavity wall. Different foundation details are often used for the same
cavity wall depending on the soil conditions and the structural requirements; a few different
foundation details are shown in Fig. 4.39. The variety and complexity of foundation details has
implications for the interpretation of laboratory measurements on cavity walls.
To include structural coupling via the foundations into the five-subsystem SEA model, it is
possible to rely purely on measurements. Structural intensity or vibration level differences can
be used to determine the coupling loss factor as described in Section 3.12.3.3. Note that it is
not always possible to use vibration level differences if there are other significant transmission
paths such as those involving the cavity or wall ties. By measuring the structural coupling, the
foundation detail is effectively treated as a black box; bending waves enter the black box on
one leaf, and emerge on the other leaf of the cavity wall. Example measurements are shown
476
Chapter 4
Masonry wall leaves
Lx ⫽ 4 m, Ly ⫽ 2.5 m, h ⫽ 0.1 m
ρs ≈ 150 kg/m2
cL ≈ 2500 m/s
Coupling loss factor (dB)
120
110
Foundation detail
0.15 m concrete slab on
≈ 0.7 m bricks on ≈ 1m concrete
100
90
80
70
100
125
160
200
250
315
400
500
630
One-third-octave-band centre frequency (Hz)
800
1000
Figure 4.40
Measured coupling loss factors for vibration transmission via a split foundation (two separate foundations separated by a
50 mm gap), and on one side of the split foundation. Measurements were taken using structural intensity and are shown with
95% confidence intervals. Note that the coupling loss factor depends upon the junction length with the foundation and the
plate area. Measured data from Hopkins are reproduced with permission from ODPM, BRE, and BRE Trust.
in Fig. 4.40 for a masonry cavity wall placed in two different positions on a split foundation
in a flanking laboratory. Whilst measurements are essential for quantifying the performance
of complicated foundations, reliance on measured data is not particularly desirable as there
are many possible foundation details and many different types of soil. To give an indication of
whether different soils are likely to be a significant factor it is necessary to use a theoretical
model for vibration transmission via the foundation.
Complete models of many foundation details are made complicated by the fact that they are
built from several components that don’t form simple beams or plates, and that are highly
damped by the soil. In addition, these components are not all rigidly bonded together because
they must also provide thermal insulation and prevent ingress of moisture. However, a model
for bending wave transmission via a simplified strip foundation is useful in highlighting the
role of soil stiffness in this structural transmission path (Wilson and Craik, 1995). This model
is described in Section 5.2.4 from which Fig. 4.41 shows the calculated coupling loss factors for three different soils (soil properties are given in Table 5.1). The peaks are due
to the soil being modelled as a lump spring; in practice the soil is likely to be anisotropic
so the coupling loss factors should be used to identify general trends rather than specific
resonance peaks. The measured and predicted coupling in this section indicates that for
masonry cavity walls there can be differences up to 10 dB with different foundation details.
Earlier analysis in Section 4.3.5.2.2 indicates that the transmission path via the foundations
477
S o u n d
I n s u l a t i o n
Masonry wall leaves
Lx ⫽ 4 m, Ly ⫽ 2.5 m, h ⫽ 0.1 m
rs ⫽ 140 kg/m2
cL ⫽ 2200 m/s
Coupling loss factor (dB)
120
Stiff clay with large
stones (0.45 m)
Lower green sand
overlying
sandstone (0.9 m)
Strip foundation
d1 ⫽ 0.25 m, d2 ⫽ 0.6 m
r ⫽ 2000 kg/m3
110
London clay – wet
(1.5 m)
100
90
80
70
50
80
125
200
315
500
800
1250
2000
3150
5000
One-third-octave-band centre frequency (Hz)
Figure 4.41
Predicted coupling loss factors for vibration transmission via a strip foundation with different soils.
is sufficiently strong that different foundation details and different soils are likely to account
for some of the variation in airborne sound insulation between nominally identical cavity
walls in situ.
4.3.5.4.4
Lightweight cavity walls
For many lightweight cavity walls, the complexity in predicting the structural coupling between
the two wall leaves means that reliance is generally placed upon extensive laboratory testing (e.g. see Walker, 1993). As with laboratory measurements of masonry cavity walls, it is
necessary to consider the absence or inclusion of structural coupling around the perimeter
of a lightweight cavity wall. For laboratory measurements, the relevant Standard (ISO 140
Part 1) gives requirements on the laboratory structure that forms the test aperture. Previously
there had been no specific requirements for the dimensions and properties of the materials
that form the test aperture when testing lightweight cavity walls; during this time an interlaboratory comparison of airborne sound insulation was carried out on such walls (Fausti et al.,
1999). A sample of the results are shown in Fig. 4.42 (Smith et al., 1999). These indicate
that structural coupling around the perimeter can significantly alter the airborne sound insulation over the majority of the building acoustics frequency range. It is worth noting that it
is generally easier to achieve a high degree of isolation between wall leaves in the laboratory than in situ. Rigid connections between wall leaves around the perimeter of a wall or
floor in situ may be necessary for fire stopping; these can significantly reduce the sound
insulation unless materials are carefully chosen to minimize the structural coupling (Craik
et al., 1997).
478
Chapter 4
80
Transmission suite aperture
border material:
70
Sound reduction index (dB)
Concrete
60
Steel lining
Timber lining
50
40
12.5 mm plasterboard
Metal C-studs
120 mm cavity containing 50 mm glass wool
Metal C-studs
12.5 mm plasterboard
30
20
10
50
80
125
200
315
500
800
1250
2000
One-third-octave-band centre frequency (Hz)
3150
5000
Figure 4.42
Airborne sound insulation of a lightweight cavity wall measured in laboratories with different border materials around the
perimeter of the aperture. Measurements according to ISO 140 Part 3. Measured data are reproduced with permission from
Smith et al. (1999).
s⬘a1
rs1
s⬘a2
rs2
rs3
Figure 4.43
Mass–spring–mass–spring–mass system.
4.3.5.5 Plate–cavity–plate–cavity–plate systems
For a plate–cavity–plate–cavity–plate system modelled with lump masses and springs as
shown in Fig. 4.43, there are two resonance frequencies to consider (Blevins, 1979)
fmsmsm =
where
X =
1
23/2 π
!
"
"
#X ± X 2 − 4s′ s′
a1 a2
1
1
1
+
+
ρs1 ρs2
ρs2 ρs3
ρs1 ρs3
′
′
+ sa2
s′
s′
sa1
+ a1 + a2
ρs2
ρs1
ρs3
(4.90)
where the dynamic stiffness per unit area for each air-filled cavity is given by Eq. 4.72.
The lower resonance frequency is calculated using the negative sign in Eq. 4.90. Note that the
frequency range between the two resonances can span one or several one-third-octave-bands.
479
S o u n d
I n s u l a t i o n
h1
h2
(1) Plate
h3
(2) Core
(3) Plate
Figure 4.44
Example sandwich panels.
The SEA model for a plate–cavity–plate system can be extended to a plate–cavity–plate–
cavity–plate system. An SEA model is appropriate above the higher of the two resonance
frequencies. One application is triple glazing in studios which often has individual cavity depths
of at least 100 mm; hence the resonance frequencies tend to be in or below the low-frequency
range. With increasing frequency, transmission paths involving structural coupling between
the plates will start to dominate over paths involving the sound fields in the cavities.
For three closely spaced sheets of glass it is only possible to draw limited conclusions from
laboratory measurements in which the individual cavity depths are at most 20 mm and the
lowest resonance falls within the frequency range of interest. These only tend to show a discernible dip in the sound insulation at the lower resonance frequency and generally offer small
improvements over what can be achieved with double glazing (Brekke, 1981; Quirt, 1983).
However, it has been shown that the sound insulation between the resonance frequencies can
be higher when the ratio of the lower to the higher resonance is much less than unity and the
layout of the plates and cavities is not symmetrical about the central plate (Vinokur, 1990).
4.3.6 Sandwich panels
Sandwich panels are sometimes used to form internal walls, external walls, doors, or roof
elements. They usually consist of a relatively lightweight core material with a plate (e.g. plasterboard, cement particle board, steel) bonded to this core on each side (see Fig. 4.44).
A variety of core materials are used such as foam, mineral wool, or a cardboard honeycomb.
Initial design considerations are usually given to the structural and/or thermal performance so
the properties of the core may end up being isotropic, orthotropic, or anisotropic, with or without
rigid connections between the two plates across the core.
Early work by Kurtze and Watters (1959) was concerned with identifying a type of plate that
would have a high static stiffness and low weight, but without the high bending stiffness that
causes resonant transmission to become as dominant as it does with solid plates. Their sandwich panel design was based on ‘promoting’ transverse shear motion rather than bending wave
motion. This required a core with low shear stiffness, and thin and relatively stiff plates on either
side. Wave motion on the panel was described by combining the panel bending stiffness and
shear stiffness in parallel. The resulting motion is advantageous because of the non-dispersive
nature of transverse shear waves. It extends the frequency range over which the phase velocity
for wave motion on the panel is lower than the phase velocity in air. Therefore by moving the
lowest coincidence frequency to higher frequencies, the panel can act as a limp mass over
a larger part of the frequency range. Later work identified adverse resonance effects relating
to thickness deformation of the core (Ford et al., 1967). In practice this limits the ability to
480
Chapter 4
Symmetric wave motion
Plate
Core
Plate
Antisymmetric wave motion
Plate
Core
Plate
Figure 4.45
Sandwich panel – wave motion.
achieve the non-resonant (mass law) sound reduction index at all frequencies below coincidence. Sandwich panels are commonly used in automotive or aeronautic industries because
their acoustic performance can usually be optimized for a specific noise source (e.g. see Makris
et al., 1986; Moore and Lyon, 1991; Wen-chao and Chung-fai, 1998).
For porous core materials the sound reduction index of a sandwich panel can be calculated
using infinite plate assumptions and Biot theory to model two compressional waves and a shear
wave propagating in the core (Bolton et al., 1996; Lauriks et al., 1992).
For a non-porous core with propagating shear and dilatational waves, resonant transmission
across a sandwich panel can be modelled in terms of symmetric and antisymmetric wave motion
on the panel (Dym et al., 1974, 1976; Moore and Lyon, 1991). Symmetric and antisymmetric
motions, correspond to dilatational and bending wave motion respectively (see Fig. 4.45).
Symmetric motion involves thickness deformation of the core. This gives rise to a dilatational
resonance that usually causes an adverse dip in the sound reduction index. For a sandwich
panel where both plates are identical, the dilatational resonance frequency, fd , is given by
(Moore and Lyon, 1991)
!
"
" 2(λ2 + 2μ2 )
"
1 "
h2
"
fd =
(4.91)
ρ2 h2
2π #
ρ1 h 1 +
6
where subscript 1 refers to either plate (assuming h1 = h3 and ρ1 = ρ3 ), and subscript 2 refers
to the core for which μ2 and λ2 are Lamé’s constants given by Eqs 4.47 and 4.48.
Reasonable estimates for the dilatational resonance frequency can also be determined from
the measured dynamic stiffness (or Young’s modulus) of the core by calculating the mass–
spring–mass resonance frequency; i.e. replacing sg′ with s′ of the core in Eq. 4.71.
Below fd , an estimate for the sound reduction index can be found by assuming only nonresonant (mass law) transmission. This can be calculated from the mass per unit area of
the sandwich panel using either the diffuse or field incidence mass law for an infinite plate
(Section 4.3.1.2.1).
481
S o u n d
I n s u l a t i o n
90
Sound reduction index (dB)
80
Sandwich panel:
13 mm plasterboard
(each side)
70
fc ≈ 3150 Hz
Core: 55 mm
polyurethane foam
(r = 27 kg/m3,
60
E = 6 ⫻ 106 Pa)
fd = 734 Hz
Sandwich panels
on both sides of
100 mm concrete
Sandwich panels on
both sides of
100 mm mineral wool
Sandwich panel:
Predicted nonresonant
transmission
(infinite plate:
field incidence)
50
40
30
20
100 125 160 200 250 315 400 500 630 800 1000 1250 1600 2000 2500 3150
One-third-octave-band centre frequency (Hz)
Figure 4.46
Measured sound reduction index for a sandwich panel and two other constructions formed using these panels. Predicted
values for non-resonant (mass law) transmission are shown as an estimate for the single sandwich panel below the dilatational
resonance, fd . Measured data are reproduced with permission from Homb et al. (1983).
Measurements from Homb et al. (1983) on a sandwich panel in various configurations are
shown in Fig. 4.46. The adverse effect of the dilatational resonance at fd is apparent for the
single sandwich panel; there is a deep dip below the infinite plate theory for non-resonant
(mass law) transmission. With many lightweight sandwich panels this resonance occurs in
the mid- or high-frequency ranges (e.g. see Jones, 1981). However, these panels are usually combined to form one or both leaves of a cavity wall, or connected to other plates to
form a multi-layer construction. Figure 4.46 shows two examples of this; in some constructions the adverse effect of the resonance dip can be avoided, in others its effect will still be
evident.
4.3.7 Composite sound reduction index for several elements
In Chapter 5 we will look at the combination of direct and flanking transmission that involves
vibration transmission between connected plates. The situation is often simpler for direct sound
transmission across partitions that are formed from more than one element. Some partitions
482
Chapter 4
such as façade walls or corridor walls contain a number of different elements; e.g., the wall
itself along with any windows, doors, holes, gaps, apertures, or ventilators. These elements
are often exposed to the same sound field, which is assumed to be diffuse. For walls, windows,
and doors it can usually be assumed that there is no significant exchange of vibrational energy
between them that alters their individual performance. It can also be assumed that there will
be no interaction between holes, gaps, apertures, and ventilators that are widely spaced apart;
the more important factor is their position in the wall (e.g. in the middle, at an edge, or in
a corner). Regardless of whether the sound reduction index of each element is measured
or predicted, they can be combined to give a single composite sound reduction index for
the wall.
For N elements each with a sound reduction index, Rn , and an area, Sn , the composite sound
reduction index is
N
1
−Rn /10
Rtotal = −10 lg N
Sn 10
(4.92)
n=1 Sn n=1
Note that Eq. 4.92 does not imply that all sound reduction indices can be scaled to different
areas. For example, the measured sound reduction index for a 1.5 × 1.25 m sheet of 6 mm
glass is not relevant when assessing a 0.1 × 0.3 m sheet to be used as a vision panel in a door.
For all plates it is worth calculating the fundamental bending mode and the critical frequency
to help decide whether using a test result for a different size of test element is appropriate for
the dominant sound transmission mechanism.
Section 4.3.10 contains calculation of the sound reduction index for slit-shaped and circular
apertures for which the opening area needs to be used in Eq. 4.92. As these are idealized
models, the opening area of each aperture is well-defined and it is clear that the sound reduction
index only applies to specific aperture dimensions. For practical purposes, measurements on
holes, gaps, slits, and ventilators are not normalized to an opening area and are almost always
described using Dn,e rather than R.
4.3.8
Surface finishes and linings
Surface finishes and linings not only increase or decrease the radiated sound from a base
wall or floor, they can also alter its basic properties (e.g. mass, stiffness, damping, airflow
resistance). A bonded surface finish can usually be considered as an integral part of the plate
that forms the wall or floor. In contrast, linings are often considered as a ‘bolt-on’ component
that is interchangeable; hence laboratory measurements of the sound reduction improvement index are commonly quoted for linings. This does not imply that these measurements
apply to all other base walls and floors; the result is usually specific to one type of base wall
or floor.
4.3.8.1 Bonded surface finishes
Many walls and floors have a bonded finish over the entire surface such as plaster, render, or
screed. These typically have a minimum thickness of 5 mm and can alter the airflow resistivity,
internal loss factor, the mass per unit area, and bending stiffness of the base plate. Such
changes subsequently affect the radiation and structural coupling losses to and from the plate.
483
S o u n d
I n s u l a t i o n
70
Measured (fair-faced)
60
Measured (13 mm plaster on one side)
Sound reduction index (dB)
SEA: Resonant and non-resonant transmission
50
40
30
20
10
0
50
80
125
200
315
500
800
1250 2000
One-third-octave-band centre frequency (Hz)
3150
5000
Figure 4.47
Effect of a bonded surface finish on a highly porous plate. 100 mm masonry wall (expanded clay blocks) with a 13 mm
lightweight plaster finish (one side). Measurements according to ISO 140 Part 3. Plate properties: Lx = 3.53 m, Ly = 2.63 m,
h = 0.1 m, ρs = 77 kg/m2 , cL = 2330 m/s, ν = 0.2, measured total loss factor. Fair-faced blocks have an airflow resistivity,
r = 6400 Pa.s/m2 . Plaster finish: ρs = 10 kg/m2 . Measured data from Hopkins are reproduced with permission from ODPM
and BRE.
A common bonded surface finish on a masonry wall is plaster or render. Most solid masonry
walls with plaster or render on one side (or both sides) can be adequately modelled as a
homogeneous plate (refer back to the examples in Section 4.3.1.3). In some cases, the surface
finish changes the bending stiffness and the total loss factor that is measured on the base plate;
hence these properties need to be used in any model. For a porous masonry wall, a bonded
surface finish can effectively remove the non-resonant transmission path via the pores. An
example is shown in Fig. 4.47; this fair-faced wall will later be modelled as a porous plate
in Section 4.3.9.2. With a plaster finish on one side, this highly porous wall can be modelled
as a solid homogeneous plate using SEA in the same way as for other masonry walls. This
suggests that despite the porous interface with the air on the fair-faced side of the wall, the
estimate for the radiation efficiency (method no. 3 in Section 2.9.4.3) is still reasonable over
the building acoustics frequency range.
When using transmission suite measurements to assess any differences between a fair-faced
masonry wall and the same wall with a bonded finish it is very useful to measure the total loss
factor. A surface finish can cause changes to the internal and coupling loss factors that result
in either an increase or a decrease in the total loss factor over various parts of the building
acoustics frequency range.
Measurements on 100 and 200 mm brick walls (1600 kg/m3 ) indicate that the addition of plaster
increases the quasi-longitudinal phase velocity (Craik and Barry, 1992). This increase also
occurs with other walls built from solid masonry blocks (densities up to 2000 kg/m3 ) and causes
the plastered wall to have a lower critical frequency than when it is fair faced. This shift in the
critical frequency is sometimes observed in laboratory measurements.
484
Chapter 4
4.3.8.2
Linings
Linings are used on both separating and flanking elements (e.g. floating floors, thermal wall
linings on masonry/concrete walls, suspended ceilings). Most linings consist of a plate with a
low mass per unit area (compared to the base wall or floor) and a critical frequency in the highfrequency range. This plate is isolated from the base wall or floor by an empty cavity, or a cavity
that is partially or fully filled with an absorbent and/or resilient material. The plate usually needs
to be supported by the base element with structural connections across this cavity. This general
description fits the five-subsystem SEA model that was previously used to model a plate–
cavity–plate system in Section 4.3.5. This model is appropriate for simple wall linings such
as plasterboard which is screwed to timber battens and screwed to a solid masonry/concrete
wall or floor. The five-subsystem model is used in the same way as for a timber joist floor
by adding a sixth subsystem to model structural coupling via the battens (Wilson and Craik,
1996). Many linings use resilient structural connections that are too complicated to model
without measuring properties relating to their dynamic stiffness. Reliance therefore tends to be
placed upon laboratory measurement of the sound reduction improvement index.
Some linings have no structural connections to the base wall or floor; these are commonly
referred to as independent linings. A frame is needed to support the independent lining so the
perimeter of this frame will be connected to the flanking walls and floors; structural coupling
between the lining and the base wall or floor therefore depends on these flanking elements.
In laboratory measurements the boundary conditions are usually arranged to avoid vibration
transmission via such a path. In situ this structural coupling is usually more important when the
flanking walls and floors are lightweight (e.g. timber frame walls/floors) rather than heavyweight
(e.g. concrete walls/floors).
Linings are formed from plates with a high critical frequency to try and take advantage of their
low radiation efficiency at frequencies well-below the critical frequency. This is based on the
assumption that all sound radiation by the lining will be due to resonant transmission and
that transmission paths involving non-resonant (mass law) transmission via the cavity will be
unimportant. This is often the case when the cavity is filled with highly absorbent material and
the base wall or floor has a high mass per unit area. However, the air in the cavity or any resilient
connections across the cavity can act as a spring element. This results in a mass–spring–mass
resonance frequency (Sections 4.3.5.1 and 4.3.5.4.1) at which more sound can be radiated
than from the base wall or floor without the lining. The addition of linings tends to increase
the sound insulation in the mid- and high-frequency ranges whilst reducing it near the mass–
spring–mass resonance. For this reason the design of linings usually starts with the aim of
getting the mass–spring–mass resonance frequency to be well-below the lowest frequency of
interest. However, constraints on the cavity depth and available resilient materials/connectors
often results in the resonance frequency falling within the low-frequency range. Whilst this is
an important part of the building acoustics frequency range, exact resonance frequencies and
their adverse effects are not easily predicted for lightweight linings (refer back to the discussion
in Section 4.3.5.1). Hence laboratory testing is usually needed to hone this aspect of the design.
The sound reduction improvement index is usually measured with airborne excitation to give
R, but it can also be measured with mechanical excitation to give RResonant as previously
described in Section 3.5.1.2.2. For the majority of flanking walls and floors that are built from
masonry/concrete it is only resonant transmission via the lining that is relevant. This applies to
the situation where the base wall or floor is only excited by structure-borne sound or mechanical
excitation. R includes non-resonant (mass law) transmission below the critical frequency
485
S o u n d
I n s u l a t i o n
21
(a)
⌬R
18
15
100 mm aircrete block wall
( ρs⫽51 kg/m2, fc⫽341 Hz)
⌬RResonant
12
9
2
Lining: 12.5 mm plasterboard (10.8 kg/m ,
fc⫽3480 Hz) which is screw fixed to a light
steel frame (37 mm cavity containing 25 mm
rock fibre quilt, 30 kg/m3)
Mass–air–mass: fmsm⫽104 Hz
6
3
Sound reduction improvement index (dB)
0
⫺3
⫺6
⫺9
⫺12
(b)
100 mm aircrete block wall
( ρs⫽51 kg/m2, fc⫽341 Hz)
12
9
Lining: 9.5 mm plasterboard (6.3 kg/m2,
fc⫽3770 Hz) laminated with 32 mm extruded
polystyrene (1.1 kg/m2, s⬘⫽70 MN/m3)
which is screw fixed to a light steel frame
(37 mm empty cavity)
Mass–air–mass: fmsm⫽123 Hz
6
3
0
⫺3
⫺6
⫺9
(c)
6
100 mm aircrete block wall
(ρs⫽51 kg/m2, fc⫽341 Hz)
3
0
⫺3
⫺6
50
80
125
200
315
500
800
1250 2000
One-third-octave-band centre frequency (Hz)
Lining: 9.5 mm plasterboard (6.3 kg/m2,
fc⫽3770 Hz) laminated with 32 mm expanded
polystyrene (0.4 kg/m2, s⬘⫽65 MN/m3) attached
with gypsum based adhesive dabs
3150 5000
(10 mm empty cavity)
Mass–air–mass: fmsm⫽247 Hz
Mass–resilient material–mass: fmsm⫽542 Hz
Figure 4.48
Measured sound reduction improvement index for three different wall linings on the same solid masonry wall. Measurements
used to determine R were made according to ISO 140 Part 3. Note that RResonant is only shown below the thin plate limit.
Measured data from Hopkins are reproduced with permission from BRE Trust.
and if the base wall or floor is porous, non-resonant transmission through the pores. Below the
critical frequency of non-porous base walls and floors, RResonant will usually be lower than R.
For porous base walls and floors, RResonant may be lower than R across a wider frequency
range.
The measured sound reduction improvement indices for three different wall linings on the
same masonry wall are shown in Fig. 4.48. Note that whilst negative values for the sound
reduction improvement index are usually due to mass–spring–mass resonances, other peaks,
and troughs can occur near the critical frequencies of the base wall and the lining. In the
high-frequency range a dip or plateau often occurs due to the high-radiation efficiency of the
lining at and above its critical frequency. The critical frequency of the base wall and the mass–
spring–mass resonance are usually both in the low-frequency range. The mass–spring–mass
resonance may only have a distinguishable dip when it is well separated from the critical
486
Chapter 4
35
Sound reduction improvement index, ∆R (dB)
Expanded clay ( ρs ⫽ 77 kg/m2, cL ⫽ 2330 m/s,
fc ⫽ 278 Hz, r ⫽ 6400 Pa.s/m2)
30
Lightweight aggregate ( ρs⫽140 kg/m2, cL⫽2070 m/s,
fc ⫽ 313 Hz, r ⫽ 526 000 Pa.s/m2)
25
Dense aggregate (ρs ⫽ 200 kg/m2, cL⫽3290 m/s,
fc = 197 Hz, estimated r > 1 MPa.s/m2)
20
100 mm masonry wall
15
Wall lining: 9.5 mm plasterboard
(fc⫽3770 Hz) attached with gypsum-based
adhesive dabs leaving empty cavities
between dabs with a depth of ≈12 mm
Mass–air–mass: fmsm ≈ 220 Hz
10
5
0
⫺5
50
80
125
200
315
500
800
1250
2000
3150
5000
One-third-octave-band centre frequency (Hz)
Figure 4.49
Measured sound reduction improvement index for one type of wall lining on three different 100 mm solid masonry walls.
Measurements used to determine R were made according to ISO 140 Part 3. Measured data from Hopkins are reproduced
with permission from ODPM and BRE.
frequency of the base wall. Examples (a) and (b) have a single mass–spring–mass resonance
frequency due to the air in the cavity; these are well-below the critical frequency of the base wall.
Example (c) has two resonance frequencies; one associated with the air in the cavity (close
to the critical frequency of the base wall) and the other due to the resilient material acting as a
spring. Note that any calculated mass–spring–mass resonance frequency only gives a rough
indication of the frequency bands in which the sound reduction index of the wall is actually
reduced by the lining. Mechanical excitation tends to emphasize the dip at the mass–spring–
mass resonance. This aspect is worth noting for point force excitation such as with impacts
on walls and floors, or some types of machinery/equipment. However, highly negative values
at the resonance frequency are not usually observed when a flanking wall or floor is excited
along its boundaries by structure-borne sound waves transmitted from other connected plates.
The measured sound reduction improvement indices for the same wall lining on three different
masonry walls are shown in Fig. 4.49. These walls all have very different airflow resistivities
and mechanical impedances. This indicates that one type of lining can have significantly different sound reduction improvement indices depending on the base wall or floor. Different
487
S o u n d
I n s u l a t i o n
base elements usually have different impedances and this will change vibration transmission
between the lining and the base plate across the structural connections; this can be seen
by referring back to the theory for point connections in Section 4.3.5.4.1. Laboratory measurements are usually only taken on base walls or floors that have been sealed with paint or
plaster, so non-resonant transmission due to porosity is usually excluded from the measurement. Outside of the laboratory, wall linings are usually applied to fair-faced walls and it is
not appropriate to assume that the measurement of the sound reduction improvement index
applies to any fair-faced wall. It has been shown that a highly porous masonry wall may reduce
the depth of the mass–spring (air)–mass dip in the sound reduction index due to a lining with a
cavity behind it (Heckl, 1981; Warnock, 1992). The majority of masonry walls do not have very
low airflow resistivities, and it is difficult to confirm that this effect is always measurable when
r > 20 000 Pa.s/m2 because of other variables that affect the level of sound insulation. Some
laboratory measurements indicate that applying a bonded surface finish such as render onto
a 200 mm dense aggregate masonry wall (2100 kg/m3 ) to seal the wall before applying a plasterboard lining gives a lower sound reduction index near the mass–spring–mass resonance
compared to when the base wall is fair faced (Mackenzie et al., 1988).
The sound reduction improvement index is only measured with a lining on one side of a base
wall or floor. Estimates of the sound reduction index of the same base element with the same
or different linings on each side can be calculated by simply adding R for each lining to R
for the base wall. Measurements on non-porous walls show that this approach can give good
estimates (Warnock, 1991) although there is also evidence that simple addition of R near
the mass–spring–mass resonance should only be considered as a rough estimate (Fothergill
and Alphey, 1989).
4.3.9
Porous materials (non-resonant transmission)
We now look at sound transmission through porous materials that form a partition, or form
part of a wall or floor. Fundamental aspects of the theory relating to porous materials were
introduced in Section 1.3.2, and provide the necessary background for this section.
The model that is described here gives the normal incidence sound reduction index, R0◦ , for
single sheets of homogeneous porous materials (Bies and Hansen, 1980; Schultz, 1988). For
porous materials in long thin cavities, normal incidence is relevant because it applies to the
axial modes. However, it is usually diffuse incidence sound fields that are of more interest. Fortunately, the sound reduction index, R, measured in the laboratory for these materials is often
similar to R0◦ for normal incidence (Schultz, 1988). Hence R0◦ can be used as an estimate for R.
Three frequency ranges are used in the model, A, B, and C. Each range is defined in terms
of the material thickness, d, relative to the wavelength of sound within the equivalent gas, λpm
(Section 1.3.2.2). For fibrous materials and some foams this can be calculated from empirical
equations such as those of Delany and Bazley (Section 1.3.2.2).
The calculations for frequency ranges A, B, and C are described below.
Frequency range A: d < λpm /10
In this frequency range the skeletal frame of the porous material has a low mass impedance,
therefore the compressions and rarefactions of the air particles in the pores of the material cause the entire frame to move. This allows sheets of porous material to be modelled
488
Chapter 4
using a lumped parameter approach from electrical circuit theory. The normal incidence sound
reduction index is (Schultz, 1988)
⎡
⎢
R0◦ = 10 lg ⎢
⎣1 +
rd
ρ 0 c0
4
2πf ρs
ρ0 c0
rd
ρ 0 c0
2
2
+
4+
rd
ρ0 c 0
2πf ρs
ρ0 c 0
2
⎤
⎥
⎥
⎦
(4.93)
where r is the airflow resistivity, d is the thickness of porous material, and ρs is the mass per
unit area of the sheet of porous material.
Frequency range B: λpm /10 ≤ d < λpm
There is no specific model for this frequency range. Hence in range B, R0 is determined by
fitting a curve to the R0 values that have been calculated for ranges A and C. A suitably
smooth transition from range A to B to C can usually be achieved by using at least a third-order
polynomial curve. When the errors in the curve fit are negligible, this smooth curve may be
used to represent R0 in ranges A, B, and C.
Range B can cover a large part of the building acoustics frequency range. To carry out the
curve fitting it is sometimes necessary to calculate values that are below the 50 Hz one-thirdoctave-band in range A, and above the 5000 Hz one-third-octave-band in range C.
Frequency range C: d ≥ λpm
In this range the frame can be considered as rigid and the concept of an equivalent gas can be
used to represent sound propagation within the porous material. Hence the calculations require
Z0,pm and kpm for the equivalent gas. These may be calculated from empirical equations such
as those of Delany and Bazley (Section 1.3.2.2).
A fraction of the sound energy that is incident upon the porous material will enter the material;
the remaining fraction is reflected back from the entry surface. Once inside the porous material,
the sound is attenuated as it propagates through the material towards the exit surface. At the
exit surface, a fraction is reflected and the remaining fraction is transmitted; the subsequent
travels of the reflected wave within the material are assumed to have an insignificant effect on
the overall sound transmission.
The propagation loss, LP (dB) within the porous material is (Eq. 1.177)
LP =
20
Im{kpm } d
ln 10
(4.94)
The entry/exit loss, LE (dB) at the entry or exit surface is determined from the reflection
coefficient for the porous material, Rpm , using
2
LE = −10 lg (1 − Rpm )
(4.95)
where the reflection coefficient for normal incidence is (Eq. 1.74)
Rpm
Z0,pm
−1
ρ c
= 0 0
Z0,pm
+1
ρ0 c0
(4.96)
489
S o u n d
I n s u l a t i o n
(a) Rock wool
45
Measured: Sound reduction index, R (dB)
Predicted: Normal incidence sound reduction index, R0˚ (dB)
40
35
Measured: Rock wool
r ⫽ 80 kg/m3, r ⫽ 37 000 Pa.s/m2
Measured: Rock wool
r ⫽ 45 kg/m3, r ⫽ 15 000 Pa.s/m2
Measured: Rock wool
r ⫽ 33 kg/m3, r ⫽ 9000 Pa.s/m2
Measured: Rock wool
r ⫽ 28 kg/m3, r ⫽ 7000 Pa.s/m2
Predicted: Ranges A and C
30
25
Predicted: Range B
20
15
Predicted: Range C
10
5
0
50
80
125
200
315
500
800
1250
One-third-octave-band centre frequency (Hz)
2000
3150
5000
Figure 4.50
Sound reduction index for 100 mm thick sheets of mineral wool – comparison of measured and predicted data. Symbols
used for the predicted values correspond to those used for the measurements. Measurements according to ISO 140 Part 3.
Measured data from Hopkins are reproduced with permission from ODPM and BRE.
Therefore the entry/exit loss is
⎡
2
;
< 2
<
< 2
;
;
Z
Z0,pm
Z0,pm
− 1 + 4 Im ρ0,pm
+ Im ρ0 c0
Re ρ0 c0
⎢
c
0 0
⎢
LE = −10 lg ⎢1 −
;
< 2
;
< 2 2
⎣
Z
Z
1 + Re ρ0,pm
+ Im ρ0,pm
0 c0
0 c0
2
The normal incidence sound reduction index, R0◦ , can then be calculating using
R0◦ = LP + 2LE
⎤
⎥
⎥
⎥
⎦
(4.97)
(4.98)
4.3.9.1 Fibrous sheet materials
We will now look at some examples for transmission across 100 mm thick sheets of rock
and glass wool (average fibre diameter ≈ 5 µm) in the direction of longitudinal airflow resistivity. Comparison of measured R from a transmission suite (S ≈ 2 m2 ) and predicted R0◦ are
shown in Fig. 4.50. Although the model gives sound transmission at normal incidence, this
490
Chapter 4
(b) Glass wool
45
Measured: Sound reduction index, R (dB)
Predicted: Normal incidence sound reduction index, R0˚ (dB)
40
Measured: Glass wool
r ⫽ 32 kg/m3, r ⫽ 15 000 Pa.s/m2
Measured: Glass wool
r ⫽ 24 kg/m3, r ⫽ 7000 Pa.s/m2
35
Predicted: Ranges A and C
30
Predicted: Range B
25
20
15
10
5
0
50
80
125
200
315
500
800
1250
One-third-octave-band centre frequency (Hz)
2000
3150
5000
Figure 4.50
(Continued)
is a reasonable estimate of the sound reduction index with a diffuse incidence sound field.
For 100 mm thick fibrous materials, the difference between measured R and predicted R0◦
is typically up to ±5 dB. Differences between rock and glass wool are not identified by this
simple model. This would require more thorough models that also account for structural waves
in the porous material (Biot theory); these waves become important for very thick layers. Such
models for porous materials can be found in the book by Allard (1993).
4.3.9.2 Porous plates
Sound transmission across porous plates depends on the mass, stiffness, and damping of
the plate as well as the airflow resistance, porosity, and other parameters that describe sound
propagation through the pores.
Porous plates in buildings are most commonly encountered in the form of fair-faced masonry
walls. For these walls there may also be air paths due to gaps or slits formed at the joints
between the blocks. It is simplest to consider two types of masonry wall; those with mortar
joints along all edges where all airflow is assumed to occur through the blocks, and those
491
S o u n d
I n s u l a t i o n
where distinct slits are created between blocks that only touch, or slot together. For the latter
type, sound transmission through the slit can be measured if the slit is particularly complicated,
or calculated for a simple straight slit (see Section 4.3.10.1). For both types, the air path
introduces a non-resonant transmission path across the plate. The sound reduction index
of a porous blockwork wall is therefore determined by sound transmission due to resonant
transmission as well as three types of non-resonant transmission: (1) mass law, (2) slits at the
block joints, and (3) sound propagation through the porous blocks.
We have already seen that non-resonant transmission through a porous material with a rigid
frame can be calculated by treating the material as an equivalent fluid. This requires knowledge of the complex wavenumber and the characteristic impedance. For porous plates such
as masonry walls it is necessary to either measure these parameters, calculate them by assuming an idealized geometry for the porous structure, or to use empirical models (Allard, 1993;
Attenborough, 1993; Voronina, 1997; Wilson, 1997). Information on the microstructure of a
masonry block from which to choose a model is rarely available. However, measurement of
the airflow resistance on small samples cut from individual blocks is relatively quick and simple.
This is measured in the same way as for other porous materials (Section 3.11.1). A model is
then needed to link the airflow resistivity to the complex wave number and the characteristic
impedance. Unless a specific empirical model is available it is possible to fall back on the
empirical equations of Delany and Bazley (Section 1.3.2.2). Although these were based on
measurements of fibrous materials with porosities close to unity, the resulting equations are
semi-empirical due to the way in which the data was normalized. For this reason they have a
slightly more general application to porous materials. (An example of this is the way that these
equations were previously used to estimate the ground impedance for outdoor sound propagation models, even though ground porosities are often well-below unity. Nowadays they are
rarely used because more accurate models have been developed for specific types of ground.)
Figure 4.51 compares the measured and predicted sound reduction index for two fair-faced
masonry walls built from highly porous blocks (mortared along all edges). One wall is built from
wood fibre aggregate blocks and the other with expanded clay blocks; both of which have airflow
resistivities less than 20 000 Pa.s/m2 . The equivalent gas model gives a reasonable estimate
of the sound reduction index with differences between measurements and predictions typically
up to ±5 dB; a similar accuracy to the previous examples shown for mineral wool.
Block density is not an indicator of airflow resistivity; it depends on the material and the manufacturing process. For this reason there are blocks with identical densities that have very
different airflow resistivities as well as low-density blocks which have a high airflow resistivity.
Available data indicates that for design purposes, the sound reduction index of fair-faced
masonry walls (up to 200 mm thick) may be considered in three main groups: (1) low airflow
resistivity: r ≤ 20 000 Pa.s/m2 where the sound reduction index may be estimated as described
above; (2) intermediate airflow resistivity: 20 000 < r ≤ 300 000 Pa.s/m2 where the sound reduction index can only be accurately determined by laboratory measurements on full-size walls;
and (3) high airflow resistivity: r > 300 000 Pa.s/m2 where gaps at the mortar joints start to
become more important than airflow through the blocks, (Richards, 1959; Warnock, 1992;
Watters, 1959; Williamson and Mackenzie, 1971). For groups (1) and (2), the sound reduction
index can often be increased above that of the fair-faced wall by using paint to seal one or both
surfaces of the porous blocks. Note that bonded surface finishes such as plaster may not only
change the airflow resistivity but other wall properties too (Section 4.3.8.1).
492
Chapter 4
30
Measured: Sound reduction index, R (dB)
Predicted: Normal incidence sound reduction index, R0˚ (dB)
25
20
Masonry wall: Wood fibre aggregate blocks
h ⫽ 0.09 m, ρs ⫽ 98 kg/m2, r ⫽ 14 000 Pa.s/m2
(Measured data – Warnock, 1992)
15
10
5
Measured
Predicted: Ranges A and C
0
Predicted: Range B
20
Masonry wall: Expanded clay blocks
h ⫽ 0.1 m, ρs ⫽ 77 kg/m2, r ⫽ 6400 Pa.s/m2
15
10
5
0
50
80
125
200
315
500
800
1250
2000
One-third-octave-band centre frequency (Hz)
3150
5000
Figure 4.51
Sound reduction index for porous masonry blockwork walls – comparison of measured and predicted data. Measured data for
wood fibre blocks according to ASTM E90 are reproduced with permission from Warnock (1992) and the National Research
Council of Canada. Measured data for expanded clay blocks according to ISO 140 Part 3 from Hopkins are reproduced with
permission from ODPM and BRE.
The sound reduction improvement index for a lining on a wall or floor is also affected by the
porosity of the base wall or floor (Section 4.3.8.2).
4.3.9.3 Coupling loss factor
Sound transmission across a sheet of porous material or a porous plate is incorporated into the
SEA framework as a non-resonant transmission mechanism between two spaces. The coupling
loss factor between two rooms is calculated using Eq. 4.26 where S is the surface area of the
porous sheet or porous plate. The non-resonant transmission coefficient is calculated from R0◦ .
The coupling loss factor between cavities or between a cavity and a room is determined using
the same approach as in Section 4.3.1.2 for a one- or two-dimensional incident sound field.
4.3.10
Air paths through gaps, holes, and slits (non-resonant transmission)
Air paths tend to reduce the airborne sound insulation so sealing is an important issue with all
wall and floor constructions. Slits can be found along unsealed edges of walls and floors and
493
S o u n d
I n s u l a t i o n
w
d
d
2a
Figure 4.52
Slit-shaped and circular apertures.
between bricks, blocks, or sheet materials whereas circular apertures can be found as small
drill holes, large holes formed for ventilation purposes, or pipes used for cables. To identify the
importance of air paths it is useful to be able to estimate their effect on the sound insulation.
For slit-shaped apertures with straight-edges or for circular apertures (see Fig. 4.52) there are
approximate diffraction models to determine sound transmission for a plane wave impinging
upon an aperture at normal incidence. These models also provide reasonable estimates for
diffuse sound fields. For this reason they are useful in identifying the basic features that affect
the transmission loss particularly when trying to identify possible leakage problems. In addition
they show that the sound reduction index for the slit cannot simply be assumed to be 0 dB at
all frequencies.
Real air paths often take tortuous routes through more complex cross-sections than a simple
slit-shaped or circular aperture. Accurate prediction is fraught with difficulty for many real apertures, especially with resilient seals around windows and doors where the sound transmission
is dependent upon their compression in situ. However, the models for simple apertures are
useful in validating laboratory measurements before taking measurements on more complex
apertures.
4.3.10.1 Slit-shaped apertures (straight-edged)
For a slit-shaped aperture in the middle of a large plate, or along an edge formed by perpendicular plates, the transmission coefficient is (Gomperts, 1964; Gomperts and Kihlman, 1967)
τ=
mK cos2 (Ke)
2n2
5
K2
sin2 (KX + 2Ke)
+ 2 [1 + cos (KX ) cos (KX + 2Ke)]
2
cos (Ke)
2n
6
(4.99)
where the end correction, e, for a straight-edged slit (assuming a slit of infinite length in an
infinite baffle) is
1
8
e=
ln − 0.57722
(4.100)
π
K
494
Chapter 4
and X = d/w (where d is the depth of the slit, w is the width of the slit), K = kw (where k is the
wave number), m is a constant for the incident sound field (m = 8 for a diffuse field or m = 4
for a plane wave at normal incidence), and n is a constant depending on the position of the slit
(n = 1 for a slit in the middle of a plate such as along one or both vertical edges of a door, and
n = 0.5 for a slit along an edge such as along the threshold of a door).
Equation 4.99 assumes cylindrical waves are radiated from the slit and that w ≪ λ. This allows
calculations over the building acoustics frequency range for most slits. Equation 4.99 also
assumes that the slit is of infinite length. In practice, sound transmission via a slit can be
considered to be independent of the slit length, l, when l > λ , and will depend only on its width
and depth (Gomperts and Kihlman, 1967). To assess its suitability for typical slit lengths in the
low-, mid- and high-frequency ranges we can look at the wavelength for the lowest band centre
frequency in each range; λ50Hz = 6.86 m (low), λ200Hz = 1.72 m (mid), λ1000Hz = 0.34 m (high).
Hence it is suited to slits running along the longest dimension of most walls and floors over most
of the building acoustics frequency range. However for most windows and doors it only tends
to be appropriate in the mid- and high-frequency ranges; this is sufficient for most calculations.
On the basis of comparisons with measurements, it can be assumed that the actual sound
reduction index in the low-frequency range will be higher than the calculated value (Gomperts
and Kihlman, 1967).
Maxima in the transmission coefficient occur at the resonance frequencies across the slit. For
a plane wave at normal incidence, resonance frequencies occur when
d + 2e = z
λ
2
(4.101)
where z = 1, 2, 3 etc.
Figure 4.53 shows the sound reduction index for various different slits that can be found in
buildings using Eqs 4.99 and 4.37. These assume that the incident sound field is diffuse (m = 8).
Note that the sound reduction index for an aperture can be negative at some frequencies and
that dips occur at the resonance frequencies. Below the first resonance frequency the sound
reduction index tends to be higher for a slit in the middle of a plate rather than along an edge.
The deepest dip tends to occur at the first resonance frequency. The resonance dips become
more prominent in the building acoustics frequency range as the slit depth increases from 6 to
200 mm.
The presence of apertures with prominent dips at the resonance frequencies needs to be borne
in mind when scrutinizing sound insulation test results for indications of air paths. Plates with
critical frequencies in the high-frequency range (e.g. glass, plasterboard) may have critical
frequency dips at similar frequencies to resonance dips associated with the aperture.
The model uses normal incidence, but the resonance frequencies depend on the angle of
incidence; for an incident angle of 45◦ the resonance frequency tends to be higher than for
normal incidence (Mechel, 1986). Very few slits in buildings are straight-edged rectangular
slits with perfectly uniform dimensions and it is diffuse incidence sound fields that are of most
interest; hence it is rarely possible to accurately predict their resonance frequencies.
Laboratory measurements indicate that Eq. 4.99 is valid when w < 0.3λ (Bodlund and Carlsson,
1989; Gomperts and Kihlman, 1967; Oldham and Zhao, 1993). These measurements also
indicate that deep dips at the resonance frequencies predicted by Eq. 4.99 do not always
occur in practice. This model assumes that the air is non-viscous; hence no account is taken
495
S o u n d
I n s u l a t i o n
15
Sound reduction index (dB)
10
15
Upper set of three curves: In the middle of a plate
Lower set of three curves: Along the edge of a plate
Upper set of four curves: In the middle of a plate
Lower set of four curves: Along the edge of a plate
10
5
5
0
0
⫺5
⫺5
⫺10
⫺10
Slit depth: 6 mm
⫺15
Slit depth: 12 mm
Width: 0.5 mm
⫺15
Width: 0.5 mm
Width: 1 mm
Width: 1 mm
⫺20
⫺20
Width: 2 mm
Width: 2 mm
⫺25
100
1000
Width: 4 mm
10 000
⫺25
100
Frequency (Hz)
Sound reduction index (dB)
15
15
All curves: In the middle of a plate
10
5
5
0
0
⫺5
⫺5
⫺15
All curves: Along the edge of a plate
⫺10
Slit depth: 50 mm
Slit depth: 50 mm
Width: 0.5 mm
Width: 2.5 mm
⫺15
Width: 1 mm
⫺20
Width: 5 mm
Width: 10 mm
⫺20
Width: 2 mm
Width: 20 mm
Width: 4 mm
⫺25
100
1000
10 000
⫺25
100
15
Sound reduction index (dB)
All curves: Along the edge of a plate
10
10
5
5
0
0
⫺5
⫺5
Slit depth: 100 mm
⫺10
Slit depth: 200 mm
Width: 1 mm
Width: 1 mm
⫺15
⫺15
Width: 2 mm
Width: 2 mm
Width: 4 mm
⫺20
1000
Frequency (Hz)
Figure 4.53
Predicted sound reduction index for different size slits.
496
Width: 4 mm
Width: 8 mm
Width: 8 mm
⫺25
100
10 000
15
All curves: Along the edge of a plate
⫺20
1000
Frequency (Hz)
Frequency (Hz)
⫺10
10 000
Frequency (Hz)
10
⫺10
1000
10 000
⫺25
100
1000
Frequency (Hz)
10 000
Chapter 4
of any damping mechanism. Viscous losses at resonance tend to be more important for long
hairline cracks with widths less than a few millimetres. Validation of more accurate theories
which incorporate viscous losses is awkward because of the variation between laboratory
measurements; this is possibly due to the effect of different materials that form the slit, and
the accuracy with which narrow slits can be created and measured (Lindblad, 1986). As a
rule-of-thumb it can be assumed that the deep dips predicted by Eq. 4.99 at the resonances
are unlikely to occur in practice when d/w > 50.
4.3.10.2 Circular aperture
Sound transmission through a circular aperture can be predicted by assuming infinitely thin,
rigid pistons at each end of the circular aperture to simulate the motion of air particles at the
entry and exit points of the aperture; it is then assumed that plane waves propagate inside the
aperture (Wilson and Soroka, 1965).
The transmission coefficient for a plane wave at normal incidence on a circular aperture with
radius, a, and depth, d, in an infinite baffle is (Wilson and Soroka, 1965)
τ=
4R02 [ cos (kd)
− X0
sin (kd)]2
4R0
+ [(R02 − X02 + 1) sin (kd) + 2X0 cos (kd)]2
(4.102)
where R0 and X0 are the resistance and reactance terms for the radiation impedance of a
piston in an infinite baffle given by
Zrad = ρ0 c0 πa2 (R0 (2ka) + iX0 (2ka))
(4.103)
The resistance term, R0 is given by
R0 (2ka) = 1 −
(2ka)2
(2ka)4
(2ka)6
2J1 (2ka)
=
−
+
− ···
2ka
2×4
2 × 4 2 × 6 2 × 42 × 62 × 8
(4.104)
and the reactance term, X0 is given by
4
X0 (2ka) =
π
2ka
(2ka)3
(2ka)5
− 2
+ 2
− ···
3
3 × 5 3 × 52 × 7
(4.105)
For 2ka > 17.8 the following approximations can be used (Wilson and Soroka, 1965)
1
ka
&
π
1
sin 2ka −
πka
4
'
2
1
π
1−
X0 (2ka) =
sin 2ka +
πka
ka
4
R0 (2ka) = 1 −
(4.106)
Although the derivation is based upon a plane wave at normal incidence it was originally
validated against transmission suite measurements; for practical purposes it can be assumed
to be equally applicable to reverberant or diffuse incidence sound fields (Wilson and Soroka,
1965). The assumption of plane waves propagating in one dimension within the aperture implies
that λ ≫ 2a. The upper frequency limit for a cylindrical tube can be taken as 2a/λ ≈ 0.6, but
validation against measurements indicates that it still gives reasonable estimates even when
497
Sound reduction index (dB)
S o u n d
I n s u l a t i o n
30
30
25
25
20
20
15
15
10
10
5
5
0
0
⫺5
⫺5
Aperture radius: 5 mm
⫺10
⫺10
Depth: 12.5 mm
Aperture depth: 200 mm
⫺15
Depth: 25 mm
⫺15
Radius: 5 mm
⫺20
Depth: 50 mm
⫺20
Radius: 10 mm
⫺25
100
Radius: 20 mm
Depth: 100 mm
1000
Frequency (Hz)
10 000
⫺25
100
1000
Frequency (Hz)
10 000
Figure 4.54
Predicted sound reduction index for different size circular holes.
2a/λ is slightly greater than unity (Oldham and Zhao, 1993; Wilson and Soroka, 1965). Other
models for oblique incidence allow for holes positioned at an edge or corner, and for the effect
of sealing and filling the holes (Mechel, 1986).
In the far field, the directivity of the radiated sound from the aperture can be inferred from
the directivity of a piston. When λ ≫ 2πa there is uniform radiation in all directions over a
hemisphere, but the radiation becomes increasingly directional when λ < 2πa with sound mainly
radiated in the direction perpendicular to the surface of the circular aperture (Morse and Ingard,
1968).
Figure 4.54 shows the sound reduction index from Eqs 4.102 and 4.37 for different circular
apertures that can be found in buildings. The sound reduction index below the first resonance
(which often occurs in the low- and mid-frequency ranges) can be much higher than with typical
slits (refer back to Fig. 4.53). However, in the same way as with slits, dips also occur in the
sound reduction index at the resonance frequencies (Eq. 4.101). Deep dips can be predicted
for long, small holes but they don’t tend to occur in practice due to viscous losses of the air. A
damping factor, D, to account for these losses at the resonance frequencies is (Gomperts and
Kihlman, 1967; Oldham and Zhao, 1993)
⎞−2
⎛
d + 0.5πa
ρ0 2μω
⎜
⎟
πa3
⎟
D=⎜
(4.107)
⎝1 +
⎠
2
ρ0 ω
πc0
where μ is the coefficient of viscosity for air (1.56 × 10−5 m2 /s at 20◦ C and 0.76 mHg).
The correction −10 lg D in decibels is added to the sound reduction index at the resonance
frequency (or each frequency band containing a resonance). This correction has been found
to give good agreement with measurements at the first resonance frequency (Oldham and
Zhao, 1993).
498
Chapter 4
4.3.10.3
More complex air paths
Many air paths are formed by apertures that are neither slit shaped nor circular. Trying to apply
idealized shapes of apertures to more complex shapes is rarely appropriate. For this reason,
laboratory measurements are usually necessary to quantify their effect.
Particularly complicated slits can occur around window frames; these slits may have one or
more internal bends, or the slits on either side of the frame may lead into a larger internal
volume embodied within the frame. The resulting decrease in sound insulation is not always
confined to the high-frequency range, it can also occur in the low- and mid-frequency ranges
too. Measurements confirm that gaps around frames rarely act as simple rectangular slits, and
sometimes act similarly to Helmholtz resonators if there are thin slits on either side that lead
to a larger volume of air within the frame (Burgess, 1985; Lewis, 1979).
Electrical socket boxes in separating walls are often blamed for poor airborne sound insulation
in the field, particularly when they are back to back; although this is not always justified. With
lightweight walls, the adverse effects (if any) tend to occur in the mid- and high-frequency
ranges and are highly dependent upon the wall construction, workmanship, and the type of
socket box (Nightingale and Quirt, 1998; Royle, 1986).
Lining the interior surfaces of a slit-shaped aperture with a porous absorbent material will not
necessarily increase the sound reduction index at all frequencies. For two slits with the same
width and length, but where the interior of one slit has an absorbent lining, the sound reduction
index for the latter may be higher at high frequencies when the absorption coefficient of the
lining is high, but it can be lower at low frequencies (Bodlund and Carlsson, 1989).
4.3.10.4 Using the transmission coefficients
To determine transmission coefficients for one-third-octave or octave-bands, the values at
single frequencies are averaged in each frequency band. Below the first resonance frequency,
reasonable estimates can be found by taking the value at the band centre frequency. Assuming
that we already have the sound reduction index for a wall or floor we can estimate the adverse
effect of introducing apertures. This can be calculated using Eq. 4.92 with the predicted sound
reduction index for a slit-shaped or circular aperture.
The transmission coefficients in this section are for single apertures in a baffle, not closely
spaced arrays of slit-shaped or circular apertures that are sometimes used to form ventilation
devices. When there are N circular apertures that are spaced apart by at least λ/4, the resulting
sound reduction index is 10 lg N lower than the sound reduction index for a single aperture
(Morfey, 1969). Note that when calculating the effect of more than one aperture using Eq. 4.92
it is the sound reduction index for a single aperture that is used along with the total area of all
the apertures.
The change to the sound reduction index, R, of a wall (area, S) due to the introduction of N
apertures (each with an area, Sa,n and a sound reduction index, Ra,n ) can be described by R,
where
⎡
⎢
R = Rwith apertures − Rwithout apertures = −10 lg ⎣
S+
1
N
n=1
Sa,n (10−Ra,n /10 )
10−R/10
S+ N
n=1 Sa,n
⎤
⎥
⎦ (4.108)
499
S o u n d
I n s u l a t i o n
80
Lightweight wall (sealed)
Predicted with slit (width: 0.5 mm, depth: 70 mm, length: 4 m)
70
Predicted with slit (width: 0.5 mm, depth: 80 mm, length: 4 m)
Sound reduction index (dB)
Predicted with slit (width: 0.5 mm, depth: 90 mm, length: 4 m)
60
50
40
30
20
10
100 125 160 200 250 315 400 500 630 800 1000 1250 1600 2000 2500 3150 4000 5000
One-third-octave-band centre frequency (Hz)
Figure 4.55
Predicted effect of a long, narrow slit-shaped aperture along one edge of a lightweight wall on the sound reduction index.
Note that R is usually referred to as the sound reduction improvement index. For apertures
this is a rather positive way of describing a generally negative effect.
Narrow slits along the top and/or bottom of a wall can go unnoticed visually. These sometimes
occur at the top of heavyweight walls where they meet the roof or underside of a floor. They
can also occur along unsealed tracks of the framework along the perimeter of lightweight walls.
These slits may only be 0.5 mm wide but they are still capable of significantly reducing the sound
insulation. This is illustrated by the example in Fig. 4.55. Note that gaps around the perimeter of
lightweight walls do not usually form simple slit-shaped apertures, but the general trends shown
in this example correspond to those observed in measurements (Royle, 1986). In practice the
resonance dips are not usually quite as deep because of damping due to air viscosity.
An example showing the effect of circular apertures is shown in Fig. 4.56 for 50 small holes
drilled through a 215 mm solid brick wall to give a total open area of 6637 mm2 (Fothergill and
Alphey, 1987). For each aperture, the first, second, third, and fourth resonances occur in the
800, 1600, 2500, and 3150 Hz bands respectively. These circular apertures are narrow and
long, and the deep dips predicted by Eq. 4.102 (particularly at the first resonance) do not occur
in practice because no account has been taken of viscous losses in the aperture. Using a
damping correction (Eq. 4.107) in bands that contain resonances gives close agreement with
the measurements. Of course the sound insulation is unlikely to be your primary concern if
the neighbours start to drill 50 holes through your bedroom wall. A more important point is
illustrated in the upper part of Fig. 4.56. This shows that with a single sheet of plasterboard
on each side of the wall the effect of the holes in the brick wall is hardly identifiable; note that
the holes were not covered up by the adhesive dabs used to fix the plasterboard. When a
construction fails a field test and the airborne sound insulation in the mid- and high-frequency
500
Chapter 4
70
60
215 mm solid brick wall with plasterboard lining attached
using adhesive dabs (≈10 mm cavity) on both sides
of the brick wall with and without holes
Measured (no holes)
Apparent sound reduction index (dB)
50
Measured (with holes)
40
30
215 mm solid brick wall with and without holes
Measured (no holes)
Measured (with holes)
Predicted (with holes)
60
50
Predicted (with holes) assuming
viscous damping. This is
only used in frequency
bands containing
resonance frequencies
40
30
Holes:
50 circular apertures
0.0065 m radius
At least 0.5 m from perimeter
20
100 125 160 200 250 315 400 500 630 800 1000 1250 1600 2000 2500 3150
One-third-octave-band centre frequency (Hz)
Figure 4.56
Measured and predicted effect of 50 small circular holes on the airborne sound insulation of a brick wall. Measurements
according to ISO 140 Part 4. Measured data are reproduced with permission from Fothergill and Alphey (1987), ODPM,
and BRE.
ranges is lower than expected, it is common to start looking for holes, slits, and gaps. This
is a logical, and relatively simple, first step. However, most buildings contain plenty of hidden
holes and discovery of a few of them may place unwarranted emphasis on poor workmanship
as the cause of failure. Whilst poor workmanship is often to blame, this example provides a
reminder that holes do not always have a significant effect on the overall sound insulation. In
practice it is necessary to pay equal (if not more) consideration to flanking transmission as a
possible cause of failure.
Sound transmission across an aperture is incorporated into the SEA framework as a nonresonant transmission mechanism between two spaces. Hence the coupling loss factor
between these two spaces is calculated using Eq. 4.26 where S is the surface area of the
aperture opening. Although transmission through a slit is predicted on the assumption that the
slit is infinitely long, its actual opening area is used when calculating the coupling loss factor.
For continuous holes or slits across cavity walls such as those formed by pipes or box-shaped
sections, the effect of the aperture can be calculated in the same way as for an aperture in a
solid plate. However, the device/object that forms this aperture may also introduce structural
501
S o u n d
I n s u l a t i o n
coupling across the cavity wall. For holes or slits in one or both leaves of a plate–cavity–plate
system, it may be appropriate to use an SEA model to determine sound transmission into
and out of a reverberant cavity. This is not suitable when the slits or holes in each plate are
opposite each other because the sound radiated by an aperture is highly directional. If there is
an aperture in the plate that faces into the receiving room, use of the transmission coefficients
requires that the cavity is large enough to support a three-dimensional sound field so that there
is sound incident upon the aperture.
4.3.11
Ventilators and HVAC
Most ventilators have quite complex forms and incorporate absorbent material; hence it is
necessary to rely on laboratory measurements to quantify their sound insulation. A few throughwall vents have very simple grilles and are not vastly different to the simple circular apertures
discussed in Section 4.3.10.2; these can sometimes be adequately modelled as circular apertures in the low- and mid-frequency range (Ohkawa et al., 1984). Equation 4.99 for slit-shaped
apertures highlights an important point for the sound insulation of any ventilator; the effect of
the incident sound field. The sound insulation depends upon whether a ventilator is positioned
in the middle of a wall, along an edge, or in a corner.
Prediction of sound transmission from HVAC systems is thoroughly covered in industry guidance documents such as those by ASHRAE (Anon, 2003) and CIBSE (Leventhall et al., 2002).
These usually give reverberant sound pressure levels or sound power levels that can simply
be combined with predictions of the sound transmitted by direct and flanking transmission in
the building.
4.3.12
Windows
Many window constructions are formed from an insulating glass unit (IGU). These are also
referred to as thermal glazing units, or double glazing, and consist of two panes of glass
separated by a cavity (typically <20 mm) in a hermetically sealed unit. It is difficult to predict
accurately the sound insulation of windows formed from an IGU; hence there is a dependence
on laboratory measurements. In addition, the effect of seals around an openable window
can only be assessed through measurement. However, most of the important features that
determine the sound reduction index can be discussed in terms of the simpler models that
have already been introduced in this chapter. These can often be used to help make design
decisions alongside laboratory measurements.
4.3.12.1 Single pane
We start by considering a single pane of float glass. This can be considered as a solid homogenous isotropic plate for which its sound reduction index is determined by non-resonant and
resonant transmission for a finite plate. An example for 6 mm glass has already been discussed
in Section 4.3.1.3.1. Other types of glass such as wired, textured/patterned, or toughened glass
have nominally identical material properties to float glass; although an average thickness will
need to be assumed for textured/patterned glass. The main factors that cause significant
changes in the measured sound reduction index are the boundary conditions and the niche in
which the glass is mounted (refer back to Section 3.5.1.3.3 on the niche effect). Concerning
resonant transmission, the radiation efficiency below the critical frequency can be calculated
by assuming that the boundaries are simply supported or clamped (Section 2.9.4). Glass is
502
Chapter 4
mounted in many different types of frame; hence either boundary condition is possible, as well
as various degrees of clamping in-between. The mounting can also affect the total loss factor
by introducing edge damping (an additional internal loss) as well as changing the structural
coupling losses from the glass. Non-resonant transmission usually dominates over resonant
transmission below the critical frequency, so changes to the boundary conditions only tend to
become apparent at frequencies near, at and above the critical frequency.
Non-resonant transmission across finite plates depends on plate size (Section 4.3.1.2.2);
smaller plates tend to transmit less than larger plates. For this reason, it is possible for a
perfectly airtight window formed from a large number of small glass panes to have higher
sound insulation than a window with a single pane (identical area) at frequencies below the
critical frequency. In practice this improvement may only be seen in the low-frequency range
because of the increased length of slits/gaps around the frame with windows consisting of
multiple panes that are not perfectly sealed.
4.3.12.2 Laminated glass
We now consider laminated glass. This is formed from two sheets of glass that are permanently
bonded together by a relatively soft interlayer such as polyvinyl butyral (PVB) or polymethyl
methacrylate (PMM). A laminate is advantageous because of the high damping that is achieved
by constraining an interlayer between two plates, as well as a reduction in the bending stiffness
compared to a solid plate. The internal loss factor of a laminate plate is high because of energy
losses associated with shear deformation of the interlayer. For this reason they are useful
for attenuating bending wave motion rather than in-plane wave motion. However, the material
properties of laminate plates are not as simple as with most solid plates because both the bending stiffness and the internal loss factor vary with frequency and temperature. Therefore reliance
tends to be placed upon measurement of these properties rather than trying to predict them from
the individual properties of the interlayer and the plates. These measurements were discussed
in Sections 3.11.3.4 and 3.11.3.7. An example of the frequency and temperature dependence
of the internal loss factor for laminated glass was previously shown in Fig. 3.91. There can be
large differences between laminates with different types of interlayer. They are often optimized
to give the highest damping at a specific temperature and/or a specific frequency, as well as
to shift the critical frequency up to higher frequencies where the adverse critical frequency
dip may be more tolerable. With increasing frequency and/or increasing temperature over the
building acoustics frequency range, the internal loss factor tends to increase, and the bending
stiffness tends to decrease (Kerry and Ford, 1983; Yoshimura and Kanazawa, 1984).
The benefits of laminate glass compared to a single pane of glass with the same thickness are
that the critical frequency is shifted to a higher frequency and the depth of the critical frequency
dip is reduced by the higher damping. To take advantage of the higher internal losses at higher
temperatures it is beneficial for the laminate pane in an IGU to face the side with the higher
temperature; this is dependent on the climate and could be indoors or outdoors. Temperature
effects for a single pane of laminate glass tend to be important near, at and above the critical
frequency. This needs to be considered when comparing measured sound insulation data
from the laboratory and the field, and when making comparisons between measurements on
different laminates. An example of the effect of temperature on the sound reduction index of a
single pane of laminate glass is shown in Fig. 4.57 (Yoshimura and Kanazawa, 1984).
Material property measurements on laminates sometimes give values at individual frequencies so it is necessary to interpolate between them in order to find values of the bending
503
S o u n d
I n s u l a t i o n
50
30˚C
Sound reduction index (dB)
45
19˚C
10˚C
40
35
30
25
20
50
80
125
200
315
500
800 1250 2000
One-third-octave-band centre frequency (Hz)
3150
5000
Figure 4.57
Effect of surface temperature on the measured sound reduction index of a 1.7 × 1.3 m pane of laminated glass (4 mm
glass/1.1 mm interlayer/4 mm glass) in an aluminium frame. Measured data are reproduced with permission from Yoshimura
and Kanazawa (1984).
stiffness and the internal loss factor for each one-third-octave-band centre frequency. With
these values it is possible to estimate the sound reduction index by assuming that the plate is
solid, homogenous, and isotropic as described in Section 4.3.1. Note that it is the total loss factor that needs to be used in the calculation and this will depend on the mounting in the frame.
This approach to prediction generally gives good agreement with laboratory measurements
(Ford, 1994; Yoshimura and Kanazawa, 1984).
4.3.12.3 Insulating glass unit (IGU)
The next step is to consider an IGU comprised of single or laminated glass panes. For these
units the modal behaviour of the two panes is tightly coupled together by the air or other gas in
the cavity as well as the structural connections (spacer bar and sealant) around the perimeter.
It is therefore not possible to satisfy the SEA assumption that net power transferred between
the two panes (or between each pane and the cavity) is proportional to the difference in their
modal energies. Other prediction models are generally needed for these types of systems such
as finite element methods, or methods based on multi-layered infinite structures (Villot et al.,
2001). However, these units are so common in buildings it is useful to be able to identify the
basic features that affect the sound reduction index. Two important features that cause dips
in the sound reduction index are the mass–spring–mass resonance frequency and the critical
frequencies of each pane.
The mass–spring–mass resonance frequency for air-filled units can be estimated using
Eq. 4.72. For units filled with other gases (e.g. argon, sulphur hexafluoride) Eqs 4.70 and 4.71
can be used. Compared with air, a different gas may only shift the mass–spring–mass frequency into an adjacent one-third-octave-band. The mass–spring–mass resonance frequency
504
Chapter 4
for most IGUs occurs in the range 100–400 Hz. In this frequency range there are also vibrational modes of the unit (Pietrzko, 1999); these are better described as global modes of the
unit due to the strong coupling between the individual panes. Therefore the individual panes
of glass do not act as simple lump masses at the mass–spring–mass resonance and sound
transmission is determined by a combination of modal behaviour and the air acting as a spring.
Any change in the vibrational modes can change the mass–spring–mass resonance frequency
and the depth of the dip. These global modes are altered by the structural connections at the
perimeter of the unit as well as the dimensions of the unit (Rehfeld, 1997). This means that it
is not always appropriate to assume similar performance from a nominally identical unit with
different dimensions. Variation in the depth of the dip is exacerbated when the mass–spring–
mass resonance lies in the low-frequency range because the sound reduction index may also
be affected by the modes of the source and/or receiving room (Section 4.3.3). The depth of
the mass–spring–mass resonance dip can be deep when the width and height of the IGU are
identical; this can occur even when the panes have different thickness or one is a laminate
(Michelsen, 1983). If we temporarily consider the local mode model for each individual pane
with identical dimensions then we would expect the modal response of each pane to be characterized by degenerate modes at which there will be a strong response. This strong response
will also be found in the global modes of the unit.
We now look at the frequency range in-between the mass–spring–mass resonance and the
lowest critical frequency of an IGU. The most common range of cavity depths is between 6
and 16 mm. Over this range the panes are still tightly coupled together. The change in sound
reduction index over this range of cavity depths is relatively small; hence tabulated sound
reduction indices often quote average values for this range of cavities (e.g. see EN 12758).
However, the type of spacer and sealant can still significantly change the sound reduction
index in this frequency range due to its effect on the structural coupling around the perimeter of
the panes (Gösele et al., 1977). Any advantage in choosing a unit with low structural coupling
can be negated by the non-resonant transmission path from the source to the receiving room
if there are gaps or slits around the frame of an openable window. If these airpaths are present
they usually reduce the sound reduction index in the mid- and high-frequency ranges (e.g. see
Michelsen, 1983). For argon-filled units, the phase velocities, and densities for air and argon are
not vastly different, hence for most practical purposes the sound reduction index of these units
is considered to be the same (Inman, 1994). However large differences occur when the gas
is sulphur hexafluoride. This gas has a much lower phase velocity and a much higher density
than air (see Table A1). The mass–spring–mass resonance frequency may not be significantly
different to air-filled units, but coupling via the sound field in the cavity is significantly reduced.
This leads to large increases in the sound insulation above the mass–spring–mass resonance
frequency (Gösele et al., 1977; Kerry and Inman, 1986; Rückward, 1981). An example of the
effect of different gas fills in an IGU is shown in Fig. 4.58 (Kerry and Inman, 1986).
If the IGU has panes with identical thickness there is usually a distinct dip at the critical frequency. For this reason most manufacturers produce units where one pane is thicker than the
other to avoid a single critical frequency dip that is overly deep. Using panes with different
thickness gives two shallower critical frequency dips. A single deep dip can also be avoided
by using one laminate and one non-laminate pane; the higher internal damping of the laminate
also being beneficial at the critical frequency. For units filled with sulphur hexafluoride, each
plate radiates sound into the air in the room on one side, and into the gas in the cavity on the
other. The plate therefore has a very different critical frequency on each side due to different
phase velocities in air and in sulphur hexafluoride. The latter has a much lower phase velocity
505
S o u n d
I n s u l a t i o n
60
4-12-6 IGU filled with:
55
Air
Sound reduction index (dB)
50
Argon
45
Sulphur hexafluoride
40
35
30
25
20
15
10
50
80
125
200
315
500
800
1250
2000
3150
5000
One-third-octave-band centre frequency (Hz)
Figure 4.58
Effect of different gas fills on the measured sound reduction index of an IGU. Measurements according to ISO 140 Part 3.
Measured data are reproduced with permission from Kerry and Inman (1986).
than air hence the critical frequency for radiation into the cavity is significantly lower than into
air. These different critical frequencies are sometimes apparent as dips in the sound reduction
index (Gösele et al., 1977; Kerry and Inman, 1986).
4.3.12.4 Secondary/multiple glazing
Whilst different gas fills can be used to increase the sound insulation of an IGU in the midand high-frequency ranges, it is difficult to increase it in the low-frequency range without using
secondary (or multiple) glazing. This requires two panes of glass separated by a wide air gap;
note that one or both panes could be an IGU or a laminate. With a wide airspace of at least
50 mm the two panes are no longer tightly coupled. Most practical designs for secondary glazing
use at least 6 mm thick glass and have at least a 50 mm cavity; hence the mass–spring–mass
resonance frequency tends to be below 100 Hz.
If the two panes are structurally isolated from each other and there are no air paths around
the frame, then at frequencies between the mass–spring–mass resonance and the first crosscavity mode there is usually the potential to increase the sound insulation by up to 10 dB by
at least doubling a cavity depth within the range 50–150 mm; so the resulting cavity depth is
within the range 100–300 mm. When the panes and the cavity are large enough to support
local modes, the five-subsystem SEA model for a plate–cavity–plate system can be used to
aid design decisions. Note that any increase in sound insulation with increasing cavity depth
may not be identified by an SEA model depending on how the non-resonant transmission is
modelled (Sections 4.3.5.2.1 and 4.3.5.3). For practical window constructions any increase in
the cavity depth beyond 300 mm tends to give negligible improvement due to the existence of
flanking transmission, structural coupling, and air leakage (e.g. see Inman, 1994). Deep critical
frequency dips can be avoided by using panes with different critical frequencies; this is seen
in the example from Section 4.3.5.2.1.
506
Chapter 4
Absorbent reveals in the cavity are used to reduce sound transmission via paths involving the
cavity. Most reveal linings are only highly absorbent in the high-frequency range. This is usually
above the first cross-cavity mode so the five-subsystem SEA model can be used to assess the
effect of different reveal linings.
Infinite plate models for plate–cavity–plate systems show dips in the sound insulation due to
the cross-cavity modes (Fahy, 1985; London, 1950). These dips are rarely (if ever) observed
in one-third-octave-band measurements. However such models have led to designs of plate–
cavity–plate systems with non-parallel plates on the basis that this will prevent or suppress
any adverse effects due to cross-cavity modes. By placing one pane of glass at an angle to
the other pane, the cavity depth will vary between Lz,min and Lz,max . When Lz,max = 3Lz,min ,
laboratory measurements from Quirt (1982) indicate that the sound insulation is nominally the
same as a rectangular cavity with a depth of 2Lz,min ; the sound insulation is lower than a
rectangular cavity with a depth of 3Lz,min ; and the sound insulation is higher than a rectangular
cavity with a depth of Lz,min . In practice this means that there is little to be gained from using
non-parallel plates. Above the first cross-cavity mode this is apparent from the five-subsystem
SEA model because the three-dimensional sound field in the cavity is assumed to be diffuse.
Hence if the cavity volume and absorption area are unchanged, there will be no increase in
sound insulation if it is only the shape of the cavity that is changed. It has also been shown that
the sound insulation does not increase if the central pane of glass in a triple-glazed window is
placed at an angle to the parallel outer panes (Rose, 1990).
4.3.13
Doors
Doors are sometimes formed from solid homogeneous plates, but more often from closely connected plates, sandwich panels, or plate–cavity–plate systems on a frame. There tends to be a
dependence on laboratory measurements for all door constructions because the effectiveness
of the seals around the perimeter is critically important in determining the achievable sound
insulation. For compression seals (as opposed to drop-down or wipe seals) the performance
may be specific to the type of door and door closer. The design aim with doors and doorsets is
often to minimize the door mass and the required opening force whilst maximizing the sound
insulation. This requires simultaneous consideration of sound transmission via the door and
the seals (e.g. see Hongisto, 2000; Hongisto et al., 2000).
For doors without any seals, sound transmission via the gaps around the perimeter means that
accurately predicting transmission via the door itself can become less important than predicting
transmission via the gaps. For simple slit-shaped apertures between the door and the frame,
the sound reduction index can be predicted as described in Section 4.3.10.1. An example is
shown in Fig. 4.59 for a 12 mm thick glass door (such as in commercial buildings) with and
without slit-shaped apertures at one door jamb or the threshold. Due to limitations of the slit
model, the effect of the slits is only shown where l > λ. The slits cause a significant decrease
in the sound reduction index in the mid- and high-frequency ranges. In this example the door
is relatively thin so there are no resonance dips due to the slits within the building acoustics
frequency range.
4.3.14
Empirical mass laws
In Section 4.3.1.2 we looked at non-resonant (mass law) transmission across infinite and finite
plates for which the mass per unit area plays the lead role (for infinite plates) or a dominant role
507
S o u n d
I n s u l a t i o n
50
12 mm glass door (fc ⫽ 1039 Hz)
With slit-shaped aperture (width: 2 mm) at one door jamb
Sound reduction index (dB)
40
With slit-shaped aperture (width: 5 mm) at the threshold
30
20
10
50
80
125
200
315
500
800
1250
2000
3150
5000
One-third-octave-band centre frequency (Hz)
Figure 4.59
Predicted effect of slit-shaped apertures on the sound reduction index of a 12 mm thick glass door (0.8 × 2.0 m). The door
jamb is assumed to be in the middle of a wall, rather than along an edge like the threshold.
(for finite plates) in determining the non-resonant transmission coefficient. On the basis that
sound insulation is complex and should be simplified as far as possible, it has been the custom
for many years to create empirical relationships that link the mass per unit area to the measured
weighted sound reduction index. The concept is attractive because of its simplicity and if they
are used to give rough estimates (say ±5 dB Rw ) they can be very useful. However, looking
back at the examples in this chapter it is clear that bending stiffness, damping (internal and
total loss factors), plate size, thickness resonances, and porosity also play important roles in
determining the sound reduction index; not all of these variables are simply related to the mass
per unit area. For masonry walls this has led to different empirical mass laws for different types
of blocks in different countries. There are now so many empirical mass laws in existence without
traceability to specific types of block properties, surface finish, and boundary conditions that
the concept has become devalued. An empirical mass law of any worth is limited to a specific
end use. For example, interpolating between measured sound reduction indices for one type
of masonry wall with different thicknesses, or producing a design curve for quick calculations
where the sound reduction indices have been normalized to a stated total loss factor.
Before establishing an empirical mass law it is important to identify the dominant sound transmission mechanisms that primarily determine the weighted sound reduction index. For many
solid masonry walls with high airflow resistivity, resonant transmission tends to be the dominant
transmission mechanism over most of the building acoustics frequency range. For this reason
the sound reduction index that is measured in the laboratory depends upon the total loss factor.
This will vary depending on the material properties of the test element and the laboratory structure. Unfortunately the potential for normalizing measurements to a reference total loss factor
over the entire building acoustics frequency range does not extend to all types of masonry
walls (refer back to Sections 3.5.1.3.2 and 4.3.1.4). There is some logic in forming separate
508
Chapter 4
empirical mass laws for fair-faced masonry walls with high airflow resistivity, and masonry walls
with a bonded surface finish (e.g. plaster). This is because a bonded surface finish tends to
alter the bending stiffness as well as sealing the pores on the surface of a masonry wall. It may
be possible to combine them into a single dataset afterwards.
For non-porous, homogeneous, isotropic solid plates it is much more convenient to create theoretical mass laws rather than using laboratory measurements (Gerretsen, 1999); this allows
the total loss factor to be chosen in terms of its relevance to the field or the laboratory situation.
4.4 Impact sound insulation
For impact sound insulation it is simplest to consider heavyweight and lightweight floors separately. This is partly due to the complexity of many lightweight floor constructions and the use
of more than one impact source to measure their impact sound insulation.
Heavyweight or lightweight base floors often use floor coverings to improve the impact sound
insulation. These tend to be soft coverings or floating floors; both of which are usually least
effective in the low-frequency range and most effective in the high-frequency range. Floor
coverings are used on top of the base floor to increase the impact sound insulation by changing
the power input from the impact source and/or isolating the walking surface from the base floor.
The resulting improvement of impact sound insulation depends upon the impact source (e.g.
ISO tapping machine, ISO rubber ball) and the base floor. As a first step it is better to reduce the
power input into the base plate using a floor covering rather than trying to redesign the ceiling
to reduce the radiated sound. The latter approach may lead to significant flanking transmission
from the base plate into the flanking walls that subsequently radiate into the receiving room.
4.4.1 Heavyweight base floors
The simplest situation is where the ISO tapping machine excites a solid homogeneous isotropic
plate that radiates sound into the receiving room. This is described by a two-subsystem SEA
model where subsystem 1 is the plate and subsystem 2 is the receiving room (see Fig. 4.60).
To simplify matters it is reasonable to assume that the power flow that returns to the plate from
the room is negligible (i.e. η21 = 0). Recalling Eqs 4.8 and 4.9 and writing them in terms of the
total loss factor gives
Win(1) ≈ ωη11 E1 + ωη12 E1 = ω(η11 + η12 )E1 = ωη1 E1
ωη12 E1 = ωη2 E2
(4.109)
(4.110)
The energy in the receiving room is found by substituting Eq. 4.109 into Eq. 4.110 to give
η12 Win(1)
(4.111)
E2 =
η1 η2 ω
Equation 4.111 can now be used to give the required impact sound pressure level. The coupling
loss factor from the plate to the room, η12 , is given by Eq. 4.21. The power input from the
ISO tapping machine was discussed in detail in Section 3.6.3. Here we will take the simplest
approach by assuming that there are short-duration impacts on a thick concrete floor slab. The
hammer impedance is assumed to be negligible compared to the driving-point impedance of
the plate, and the latter is assumed to equal that of an infinite plate. Hence the power input is
2
Win = Frms
3.9B
2.3ρcL h2
≈
(2.3ρcL h2 )2 + (ωm)2
2.3ρcL h2
(4.112)
509
S o u n d
I n s u l a t i o n
Wd(1)
Win(1)
ISO tapping
machine
Subsystem
1
Plate
(1)
W21
Wd(2)
W12
Subsystem
2
Receiving
room
(2)
Figure 4.60
Two-subsystem SEA model for impact sound insulation of a solid floor.
which gives the normalized impact sound pressure level as
ρ02 c02 σ
Ln = 10 lg
+ X dB
ρ2 h3 cL η
(4.113)
where η is the total loss factor of the plate, X = 78 dB for one-third-octave-bands and X = 83 dB
for octave-bands.
Figure 4.61 compares measurements on a 140 mm concrete floor slab with two predicted values; one prediction uses the infinite plate mobility (Eq. 4.113), and the other uses the measured
driving-point mobility (50–1250 Hz). Below the critical frequency where Ns < 1 the differences
between measurements and predictions are typically up to 5 dB. The use of measured mobility does not significantly improve the prediction in the low-frequency range where the mode
count is low and the modal fluctuations are pronounced; this is partly due to uncertainty in the
measured structural reverberation time and the predicted radiation efficiency. At and above the
critical frequency there is close agreement between measurements and predictions. In the highfrequency range the difference often increases because the force spectrum is no longer flat and
starts to tail off with increasing frequency (Section 3.6.3.2); this effect is often exacerbated in
field measurements when the surface is not smooth and clean (Section 3.6.5.4). In addition, the
thin plate assumption for the driving-point mobility starts to break down above the thin plate limit.
Heavyweight floors such as beam and block floors tend to be more complex to model due to
their modular nature. Beam and block floors with a bonded surface finish sometimes allows
their driving-point mobility to be approximated by an infinite homogeneous plate (refer back
to Fig. 2.54). This does not automatically imply that they act as homogeneous plates in all
respects. When these floors are bonded across their surface they do not always behave as
homogeneous plates concerning their airborne sound insulation (refer back to Fig. 4.24). In
addition there may be a significant decrease in vibration level with distance across a beam
and block floor (refer back to Fig. 2.42). A comparison of the impact sound insulation for beam
and block floors with a concrete floor slab is shown in Fig. 4.62. Although they have a similar
performance to the concrete slab, the model for a homogeneous plate does not necessarily
provide an adequate model for how these particular beam and block floors behave. There is
no general rule that covers all types of modular floor. However, there are some designs of
510
Chapter 4
Ns ⫽ 0.6 0.7 0.9 1.1 1.4 1.8 2.3 2.8 3.6 4.5 5.7 7.1 9.1 11 14 18 23 28 36 45 57
Normalized impact sound pressure level (dB)
M ⫽ 0.4 0.2 0.1 0.2 0.2 0.2 0.2 0.2 0.3 0.4 0.5 0.7 0.7 0.8 1.0 1.1 1.1 1.1 1.6 1.9 2.1
80
75
70
f11 ⫽ 32 Hz
fc ⫽ 122 Hz
65
fB(thin) ⫽ 1485 Hz
60
Measured
55
SEA (thin plate)
SEA (thin plate) using measured driving-point mobility
50
50
80
125
200
315
500
800
1250 2000
One-third-octave-band centre frequency (Hz)
3150
5000
Figure 4.61
140 mm concrete floor slab. Upper x-axis labels show the predicted statistical mode count and modal overlap factor for
the plate in each frequency band. Measurements according to ISO 140 Part 6. Plate properties: Lx = 4.2 m, Ly = 3.6 m,
h = 0.14 m, ρs = 345 kg/m2 , cL = 3800 m/s, ν = 0.2, measured total loss factor (approximately 0.005 + 0.3f−0.5 ). Measured
data from Hopkins are reproduced with permission from ODPM and BRE.
Normalized impact sound pressure level (dB)
80
75
70
65
150 mm concrete slab ( ρs ⫽ 320 kg/m2, fc ⫽ 108 Hz)
60
150 mm beam and block floor with 5 mm levelling
compound ( ρs ⫽ 313 kg/m2, fc,eff ≈ 150 Hz)
55
150 mm beam and block floor with 70 mm screed
( ρs ⫽ 443 kg/m2, fc,eff ≈ 150 Hz)
50
50
80
125
200
315
500
800
1250
2000
3150
5000
Figure 4.62
Comparison of a solid homogeneous concrete floor slab with an orthotropic beam and block floor (different surface finishes).
The blocks and beams are solid, and the only rigid material that bonds them together is the surface finish. The measured total
loss factors for these three floors are within 2 dB of each other. Measurements according to ISO 140 Part 6. Measured data
are reproduced with permission from ODPM and BRE.
511
S o u n d
I n s u l a t i o n
beam and block floor where the bonded finish results in a plate that can be modelled as being
homogeneous and orthotropic (e.g. see Gerretsen, 1986; Patrício, 2001).
4.4.2 Lightweight base floors
For lightweight plates that form the walking surface of the floor, the force spectrum from the
ISO tapping machine is significantly different to concrete floors (refer back to Fig. 3.32). Most
lightweight floors are formed from plates and beams so the driving-point mobility is frequencydependent and will differ depending on whether the tapping machine is above or in-between
the beams. The application of SEA models to quantify the impact sound insulation tends to be
limited; although these models are sometimes useful in a qualitative way. The impact sound
insulation of timber joist floors without a resiliently mounted ceiling or floating floor can be
predicted with more complex analytic and numerical models (Brunskog and Hammer, 2003
a, b). For such floors, these models indicate that significant increases in the impact sound
insulation can be gained by increasing the joist depth (when there is absorbent material in
the cavity), and increasing the mass per unit area of the ceiling by adding additional sheet
material such as plasterboard (Brunskog and Hammer, 2003c). Lightweight separating floors
between dwellings tend to be constructed from several layers of board materials and use
isolating elements such as resilient hangers or resilient channels to support the ceiling. These
aspects make it more difficult to predict the impact sound insulation.
Timber joist floor constructions vary around the world, and optimizing these constructions
usually relies upon laboratory measurements. There can be significant variation in the impact
sound insulation between nominally identical timber floors in situ (e.g. see Johansson, 2000)
and laboratory measurements usefully identify aspects of the construction that cause some of
this variation (e.g. see Fothergill and Royle, 1991; Warnock and Birta, 1998). In the low- and
mid-frequency ranges, the impact sound insulation spectrum is often characterized by peaks
and troughs. Many changes to the floor construction tend to enhance, reduce, or shift the
frequency of these peaks and troughs. Whether these changes are of benefit usually depends
on the rating system used to determine the single-number quantity.
In the absence of a simple model for lightweight floors it is still possible to interpret and make
decisions from measurements by using SEA path analysis in a qualitative manner. An example
using four timber joist floors is shown in Fig. 4.63. The walking surface is unchanged so the
power input from the tapping machine can be taken as being the same for each floor. Floor
A is the most basic construction to which changes can be made. Floor B is formed from A by
adding mineral wool into the cavities to absorb sound. However, the change in the impact sound
insulation is only a few decibels. We can either infer that the transmission path via the joists
dominates over any paths involving the sound field in the cavity, or that the mineral wool is not
very effective at absorbing sound in the cavity. If the former assumption is reasonable, then
reducing the structural coupling between the joists and the ceiling should give a significant
improvement. This is confirmed by floors C and D where the plasterboard is supported by
resilient channels. This does not automatically imply that we can now remove the mineral wool
from the cavities of floors C and D and achieve the same result. The resilient channels have
significantly reduced the strength of the structural transmission path, so transmission paths via
the sound field in empty cavities may now be of greater importance. Whilst resiliently suspended
ceilings tend to improve the impact sound insulation in the building acoustics frequency range
they sometimes reduce it below 50 Hz (Parmanen et al., 1998).
512
Chapter 4
90
A
85
18 mm OSB, 40 ⫻ 240 mm timber joists at 400 mm
centres, 15 mm plasterboard
Normalized impact sound pressure level (dB)
80
75
70
B
65
As A, but with 100 mm mineral wool in the cavities
60
Timber joist floor
55
C
A
50
As B, but with resilient metal channels running
perpendicular to the joists to support the
plasterboard ceiling
B
45
C
40
D
35
30
50
80 125 200 315 500 800 1250 2000 3150 5000
One-third-octave-band centre frequency (Hz)
D
As C, but with two layers of 15 mm plasterboard
for the ceiling
Figure 4.63
Impact sound insulation of different timber joist floors using the ISO tapping machine. Measurements according to ISO 140
Part 6. Measured data from Hopkins and Hall are reproduced with permission from ODPM and BRE.
4.4.3 Soft floor coverings
Soft floor coverings are commonly found in the form of carpet or vinyl. In terms of the improvement of impact sound insulation, L, their performance tends to differ depending on whether
the base floor is heavyweight or lightweight.
4.4.3.1 Heavyweight base floors
The ISO tapping machine is used as the impact source on heavyweight floors. The improvement
of impact sound insulation due to a soft floor covering on such a heavyweight floor can be
determined using the contact stiffness of the spring-like floor covering that is ‘seen’ by the
hammer of the tapping machine. This is appropriate because a soft floor covering usually has
negligible effect on the total loss factor and bending stiffness of a heavyweight base floor. It can
therefore be assumed that it only alters the force input. On heavyweight floors with or without
a soft covering, the force pulse from each hammer gives rise to an under-critical oscillation
(Section 3.6.3.1). The magnitude of the peak force from the tapping machine can be calculated
with and without a soft covering to give |Fn |with and |Fn |without respectively. The improvement of
impact sound insulation can then be calculated from (Lindblad, 1968; Vér, 1971)
|Fn |without
L = 20 lg
(4.114)
|Fn |with
Example force spectra for a 140 mm concrete floor slab with and without soft floor coverings
are shown in Fig. 4.64. For covering No. 1, E/d = 1.5 × 1011 N/m3 , which is indicative of a
few millimetres of solid PVC. For covering No. 2, E/d = 2.8 × 108 N/m3 ; this could potentially
513
S o u n d
I n s u l a t i o n
Magnitude of the peak force, |Fn| (N)
10
1
0.1
0.01
140 mm concrete slab
Slab with soft floor covering No. 1
Slab with soft floor covering No. 2
0.001
10
100
1000
10 000
Frequency (Hz)
Figure 4.64
Force spectrum for the ISO tapping machine on a concrete floor slab with and without a soft floor covering.
represent some carpet or vinyl layers with a soft resilient backing. We will shortly discuss the
difficulty in relating these force spectra to real soft floor coverings.
Whilst the force spectrum is initially flat at low frequencies, there is a cut-off frequency, fco ,
above which the force decreases. For the concrete slab, the cut-off frequency (Eq. 3.102) due
to the contact stiffness of the concrete (Eq. 3.97) is above the building acoustics frequency
range (fco ≈ 7000 Hz). When the soft floor coverings are fixed to the slab, the contact stiffness
of the covering (Eq. 3.98) gives the cut-off frequency as fco ≈ 2300 Hz for covering No. 1, and
fco = 100 Hz for covering No. 2. Below fco the soft floor covering does not significantly alter the
force input compared to the bare slab; hence it does not improve the impact sound insulation.
For soft floor coverings there are deep troughs in the force spectra above the cut-off frequency;
these occur at frequencies, nfco where n = 3, 5, 7, etc. These troughs occur because the model
does not include the internal loss factor, ηint , of the soft floor covering. This damping can be
included in the model by replacing E for the floor covering with E(1 + iηint ). However these
loss factors are not usually known, and can be frequency dependent (Pritz, 1996). This is not
particularly problematic because material damping is usually high enough for the troughs to be
shallow; hence after averaging into one-third-octave or octave-bands there will rarely be any
significant ripple in the curve.
Using the force spectrum we can now calculate the improvement of impact sound insulation
from Eq. 4.114; this is shown in Fig. 4.65. The peaks in L correspond to the troughs in
the force spectra. If internal damping is incorporated using the internal loss factor, the curves
will tend towards a straight slope of 12 dB/octave (equivalent to 40 dB/decade) for f ≥ fco . In
practice, measurements of L rarely show any ripple. Below fco , L is ≈ 0 dB. Hence by calculating fco it is possible to estimate L with two straight lines. This model assumes a linear
spring for the soft floor covering and is useful in identifying general features of L. However,
514
Chapter 4
Improvement of impact sound insulation (dB)
80
70
Soft floor covering No. 1
60
Soft floor covering No. 2
50
0 dB f < fco
40
12 dB/oct f ≥ fco
30
20
10
0
⫺10
10
100
1000
10 000
Frequency (Hz)
Figure 4.65
Improvement of impact sound insulation due to soft floor coverings on a 140 mm concrete floor slab.
we now need to return to an issue relating to the ISO tapping machine that was discussed
in Section 3.6.3.3; namely, soft floor coverings acting as non-linear springs (Lindblad, 1968).
The linear spring model results in a single characteristic frequency, fco , above which there is a
12 dB/octave increase in impact sound insulation. However, the relatively high force from the
ISO tapping machine causes a non-linear response from some soft floor coverings. This results
in L having more than one distinct slope. In practice, some materials have two or three slopes
ranging from 5 to 22 dB/octave. Considering this fact alongside the finding that some coverings
have a frequency-dependent Young’s modulus (Pritz, 1996) means that reliance tends to be
placed upon laboratory measurements of L.
Increasing the thickness of a homogeneous soft covering decreases the contact stiffness, which
in turn will reduce the cut-off frequency and subsequently increase L. Whilst increasing the
thickness is generally beneficial, materials used for soft coverings do not always show such
simple relationships and there can sometimes be negligible increase in L with increasing
thickness.
For soft floor coverings on heavyweight floors, typical ranges of L are shown in Fig. 4.66 for
common materials. The dominant role of the contact stiffness in determining L means that
laboratory measurements of soft coverings usually give good estimates of L when in situ for
many other types of concrete base floor, such as ribbed or hollow core plank floors.
4.4.3.2
Lightweight base floors
For lightweight floors the impact source may be the ISO tapping machine or a heavy impact
source such as the ISO rubber ball. With a soft floor covering the tapping machine tends to
give minor improvements in the impact sound insulation for the low-frequency range, with more
significant improvements in the mid- and high-frequency ranges. Heavy impact sources indicate
much smaller improvements, often with little or no improvement in the low- and mid-frequency
range. An example is shown in Fig. 4.67 for a carpet on a timber floor (Inoue et al., 2006).
515
S o u n d
I n s u l a t i o n
70
Improvement of impact sound insulation (dB)
Typical minimum and maximum values for:
60
Carpet
50
Vinyl floor coverings
with resilient backing
40
Vinyl floor coverings
30
20
10
0
-10
50
80
125 200 315 500 800 1250 2000 3150 5000
One-third-octave-band centre frequency (Hz)
Figure 4.66
Typical ranges for the improvement of impact sound insulation using the ISO tapping machine for soft floor coverings on
heavyweight floors.
Improvement of impact sound insulation (dB)
50
ISO tapping machine
40
Bang machine
ISO rubber ball
30
20
10
0
⫺10
31.5
63
125
250
500
1000
2000
4000
Octave-band centre frequency (Hz)
Figure 4.67
Improvement of impact sound insulation for a carpet on a timber floor (Reference floor No.1 from ISO 140 Part 11).
Measurements according to ISO 140 Part 11. Measured data are reproduced with permission from Inoue et al. (2006).
4.4.4 Floating floors
Floating floors generally consist of a rigid walking surface ‘floating’ on a resilient material with
no rigid connections to the surrounding walls at the edges of the floating floor. There are three
516
Chapter 4
Base floor
Base floor
Base floor
Figure 4.68
Main types of floating floor.
main types as shown in Fig. 4.68; the connection between the walking surface and the base
floor via the resilient material may be at individual points, continuous over their entire surface,
or along lines (i.e. along battens/joists). These resilient connections result in a mass–spring
type resonance associated with the floating floor. The performance of any floating floor relies
on its isolation from the base floor; hence it is highly dependent on the quality of workmanship.
Floating floors are easily bridged with rigid connections such as nailing through a lightweight
floating floor, screed pouring through gaps in the resilient layer, or when walking surfaces are
tightly butted up against the side walls.
4.4.4.1 Heavyweight base floors
When the base floor is much heavier than the floating floor, the mass–spring resonance frequency for the walking surface and the resilient layer modelled by a mass–spring system is
approximately the same as for a mass–spring–mass system which includes the base floor.
As with cavity walls the design aim is usually to have this resonance well-below the important
frequency range as it tends to have an adverse effect on L in the frequency band in which it
falls (i.e. negative values of L).
4.4.4.1.1 Resilient material as point connections
The SEA model described by Vér (1971) can be used to predict the improvement of impact
sound insulation when the plate that forms the walking surface supports a reverberant bending
wave field. This walking surface is connected to a heavyweight base floor by N resilient mounts
with a dynamic stiffness, k. To reduce the number of subsystems in the model it is assumed that
sound radiated into the cavity between the two plates is highly attenuated within the cavity by
an absorbent material. It is further assumed that this material does not connect the two plates
and transmit vibration between them; all transmission occurs via the mounts. This means that
there is no need to include the cavity as a subsystem in the model. The number of subsystems
can now be reduced to the bare minimum because we are only interested in the change in
sound pressure level in the receiving room. This is equal to the change in vibration level of
the base floor; hence we can also exclude the receiving room from the model. The result is a
517
S o u n d
Wd(1)
I n s u l a t i o n
Win(1)
ISO tapping
machine
Subsystem
1
Walking
surface
Floating floor
W21
Wd(2)
N resilient
mounts
W12
Subsystem
2
Base floor
Figure 4.69
Two-subsystem SEA model for a floating floor connected by resilient mounts.
two-subsystem SEA model where subsystem 1 is the plate that forms the walking surface and
subsystem 2 is the base floor (see Fig. 4.69).
This example shows the advantage of working in terms of energy and coupling loss factors
because the basic SEA model is the same as the one that has just been used for impact sound
insulation of a heavyweight base floor (Section 4.4.1). It is only the coupling loss factors and the
type of subsystem that has changed; hence the energy in the base floor is given by Eq. 4.111.
If the floating floor is now removed and the tapping machine is placed directly on the base floor,
the energy of the base floor is found from
(4.115)
Win(2) = ωη2 E2
We can assume that the total loss factor of the base floor is not increased by coupling to the
floating floor. This is reasonable for a heavyweight base floor that is connected to supporting
walls on all sides because the structural coupling losses to these walls are usually much higher.
The improvement of impact sound insulation is now given by the ratio of the base floor energy
without the floating floor to with the floating floor. Hence combining Eqs 4.111 and 4.115 gives
Win(2) η1
L = 10 lg
Win(1) η12
(4.116)
Equation 4.116 can now be expanded and simplified with the aid of a few more assumptions.
Firstly, the hammer impedance is assumed to be negligible compared to the driving-point
impedance of each of the plate subsystems. It is also assumed that both plates can be modelled
as infinite plates. The resilient mounts are modelled as springs for which vibration of the walking
surface is assumed to be transmitted to the base floor only by forces, rather than by moments.
The coupling loss factor, η12 , between the walking surface and the base floor is then given by
Eq. 4.87 for which the assumptions were previously discussed in Section 4.3.5.4.1. The final
simplifying assumption in calculating Eq. 4.87 is that the plate mobilities are sufficiently low that
|Y1 + Y2 + Yc |2 ≈ |Yc |2 =
518
ω2
k2
(4.117)
Chapter 4
This gives the improvement of impact sound insulation as
2
2.3ρs1
cL1 h1 η1 S1 ω3
L ≈ 10 lg
Nk 2
(4.118)
where k is the dynamic stiffness of each resilient mount (N/m) and N is the number of mounts.
Subscript 1 indicates properties of subsystem 1, the plate that forms the walking surface of the
floating floor.
Equation 4.118 is only appropriate above the mass–spring–mass resonance frequency of the
system (Eq. 4.89). Below this frequency it is reasonable to assume that L = 0 dB. At the
resonance frequency L can be negative and is usually in the range −10 dB ≤ L ≤ 0 dB.
The equation usefully shows that above the mass–spring–mass resonance frequency, L
increases at 30 dB/decade. The assumptions in the above derivation are suited to concrete
plates, and have been shown to give good agreement with resilient mounts such as cork (Vér,
1971). For such floors this model identifies ways of improving the impact sound insulation; e.g.,
by reducing the number of mounts, using a thicker plate for the walking surface, or by trying
to increase the total loss factor of the walking surface by increasing the internal damping of
the plate.
For more general designs of floating floor it is necessary to work back through the various
assumptions and revise the model accordingly. This is needed for floating floors where the
walking surface has a much lower mass per unit area and there is an empty cavity without
absorbent (such as for access floors used for cabling). Note that the model cannot be used for a
continuous resilient layer simply by increasing the number of mounts because of the assumption
that the vibration field at the connection points is uncorrelated. To make sure that consideration
is given to other sound transmission mechanisms it is convenient to return to the five-subsystem
model for a plate–cavity–plate system (Fig. 4.29); where subsystem 2 now forms the walking
surface and subsystem 4 is the base floor. The first step is to reconsider the power input
because more than one may now be necessary. For a structure-borne sound source such as
the ISO tapping machine or machinery mounted on the floor, the power is injected directly
into subsystem 2. However, for point excitation of a lightweight plate there may be significant
sound power radiated by the nearfield (Section 2.9.7); this can be included as a power input
directly into the cavity subsystem. In addition, structure-borne sources tend to generate some
airborne noise from their mechanical parts, so it may be necessary to have another power
input for sound power radiated into subsystem 1 (the room in which the structure-borne sound
source is operating). For most walking surfaces with a low mass per unit area, calculation of the
power input from the ISO tapping machine requires consideration of the interaction between the
hammer and the walking surface. Rather than using Fourier transforms as in Section 3.6.3.1,
estimates can be made using Eq. 4.112 but without making the approximation that the hammer
impedance is negligible. The structural coupling loss factor for the resilient mounts can then
be calculated according to Eq. 4.87 without making the approximation in Eq. 4.117. As the
resilience of the mounts is increased, the path involving structural coupling between the plates
will gradually become less important than the transmission path via the sound field in the
cavity. Path analysis can then be used to try and optimize the design. For structural reasons
the mounts may not be particularly resilient. The model can still be used for walking surfaces
with rigid mounts by setting Yc = 0, although the floor is no longer ‘floating’.
If the walking surface of any type of floating floor does not completely cover the base floor
then sound radiated by the walking surface back into the source room can subsequently
519
S o u n d
I n s u l a t i o n
excite the base floor which radiates sound into the receiving room. This transmission path,
2 → 1 → 4 → 5, is conveniently assessed using an SEA model and becomes more important
with resilient layers that are dynamically soft. For lightweight floating floors it is usually more
important in the mid- and high-frequency ranges near the critical frequency of the walking
surface.
4.4.4.1.2 Resilient material over entire surface
One example of this type of floor is a screed on top of a continuous resilient layer. This kind
of floating floor can be modelled according to Cremer (1952) where the walking surface and
the base plate act as homogeneous infinite plates and the resilient layer acts as a series
of closely spaced springs. The lump spring model that is used to transmit forces from the
walking surface to the base floor means that wave motion in the resilient layer is not included.
This approach allows use of the measured dynamic stiffness for the resilient layer (Section
3.11.3). It is further assumed that the walking surface and the base floor are formed from
heavyweight plates with similar density and similar Young’s modulus. In addition, the walking
surface is taken to be thinner than the heavyweight base floor. Excitation by a point force (i.e.
the hammer of the ISO tapping machine) excites a propagating bending wave on the walking
surface that subsequently excites bending waves on the base floor via the resilient layer. The
full derivation is given in Cremer et al., (1973) and uses two coupled bending wave equations to
describe the motion of the two plates. Despite the complexity of the transmission process, the
resulting equation is remarkably compact. Under the assumption that the hammer impedance
is negligible compared to the driving-point impedance of the plates, the improvement of impact
sound insulation is given by (Cremer et al., 1973)
f
L = 40 lg
(4.119)
fms
for which the resonance frequency (as with a mass–spring system) is given by
1 s′
fms =
2π ρs
(4.120)
where s′ is the dynamic stiffness per unit area (Section 3.11.3.1) and ρs is the mass per unit
area of the walking surface.
If the dynamic stiffness for a specific thickness is not known or a number of different materials
are combined together, a rough estimate for N resilient layers on top of each other can be
calculated from
−1
N
1
′
s =
(4.121)
sn′
n=1
This type of floor may use a lightweight plate to form the walking surface. In this case the
hammer impedance may no longer be negligible compared to the driving-point impedance as
the frequency increases. Account therefore needs to be taken of the reduction in power input
above the limiting frequency, flimit (Eq. 3.106). Assuming that the walking surface still acts as
an infinite plate, this gives L as (Cremer et al., 1973)
&
2 '
f
ωm
L = 40 lg
(4.122)
+ 10 lg 1 +
fms
2.3ρcL h2
520
Chapter 4
70
Measured: Unloaded
Improvement of impact sound insulation (dB)
60
Measured: Loaded (22 kg/m2)
50
Predicted (accounting for impedance
of the ISO tapping machine hammer)
40
30
fms ⫽ 83 Hz
flimit ⫽ 521 Hz
20
Walking surface: 22 mm chipboard
10
Resilient layer: s⬘ = 4 MN/m3
(45 mm reconstituted foam formed
from two layers of foam)
Base floor: 140 mm concrete slab
0
⫺10
50
80
125
200
315
500
800
1250
One-third-octave-band centre frequency (Hz)
2000
3150
Figure 4.70
Improvement of impact sound insulation (ISO tapping machine) due to a floating floor (lightweight walking surface on continuous
resilient layer) on a 140 mm concrete floor slab. Measurements according to ISO 140 Part 8. Measured data from Hopkins
and Hall are reproduced with permission from BRE Trust.
where ρ, cL , h correspond to the plate that forms the walking surface and m is the mass of the
ISO tapping machine hammer (0.5 kg).
For a walking surface with a driving-point impedance, Zdp , Eq. 4.122 can also be written in
terms of the limiting frequency above which the power input starts to decrease with increasing
frequency (Cremer et al., 1973)
&
2 '
f
f
L = 40 lg
+ 10 lg 1 +
(4.123)
fms
flimit
Equations 4.119, 4.122, 4.123 for L are valid above fms and have been found to give reasonable estimates in the frequency range: fms < f < 4fms (Cremer et al., 1973). When calculating
single-number quantities this is usually the most important part of the building acoustics
frequency range. Equation 4.119 indicates that L increases by 40 dB/decade (recall that
30 dB/decade was predicted for a reverberant plate on individual resilient mounts). An example
for a lightweight plate as the walking surface is shown in Fig. 4.70 to illustrate and aid discussion
on a few points. For prediction purposes it is simplest to assume that L = 0 dB in all frequency
bands below the band containing fms . A dip may occur in the frequency band containing fms ,
and sometimes in adjacent bands too. The uncertainty in predicting the frequency band with a
negative L is typically plus or minus one frequency band, as seen in this example. The depth
of the mass–spring dip is highly variable; in addition the dip may become shallower, deeper,
or remain unchanged when an additional static load is applied to the floor. A rule of thumb for
the frequency band containing fms is that L will fall in the range −5 dB ≤ L ≤ 0 dB. In the
high-frequency range the resilient layer usually supports wave motion and the resilient layer
can no longer be modelled as a simple lump spring element.
521
S o u n d
I n s u l a t i o n
In practice, not all floating floors with continuous resilient layers have a single slope of
40 dB/decade. This is mainly because the walking surface and the base floor do not act as
infinite plates and with increasing frequency the resilient layers no longer act as simple springs.
The infinite plate assumption is now considered for a floating screed and a structural concrete
floor slab. These often have similar material properties and the former is usually thinner than
the latter by at least a factor of three. Structural slabs sometimes act as infinite plates when
they span complete floors and show a noticeable decrease in vibration level with distance.
Floating screeds tend to be cast in each room. Therefore they are smaller, and by their very
nature they must be isolated from the walls at the edges; hence any structural coupling losses
to these walls are negligible and the bending waves are reflected. If the screed has high internal
damping, the amplitude of the propagating bending waves may have sufficiently decreased
with distance that reflections from the boundaries can be neglected; hence it can be treated
as an infinite plate. High internal damping has been found to occur with asphalt screeds for
which Eq. 4.119 can give good agreement with measurements (Cremer et al., 1973). However, this is not the case with many other screeds such as those formed from sand–cement
where the internal loss factor is low; these screeds usually act as finite plates with a reverberant bending wave field. For many floating screeds, Eq. 4.119 tends to overestimate L and
the frequency dependence is better described by a 30 dB/decade slope; hence the following
empirical solution is commonly used (EN 12354 Part 2)
f
L = 30 lg
(4.124)
fms
Single-number quantities predicted using Eq. 4.124 show good agreement with measurements
on floating screeds (Metzen, 1996). However, in the mid- and high-frequency ranges the slope
can be shallower than 30 dB/decade; sometimes tending towards a plateau as shown by the
example in Fig. 4.71 (Gudmundsson, 1984a). This can make it harder to identify minor bridging
defects with concrete screeds because a reduction in L due to bridging tends to occur in the
mid- and high-frequency ranges (Cremer et al., 1973; Villot and Guigou-Carter, 2003).
At low frequencies where the resilient layer does not support wave motion, it can be treated
as a lump spring element. Resilient layers used under floating floors are typically between 5
and 50 mm thick for which most materials can be treated as a lump spring in the low-frequency
range. Even when this is appropriate there are other factors that can make this spring less
than simple. From Section 3.11.3.1 we recall that for porous resilient materials the dynamic
stiffness is the combined stiffness of the skeletal frame and the air contained within it. For
mineral wool, the dynamic stiffness and loss factor of the skeletal frame can vary non-linearly
with static load (Gudmundsson, 1984b), and non-linearly with strain amplitude (Pritz, 1990).
When the dynamic stiffness of a porous resilient layer is predominantly determined by the
dynamic stiffness of the contained air, these non-linear effects for the skeletal frame tend to be
less significant. However, there is a wide variety of resilient materials and there will inevitably
be some materials that have the same dynamic stiffness in laboratory measurements but act
as springs with a different dynamic stiffness under different walking surfaces. To keep this in
perspective it is important to note that for most floating floor designs, we get a reasonable
estimate of the frequency band that contains the mass–spring resonance dip and Eq. 4.119
or 4.124 gives an adequate estimate for L in the low-frequency range. In the mid- and highfrequency ranges, reverberant bending wave fields on the walking surface and wave motion
in the resilient layer require more complex models (Gudmundsson, 1984a). Modelling wave
motion relies on properties of the resilient material that can be highly variable, are not usually
available, and are not particularly easy to model (e.g. mineral wool is anisotropic).
522
Chapter 4
Improvement of impact sound insulation (dB)
70
Resilient layer
60
30 mm (X ⫽ 1)
50
60 mm (X ⫽ 2)
120 mm (X ⫽ 4)
40
30
20
Walking surface: 50 mm dense concrete
Resilient layer: X layers of 30 mm rock
wool ( ρ = 162 kg/m3)
Base floor: 160 mm dense concrete slab
10
0
⫺10
50
80
125
200
315
500
800
1250 2000
One-third-octave-band centre frequency (Hz)
3150
5000
Figure 4.71
Improvement of impact sound insulation for a concrete screed on resilient layers of different thickness. Measurements
according to ISO 140 Part 8. Measured data are reproduced with permission from Gudmundsson (1984).
ρs2
s⬘2
ρs1
s⬘1
Figure 4.72
Mass–spring–mass–spring system on a rigid base.
If one floating floor is placed on top of another floating floor and there is a heavyweight base
floor, a mass–spring–mass–spring system is formed (see Fig. 4.72). This double floating floor
has two resonance frequencies given by
fmsms =
1
23/2 π
!
"
"
#
X±
X2 −
4s1′ s2′
ρs1 ρs2
where X =
s1′
s′
s′
+ 2 + 2
ρs1
ρs1
ρs2
(4.125)
Comparison of a single and a double floating floor is shown in Fig. 4.73 (Hopkins and Hall,
2006). A double floating floor can avoid the adverse dip in L that occurs with a single floating
floor at the mass–spring resonance frequency. However, because the double floating floor has
two resonance frequencies, the steep increase in L does not begin until frequencies above
the higher resonance frequency.
4.4.4.1.3 Resilient material along lines
A floating floor formed by a lightweight plate rigidly connected to a series of parallel timber
battens that rest upon a resilient layer is commonly referred to as a timber raft. Qualitatively
523
S o u n d
I n s u l a t i o n
Improvement of impact sound insulation (dB)
30
Walking surface: 18 mm plywood
Resilient layer: s⬘ ⫽ 7.25 MN/m3
(25 mm reconstituted foam)
Base floor: 140 mm concrete slab
25
20
As above but with one floating
floor on top of another to form
a double floating floor
15
10
5
Single floating floor
fms ⫽ 118 Hz
0
Double floating floor
fmsms ⫽ 74 Hz and 195 Hz
⫺5
⫺10
50
63
80
100
125
160
200
250
One-third-octave-band centre frequency (Hz)
315
400
Figure 4.73
Improvement of impact sound insulation (ISO tapping machine) due to a single or double floating floor on a 140 mm concrete
floor slab. Measurements according to ISO 140 Part 8. Measured data from Hopkins and Hall are reproduced with permission
from BRE Trust.
it is useful to consider the five-subsystem SEA model that was previously used for a plate–
cavity–plate system. For vibration transmission between the walking surface and the base plate
this model indicates that it is not only necessary to consider structural coupling via the battens
and resilient layer, but also via the sound field in the cavities. The latter path will be more
important when the resilient layer is dynamically soft. Using an SEA model to quantify vibration
transmission across a timber raft by treating the resilient layer as a lumped spring element has
been found to be difficult (Stewart and Craik, 2000). Other complexities also occur: there is a
different input power and different vibration transmission across the plate for hammers above
and in-between battens; the timber raft tends to act as a spatially periodic plate with a decrease
in vibration level with distance and there can be significant nearfield radiation into the cavity
directly beneath the ISO tapping machine.
4.4.4.2 Lightweight base floors
The improvement of impact sound insulation for a floating floor on a lightweight base floor is
generally specific to one type of base floor as there will usually be different sound transmission
mechanisms. Certain aspects such as the resonance frequency and driving-point mobility can
be used for basic design decisions, but calculation of L is less amenable to simple models.
For this reason, reliance is placed on laboratory measurements, particularly for heavy impact
sources. Examples for the improvement of impact sound insulation with a timber raft on different
base floors are shown in Fig. 4.74; note that this does not imply that similar trends exist for
all lightweight base floors and floating floors. When impact sound insulation is critical in the
low-frequency range, any conclusion about the efficacy of a floating floor not only depends on
the base floor but also upon the excitation. In this particular example the differences between
the base floors become particularly apparent in the mid- and high-frequency ranges.
524
Chapter 4
Improvement of impact sound insulation (dB)
35
Base floor
Base floor
30
A
A
25
B
B
20
C
C
D
D
E
E
15
F
10
5
0
ISO tapping machine
ISO rubber ball
⫺5
50
80
125 200 315 500 800 1250 2000 3150 5000 50
80
125
200
315
500
One-third-octave-band centre frequency (Hz)
One-third-octave-band centre frequency (Hz)
Floating floor: Timber raft
18 mm chipboard screwed to
timber battens (approximately 48 ⫻ 48 mm) at 400 mm centres
on 25 mm mineral wool ( ρ ⫽ 36 kg/m3, s⬘ ⫽ 11 MN/m3)
A
B
C
D
E
F
Figure 4.74
Improvement of impact sound insulation for a timber raft floating floor on different base floors measured using the ISO
tapping machine and ISO rubber ball. Measurements according to ISO 140 Part 11 (lightweight base floors) and ISO
140 Part 8 (heavyweight base floor). Base floor D corresponds to Reference floor No.2 in ISO 140 Part 11: 18 mm OSB,
40 × 240 mm timber joists at 400 mm centres, 100 mm mineral wool, resilient metal channels running perpendicular to the
joists, two layers of 15 mm plasterboard. Base floors A, B, and C use different components of base floor D. Base floor E
is comprised of 22 mm chipboard, light steel C-joists and 15 mm plasterboard. Base floor F is a 150 mm concrete floor
slab (ISO 140 Part 8 reference floor). Measured data from Hopkins and Hall are reproduced with permission from ODPM
and BRE.
525
S o u n d
I n s u l a t i o n
80
Measured: Radiated power
SEA: Radiated power
70
SEA: Power input
Artificial heavy rain
Rainfall rate: 40 mm/h
Median drop diameter: 5 mm
Fall velocity: ≈ 7 m/s
Surface area, S: 1.875 m2
Excitation area, Se: 1 m2
Sound power (dB)
60
50
40
fc ⫽ 2021 Hz
30
20
50
80
125
200
315
500
800
1250
2000
3150
5000
One-third-octave-band centre frequency (Hz)
Figure 4.75
Radiated sound power from artificial heavy rain on 6 mm glass. Plate properties: Lx = 1.5 m, Ly = 1.25 m, h = 0.006 m,
ρs = 15 kg/m2 , cL = 5350 m/s, ν = 0.24, measured total loss factor, radiation efficiency calculated assuming plate lies in an
infinite baffle. Plate orientated at an angle of 30◦ for water drainage. Sound intensity measurements according to ISO 15186
Part 1. Measured data from Hopkins are reproduced with permission from BRE.
4.5 Rain noise
The prediction of rain noise uses similar calculations to those already used for impact sound
insulation on floors with the ISO tapping machine. Calculation of the power input for artificial
and natural rainfall has previously been described in Section 3.7.1. Referring back to the SEA
assumptions in Section 4.2, excitation from artificial rain is quite well-suited to the requirement
for statistically independent excitation forces. The power input can therefore be used in an SEA
model or other types of models for multi-layered plates (Guigou-Carter et al., 2002; GuigouCarter and Villot, 2003).
In a similar way to impact sound insulation for a homogenous plate (Section 4.4.1) we can
use the same two-subsystem SEA model to calculate the reverberant sound pressure level in
the room below the element in the roof that is being excited by rain (either artificial rain or an
idealized model for natural rain). This can be converted to a sound power level for the element
if required.
A comparison of measured and predicted sound power from artificial heavy rain on 6 mm
glass is shown in Fig. 4.75. The power input from the artificial rain is also shown because
the predicted decrease that initially occurs above the cut-off frequency for the force pulse is
not observed in the measurements; this can be seen by the absence of any dip in the measurement between 1000 and 1600 Hz. However, in general there is close agreement between
measurement and prediction. This suggests that any change to the power input due to drops
526
Chapter 4
falling on the existing layer of water that is running down the glass can reasonably be ignored
in most practical calculations.
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Chapter 5
Combining direct and flanking
transmission
5.1 Introduction
F
lanking transmission is omnipresent in buildings and its effects are not confined to any
particular part of the building acoustics frequency range. In fact it is not uncommon for
the flanking structure to radiate similar or higher sound power levels than the separating
wall or floor itself.
Figure 5.1 shows field measurements of airborne sound insulation for solid brick separating
walls in dwellings with a variety of different flanking walls and floors. The variation is not only
due to the different flanking constructions but also due to variation in workmanship, material
properties, room dimensions, and measurement uncertainty. Sound insulation distributions of
single-number quantities tend to be skewed towards lower values of sound insulation. The
lower bound is only limited by the ingenuity of the builder to alter sound transmission paths and
introduce new transmission paths through the quality of workmanship as well as by substitution
of different materials to those that were specified. The wide range of results for the nominally
identical separating wall illustrates the importance of understanding how the combination of
direct and flanking transmission determines the in situ performance.
This chapter looks at general principles of predicting direct and flanking transmission based
on the well-established use of Statistical Energy Analysis (SEA) models for predicting sound
transmission in buildings (Craik, 1996; Cremer, et al., 1973; Gerretsen, 1979, 1986; Kihlman,
1967). It starts by looking at vibration transmission between plates connected at a junction. An
example building plan showing some plate junctions is shown in Fig. 5.2. The principles apply
to both lightweight and heavyweight constructions. However many lightweight walls and floors
do not act as simple reverberant subsystems, hence the examples will focus on heavyweight
walls and floors, and issues relating to low mode counts and low modal overlap. The chapter
then moves on to look at the general application of SEA models and SEA-based models (EN
12354) to the prediction of sound insulation.
Flanking transmission significantly increases the complexity of prediction. With so many transmission paths there is the potential to predict what appears to be the correct sound insulation
in situ by using a model that does not adequately describe the actual sound transmission process. It is equally possible to inadvertently compensate for errors in one part of a model with
invalid assumptions in another part of the model. For this reason, emphasis is placed on the
many assumptions and limitations that are involved in the prediction of flanking transmission.
These need to be kept firmly in perspective otherwise it is all too easy to disregard the possibility of success despite all evidence to the contrary. A degree of pragmatism is required. If the
aim is to consistently predict the sound insulation between a specific pair of rooms in a specific
building to an accuracy of ±1 dB in each frequency band across the entire building acoustics frequency range, then disappointment will generally follow. Realistic aims are to estimate
the average sound insulation for many similar constructions and to gain sufficient insight into
535
S o u n d
I n s u l a t i o n
Standardized level difference, DnT (dB)
70
60
50
40
30
100 125 160 200 250 315 400 500 630 800 1000 1250 1600 2000 2500 3150
One-third-octave-band centre frequency (Hz)
24
22
20
18
Count
16
14
12
10
8
6
4
2
0
45
46
47
48 49 50 51 52 53 54 55 56 57
Weighted standardized level difference, DnT,w (dB)
58
59
60
Figure 5.1
Field measurements from ninety-one solid brick separating walls (≈215 mm brick wall with ≈13 mm plaster finish on both
sides) with different flanking constructions. The external flanking walls are masonry cavity walls but with different materials,
cavity depths and wall ties; other flanking walls and floors varied. Room volumes were between 25 and 60 m3 . Measured data
are reproduced with permission from ODPM & BRE.
important sound transmission mechanisms that will allow design decisions to be made, and
help identify solutions to sound insulation problems in existing buildings.
5.2 Vibration transmission across plate junctions
When predicting vibration transmission across junctions of walls and floors there is more than
one model that can be used, each of which has limitations in its application to real buildings.
536
Chapter 5
T-junction
Separating
wall
X-junction
Figure 5.2
Example plate junctions in buildings.
Prediction is simplest when each plate that forms a wall or floor can be modelled as a solid,
homogeneous, isotropic plate using thin plate theory. Whilst this is often a reasonable assumption there are many walls and floors that require more complex models. Just as with direct sound
transmission in Chapter 4, idealized models are beneficial in providing an insight into the process of vibration transmission and they form a useful benchmark against which more complex
constructions can be compared.
Predictions using a wave approach based on semi-infinite plates are commonly used under
the assumption that all the plates have diffuse vibration fields. This assumption is not always
appropriate, so use is also made of modal approaches and numerical methods. With any
approach it is usually possible to model the junction so that either there is only bending wave
motion on each plate, or bending and in-plane wave motion is possible on each plate. The
output from any of these calculations can be converted into coupling parameters that can be
used in SEA or SEA-based models.
537
S o u n d
I n s u l a t i o n
In this section we start with an overview of the wave approach that considers only bending
waves, move on to a wave approach that considers both bending and in-plane waves, and
finally look at Finite Element Methods (FEM) as one example of a numerical method. (For
examples of other numerical approaches, see Cuschieri, 1990; Guyader et al., 1982; Rébillard
and Guyader, 1995.) To predict sound transmission between adjacent rooms it is usually
sufficient to use a wave approach that only considers bending waves. For non-adjacent rooms
that are some distance apart (i.e. with structure-borne sound transmission across a few or
several junctions) it is necessary to consider both bending and in-plane waves. To gain an
insight into some of the more awkward issues concerning the measurement and prediction of
vibration transmission it is necessary to compare these different approaches in this section;
this is done using SEA and numerical experiments with FEM.
Plate junctions in buildings can generally be described as one of four idealized types:
1.
2.
3.
4.
a rigid junction where the plates are rigidly connected together
a rigid junction where the plates are all rigidly connected to a beam/column
a resilient junction where one or more plates are connected via a resilient material
hinged junctions.
In this chapter the focus is on rigid junctions because these are applicable to many junctions
of brick/block/concrete walls and floors in buildings. Idealized models of junctions are very
useful tools but in practice there is more than one way of forming a rigid connection between
plates; hence measurements are sometimes used to quantify vibration transmission. This
can be fraught with difficulty due to the variation in workmanship that occurs at junctions of
nominally identical walls and floors, measurement uncertainty and the effect of different plate
properties and dimensions on vibration transmission. The examples using FEM, SEA, and
measurements in this section illustrate the features that need to be considered when relying
purely on measurements.
5.2.1 Wave approach: bending waves only
Junctions of walls and floors can often be described by an X, T, L or in-line junction as shown
in Fig. 5.3. In the following derivations it is assumed that the plates forming these junctions are
solid, homogeneous, isotropic plates modelled with thin plate theory.
In the same way that airborne sound insulation is described using the sound reduction index; we
can make use of a transmission coefficient to describe structure-borne sound transmission.
However with plate junctions there are more permutations for the source plate, i, and the
receiver plate, j. The transmission coefficient, τij , is therefore defined as the ratio of the power,
Wij , that is transmitted across the junction to plate j, to the power, Winc,i , that is incident on the
junction on plate i,
τij =
Wij
Winc,i
(5.1)
which in decibels gives a transmission loss, TLij , as
TLij = 10 lg
538
1
τij
(5.2)
Chapter 5
Plate 2
Plate 2
X-junction
L-junction
Plate 1
Plate 3
Plate 1
Plate 4
T-junction (1)
T-junction (2)
Plate 2
Plate 2
Plate 3
Plate 1
Plate 1
Plate 4
In-line junction
Plate 2
Plate 1
Figure 5.3
Common plate junctions. The hatched box represents the junction beam used for the purpose of modelling vibration
transmission with a wave approach.
For use in SEA models, the coupling loss factor, ηij , is calculated from the transmission
coefficient, τij , as described in Section 2.6.4.
5.2.1.1 Angular averaging
The wave approach considers a plane wave on plate i that impinges upon a junction at a specific
angle of incidence, θ, as shown in Fig. 5.4. This gives an angle-specific transmission coefficient,
τij (θ). In practice we are not interested in the transmission coefficient at a specific angle but in
the angular average transmission coefficient. In a diffuse vibration field it is assumed that all
angles of incidence are equally probable. As the intention is to use the transmission coefficient
within the framework of SEA, we note that the SEA assumption of equipartition of modal energy
on each plate relates to the assumption that the incident energy is uniformly distributed in angle.
The intensity that is incident upon the junction length, Lij , is proportional to Lij cos θ, hence the
power transmitted from plate i to plate j is (Cremer, 1973)
π/2
(5.3)
τij (θ)Lij cos θ dI(θ)
Wij =
−π/2
539
S o u n d
I n s u l a t i o n
Incident
wave
θ
Lij
Reflected
wave
Transmitted
wave
Plate i
Plate j
Junction
line
dθ
θ
Lij
Figure 5.4
Plane wave incident at an angle, θ , upon a junction, giving rise to a reflected and a transmitted wave.
As the incident energy is uniformly distributed in angle, the intensity dI(θ) associated with dθ
(see Fig. 5.4) is given by
cg,i Ei dθ
Si 2π
dI(θ) =
(5.4)
Equations 5.3 and 5.4 now give the transmitted power as
Wij =
cg,i Ei Lij
πSi
π/2
τij (θ) cos θ dθ
(5.5)
0
and the angular average transmission coefficient, τij , is therefore given by
τij =
π/2
τij (θ) cos θ dθ
(5.6)
0
The angular average transmission coefficient is weighted towards values at normal incidence.
For this reason, normal incidence transmission coefficients are sometimes used to give rough
estimates of the angular average value; the reason is not that it is common for all incident
waves to be normally incident upon a plate junction.
540
Chapter 5
To reduce the number of calculations it is useful to note a relationship between the angular
average transmission coefficients, τij and τji . For a bending wave that is incident upon a junction
of homogeneous isotropic plates where only reflected and transmitted bending waves are
generated at the junction, the SEA consistency relationship (Eq. 4.2) links τij to τji using
fc,j
hi cL,i
= τji
(5.7)
τij = τji
hj cL,j
fc,i
5.2.1.2 Angles of incidence and transmission
In calculating the transmission coefficient over the range of possible angles of incidence it is
necessary to relate the angle of incidence to the angle of transmission. From Snell’s law of
refraction, the incident, reflected and transmitted waves must have the same spatial dependence along the junction line in terms of their wavenumber. The angle of incidence on plate i is
therefore related to the angle of transmission on plate j by
ki sin θi = kj sin θj
(5.8)
where k is the wavenumber on each plate; in this section we are only concerned with bending
waves but it equally applies to quasi-longitudinal or transverse shear waves.
The angle of transmission can only take real (not complex) values. Therefore when ki > kj there
must be a cut-off angle, θco . For angles of incidence greater than the cut-off angle there is no
transmitted wave and the transmission coefficient is zero. From Eq. 5.8 this cut-off angle is
given by
kj
θco = arcsin
(5.9)
ki
In this section we only consider bending waves and because the frequency-dependence of
the bending wavenumber is the same for all thin plates the cut-off angle does not vary with
frequency. However, when both bending and in-plane waves are considered (Section 5.2.2),
the cut-off angle can be frequency-dependent.
5.2.1.3 Rigid X, T, L, and in-line junctions
In many buildings there is a degree of symmetry between adjacent rooms; hence some of the
plates that form the junction will be identical in terms of their thickness and material properties.
This conveniently reduces the number of variables that are needed to calculate the transmission
coefficients. Referring to the plate junctions in Fig. 5.3, the following assumptions apply:
1. X-junction: plates 1 and 3 are identical, and plates 2 and 4 are identical
2. T-junction (1): plates 1 and 3 are identical
3. T-junction (2): plates 2 and 4 are identical.
This allows the transmission coefficients to be calculated using the variables χ and ψ (Cremer
et al., 1973)
kB2
h1 cL1
fc2
4 ρs2 B1
χ=
=
=
=
(5.10)
kB1
ρs1 B2
h2 cL2
fc1
ψ=
2
B2 kB2
2
B1 kB1
=
ρs2 fc1
h2 cL2 ρs2
=
h1 cL1 ρs1
ρs1 fc2
(5.11)
541
S o u n d
I n s u l a t i o n
The wave approach used to model an incident bending wave that results in only reflected and
transmitted bending waves is thoroughly covered in the book by Craik (1981); this also contains
tabulated data for bending wave transmission across different junctions of plates and beams. In
this section a short qualitative discussion is given along with the main equations needed to calculate the transmission coefficients for the most common junctions; X, T, L, and in-line junctions.
To model the rigid junction the plates are connected by a junction beam (see Fig. 5.3). This
beam has no mass, does not support wave motion and has a rigid cross-section. We can
now define the conditions that result in an incident bending wave only giving reflected and
transmitted bending waves (not in-plane waves). This occurs when the junction beam is simply
supported (sometimes referred to as being pinned) so that it cannot undergo displacement, but
is free to rotate. Hence when a bending wave is incident upon the junction, the junction beam
can ‘transfer’ rotation to the other plates, this will only give rise to bending waves. The junction
beam is used to define equilibrium and continuity conditions that must be satisfied for this to
occur. The equilibrium conditions ensure that the sum of the moments acting on the junction
beam is zero. The continuity conditions ensure that the displacement of the beam and each
plate along their line of connection is zero, and that the rotation of the beam and the plates
is equal. The resulting transmission coefficients are independent of frequency; this simplifies
their calculation as well as simplifying further calculations in SEA or SEA-based models.
For the junctions of perpendicular plates shown in Fig. 5.3 it is convenient to refer to ‘transmission around a corner’ and ‘transmission across a straight section’. To calculate the transmission
coefficients we assume that there is an incident bending wave on plate 1. For the X- and the
L-junction, any of the plates can be chosen as plate 1 to calculate the transmission coefficients.
For T-junctions, transmission around the corner depends on the relative orientation of the other
plates and we need to refer to T-junctions (1) and (2); it is not necessary to make this distinction
for transmission across the straight section of a T-junction.
Using the plate numbering system in Fig. 5.3 with an incident bending wave on plate 1,
transmission around the corner of any rigid X- , T- or L-junction is given by (Craik, 1981, 1996).
If χ ≥ sin θ, then
0.5J1 J2 ψ cos θ χ2 − sin2 θ
τ12 (θ) =
(J2 ψ)2 + χ2 + J2 ψ
1 + sin2 θ χ2 + sin2 θ + 1 − sin2 θ χ2 − sin2 θ
(5.12)
else if χ < sin θ, then
τ12 (θ) = 0
where the constants J1 and J2 depend on the junction. For X-junctions, J1 = 1 and J2 = 1. For
T-junction (1), J1 = 2 and J2 = 0.5. For T-junction (2), J1 = 2 and J2 = 2. For L-junctions, J1 = 4
and J2 = 1.
For an incident bending wave on plate 1, transmission across a straight section of a rigid
X-junction or the straight section of rigid T-junction (1) is given by (Craik, 1981, 1996).
If χ ≥ sin θ, then
τ13 (θ) =
542
(J3 ψ)2 + χ2 + J3 ψ
0.5χ2 cos2 θ
1 + sin2 θ χ2 + sin2 θ + 1 − sin2 θ χ2 − sin2 θ
(5.13)
Chapter 5
else if χ < sin θ, then
τ13 (θ) =
where
cos2 θ
(J3 ψ) C
2J3 ψC
2+
1 + sin2 θ
+
4
χ
χ2
2
2
C = χ2 + sin2 θ + sin2 θ − χ2
and the constant J3 depends on the junction. For X-junctions, J3 = 1. For T-junction (1),
J3 = 0.5.
For an in-line junction, the normal incidence transmission coefficient gives estimates within 1 dB
of the angular average value when χ ≥ 1. When χ < 1, calculation in the reverse direction can
be carried out first and then Eq. 5.7 can be used to give the value in the other direction (Craik,
1996). For a bending wave on plate 1 at normal incidence upon the junction line, transmission
across a rigid in-line junction is given by (Cremer et al., 1973)
τ12 ≈ τ12 (0◦ ) =
5.2.1.3.1
√
2
2(1 + χ)(1 + ψ) χψ
χ(1 + ψ)2 + 2ψ(1 + χ2 )
(5.14)
Junctions of beams
The calculations for plate junctions in Section 5.2.1.3 can be adapted to beams. Calculations
for junctions of beams are simpler because all waves are at normal incidence, so only the
normal incidence transmission coefficient is used and no angular averaging is needed. The
variables χ and ψ are calculated from Eqs 5.10 and 5.11 using the bending wavenumbers for
beams, the beam bending stiffness and the mass per unit length.
Note that timber studs and joists that form beams in timber-frame constructions may need to
be modelled as hinged rather than rigid junctions (Craik and Galbrun, 2005).
5.2.2
Wave approach: bending and in-plane waves
Compared to the wave approach for only bending waves, the inclusion of in-plane waves
increases the complexity of the solution, as well as subsequent computation of the transmission coefficients. Major steps in applying the wave approach to bending and in-plane wave
generation at plate junctions were made in the work of Cremer et al. (1973) and Kihlman
(1967). Subsequent increases in computing power aided further advances in application and
development of the theory with other types of plate junctions (Craven and Gibbs, 1981; Gibbs
and Gilford, 1976; Langley and Heron, 1990; Wöhle et al., 1981). In general there are two
aspects being considered simultaneously; the physics of the problem, and numerical computation. There are advantages in approaching a derivation from one or the other viewpoint as it
is difficult to do justice to both aspects simultaneously. The derivation in this section only gives
a brief overview of the theory to try and simplify subsequent computation. It is based on the
notation and derivation given by Mees and Vermeir (1993) and Bosmans (1998); this makes a
clear link back to the theory given by Cremer et al. (1973).
An example junction is shown in Fig. 5.5 where there are four plates connected by a junction
beam. The derivation is general and equally applies to two, three, or four plates at any angle to
each other. Hence it can be used for L, T, X and in-line junctions. The coordinate systems used
543
S o u n d
I n s u l a t i o n
Global coordinate system
z
x
y
ep
Fyp
Mzp
Fxp
ζb
ζp
ξb
ξp
αzb
ηb
Fzp
ηp
αzp
Local coordinate system
on each plate
z
x
y
Figure 5.5
Exploded view of a X-junction indicating the variables used to describe wave motion on each plate (subscript p) and on the
junction beam (subscript b). The z-axis of the global coordinate system is aligned along the centre line of the junction beam.
Global coordinate system x
Junction
beam
θp
Plate, p
y
Local coordinate system
y
x
Figure 5.6
Local and global coordinate system for each plate connected to the junction beam.
to describe each plate and their position relative to the junction beam are shown in Fig. 5.6.
Wave motion on each plate is described using a local coordinate system where each plate lies
in the xz plane. Each plate that forms the junction is then connected to the junction beam using
the global coordinate system.
544
Chapter 5
The plates are assumed to be solid, homogeneous, and isotropic, and are modelled using
thin plate theory. The model uses a junction beam to connect the plates together, but it does
not represent a physical part of the real junction. The junction beam has no mass, does not
support wave motion and has a rigid cross-section. This beam is free to rotate and to undergo
displacement in the three coordinate directions; this allows generation of in-plane waves at the
junction. With each plate there is an offset, ep , from the junction beam. For rigid junctions this
offset can be set to zero; note that masonry/concrete walls or floors can be rigidly connected
together in more than one way, and there is no clear way of exactly defining any offset. However,
it is included here because the offset does have a physical meaning when altering this basic
model to account for a junction beam with mass and stiffness that supports wave motion,
resilient layers between one or more plates and the junction beam, or a hinged junction (see
Bosmans, 1998; Mees and Vermeir, 1993).
Incident waves can be bending, quasi-longitudinal, or transverse shear waves that travel in the
negative x-direction towards the junction at x = 0 (local coordinates). For any incident wave
with unit amplitude that impinges upon the junction at an angle of incidence, θi , the general
form for the wave is
exp(iki x cos θi ) exp(−iki z sin θi )
(5.15)
where ki is the wavenumber.
This gives the general form for transmitted waves as
T exp(−ikx x) exp(−iki z sin θi )
(5.16)
where T is the complex amplitude.
Note that the time-dependence, exp(iωt), has been excluded for brevity, and that although the
z-dependence is needed to derive the plate and beam parameters at the junction (x = 0) it is
omitted wherever it eventually cancels out in the final set of equations.
5.2.2.1 Bending waves
The general equations for bending wave motion are given in Section 2.3.3. However to accommodate the number of subscripts that are needed here, the subscript notation is slightly
different. The variables are shown in Fig. 5.5 for one of the plates that form the junction. The
subscript p identifies each plate that forms the junction. For lateral displacement, η, associated
with bending wave motion, there is rotation about the z-axis by an angle, αz , given by
αzp =
∂η
∂x
(5.17)
a bending moment per unit width,
Mzp = −B
∂2 η
∂2 η
+
ν
∂x 2
∂z 2
(5.18)
and a shear force per unit width,
Fyp = −B
∂3 η
∂3 η
+ (2 − ν)
∂x 3
∂x∂z 2
(5.19)
545
S o u n d
5.2.2.1.1
I n s u l a t i o n
Incident bending wave at the junction
For an incident bending wave on plate p = 1 that is described by the general form in Eq. 5.15
and impinges upon the junction at x = 0, the displacement, rotation (Eq. 5.17), bending moment
(Eq. 5.18) and shear force (Eq. 5.19) are given by
(5.20)
η1i = 1
α1i = ik B1 cos θi
(5.21)
2
5.2.2.1.2
2
Mz1i = B1 kB1
( cos2 θi + ν1 sin θi )
(5.22)
3
cos θi [cos2 θi + (2 − ν1 ) sin2 θi ]
Fy1i = iB1 kB1
(5.23)
Transmitted bending wave at the junction
For the transmitted bending wave propagating on plate p in the positive x-direction, the
displacement is given by
ηp (x, z) = [TBNp exp(−ikBpx1 x) + TNp exp(−ikBpx2 x)] exp(−iki z sin θi )
(5.24)
where TBNp (Bending wave/Nearfield) and TNp (Nearfield) are complex amplitudes.
Solutions to the wave equation lead to either real or imaginary values for the wavenumber,
kBpx1 , depending on the angle of incidence; but only give imaginary values for the wavenumber,
kBpx2 . To interpret Eq. 5.24 in terms of propagating bending waves and nearfields we use Snell’s
law (Eq. 5.8) to relate the angle of incidence to the angle at which waves are transmitted.
Snell’s law gives a cut-off angle, θco (Eq. 5.9) for the angle of incidence; at larger angles
there is no transmitted propagating wave. Therefore the first term in the square brackets of
Eq. 5.24 corresponds to a propagating bending wave with complex amplitude, TBNp , when
θi ≤ θco but changes to a nearfield when θi > θco . The second term corresponds to a nearfield
with complex amplitude, TNp , regardless of the angle of incidence. Substituting the general form
for a transmitted wave into the bending wave equation gives the wavenumbers as follows:
2
kBp
− ki2 sin2 θi
2
= −i ki2 sin2 θi − kBp
If θi ≤ θco , then kBpx1 =
else, if θi > θco then kBpx1
(5.25)
2
kBpx2 = −i kBp
+ ki2 sin2 θi
(5.26)
The displacement, rotation, bending moment and shear force for the transmitted bending wave
at x = 0 can now be written in terms of TBNp and TNp ,
(5.27)
ηp = TBNp + TNp
(5.28)
αzp = −ik Bpx1 TBNp − ik Bpx2 TNp
Mzp =
2
Bp TBNp (kBpx1
+
νp ki2
2
sin θi ) +
2
Bp TNp (kBpx2
+
νp ki2
2
sin θi )
(5.29)
2
+ (2 − νp )ki2 sin2 θi ]
Fyp = −iBp kBpx1 TBNp [kBpx1
2
+ (2 − νp )ki2 sin2 θi ]
− iBp kBpx2 TNp [kBpx2
546
(5.30)
Chapter 5
To simplify computation and calculations of wave intensities it is useful to rename the various
terms in Eqs 5.27–5.30 as constants C1 and C2 (Bosmans, 1998)
⎡
⎤ ⎡
⎤
ηp
C1ηp C2ηp &
'
⎢
⎥ ⎢
⎥
⎢ αzp ⎥ ⎢ C1αzp C2αzp ⎥ TBNp
(5.31)
⎢
⎥=⎢
⎥
⎣Mzp ⎦ ⎣C1Mzp C2Mzp ⎦ TNp
Fyp
C1Fyp C2Fyp
5.2.2.2 In-plane waves
Bending wave motion could be dealt with in isolation, but the in-plane wave motion is linked to
both quasi-longitudinal and transverse shear waves. The in-plane displacements, ξ and ζ, are
shown in Fig. 5.5, these are related by two coupled equations of motion (Cremer et al., 1973)
2
2
E
∂ ξ
∂ ξ
∂2 ζ
∂2 ζ
E
∂2 ξ
+ν
+
+
−ρ 2 =0
(5.32)
2
2
2
1−ν
∂x
∂x∂z
2(1 + ν) ∂z
∂x∂z
∂t
2
2
∂ ζ
∂ ζ
∂2 ξ
∂2 ξ
E
∂2 ζ
E
(5.33)
+
ν
+
=0
+
−
ρ
1 − ν2 ∂z 2
∂x∂z
2(1 + ν) ∂x 2
∂x∂z
∂t 2
From Cremer et al. (1973) the wave fields are described using a potential, , and a stream
function, ,
(x, z) = [+ exp(ik L x cos θ) + − exp(−ik L x cos θ)] exp(−ik L z sin θ)
(5.34)
(x, z) = [+ exp(ik T x cos θ) + − exp(−ik T x cos θ)] exp(−ik T z sin θ)
(5.35)
from which the in-plane displacements are given by
∂ ∂
+
∂x
∂z
∂ ∂
−
ζ=
∂z
∂x
ξ=
(5.36)
(5.37)
and the equations of motion are
+ kL2 = 0 for quasi-longitudinal waves
+ kT2 = 0 for transverse shear waves
(5.38)
(5.39)
where is the harmonic operator.
The in-plane waves give rise to a normal force, Fx , and an in-plane shear force, Fz , (see
Fig. 5.5) which are related to the in-plane displacements by (Timoshenko and WoinowskyKrieger, 1959)
Eh
∂ζ
∂ξ
Fx =
+ν
(5.40)
1 − ν2 ∂x
∂z
∂ξ
∂ζ
+
Fz = Gh
(5.41)
∂z
∂x
5.2.2.2.1
Incident quasi-longitudinal wave at the junction
A unit amplitude quasi-longitudinal wave on plate p = 1, that is described by the general form
in Eq. 5.15 is given by
(x) = exp(iki x cos θi ) exp(−iki z sin θi )
(5.42)
547
S o u n d
I n s u l a t i o n
When this wave impinges upon the junction at x = 0, the displacements (Eqs 5.36 and 5.37),
normal force (Eq. 5.40) and in-plane shear force (Eq. 5.41) are given by
ξ1i = ik i cos θi
(5.43)
ζ1i = −ik i sin θi
(5.44)
Fx1i =
−E1 h1 2
ki ( cos2 θi + ν1 sin2 θi )
1 − ν12
Fz1i = 2G1 h1 ki2 cos θi sin θi
5.2.2.2.2
(5.45)
(5.46)
Incident transverse shear wave at the junction
A unit amplitude transverse shear wave on plate p = 1 described by the general form in Eq. 5.15
is given by
(x) = exp(iki x cos θi ) exp(−iki z sin θi )
(5.47)
When this wave impinges upon the junction at x = 0, the displacements (Eqs 5.36 and 5.37),
normal force (Eq. 5.40) and in-plane shear force (Eq. 5.41) are given by
ξ1i = −ik i sin θi
(5.48)
ζ1i = −ik i cos θi
(5.49)
Fx1i = 2G1 h1 ki2 cos θi sin θi
Fz1i =
5.2.2.2.3
G1 h1 ki2 ( cos2 θi
(5.50)
2
− sin θi )
(5.51)
Transmitted in-plane waves at the junction
For the transmitted in-plane waves propagating on plate p in the positive x-direction, the
displacements are given by
p (x, z) = TLp exp(−ik Lpx x) exp(−ik i z sin θi )
(5.52)
p (x, z) = TTp exp(−ik Tpx x) exp(−ik i z sin θi )
(5.53)
where TLp and TTp are complex amplitudes.
Snell’s law gives the cut-off angle for the quasi-longitudinal and transverse shear waves so
their wavenumbers are given as follows:
2
kLp
− ki2 sin2 θi
2
then kLpx = −i ki2 sin2 θi − kLp
If θi ≤ θco , then kLpx =
else, if θi > θco
2
kTp
− ki2 sin2 θi
2
then kTpx = −i ki2 sin2 θi − kTp
(5.54)
If θi ≤ θco , then kTpx =
else, if θi > θco
548
(5.55)
Chapter 5
The displacements, normal force and in-plane shear force for the transmitted in-plane wave at
the junction (x = 0) can now be written in terms of TLp and TTp ,
ξp = −ik Lpx TLp − ik i sin θi TTp
(5.56)
ζp = −ik i sin θi TLp + ik Tpx TTp
(5.57)
Fxp =
−Ep hp TLp 2
(kLpx + νp ki2 sin2 θi ) − 2Gp hp kTpx ki sin θi TTp
1 − νp2
2
Fzp = −2Gp hp kLpx ki TLp sin θi − Gp hp TTp (ki2 sin2 θi − kTpx
)
(5.58)
(5.59)
As with a transmitted bending wave it is also useful to define constants C1 and C2; from
Eqs 5.56–5.59 these can be defined as (Bosmans, 1998)
⎡ ⎤ ⎡
⎤
C1ξp C2ξp & '
ξp
⎢ ⎥ ⎢
⎥
⎢ ζp ⎥ ⎢ C1ζp C2ζp ⎥ TLp
(5.60)
⎥
⎢ ⎥=⎢
⎣Fxp ⎦ ⎣C1Fxp C2Fxp ⎦ TTp
C1Fzp C2Fzp
Fzp
5.2.2.2.4
Conditions at the junction beam
Up till this point the focus has been on wave motion on each plate. To calculate the required
transmission coefficients for an incident wave on plate 1, the derived equations must satisfy equilibrium and continuity conditions at the junction beam. The equilibrium conditions in
Eqs 5.61–5.64 ensure that the sum of forces acting on the junction beam equals zero. To do this
it is necessary to make use of the offset, ep , as well as the angle, θp , between each plate and
the junction beam in the global coordinate system. The continuity conditions in Eqs 5.65–5.68
ensure continuity between plate and beam motion (displacement and rotation) along the line
that they are connected together; these equations are needed for each plate that forms the
junction.
The following set of equations are solved by converting them into matrix format and using a
matrix inversion to give the complex amplitudes (TBNp , TNp , TLp , TTp ) on each plate as well as
the junction beam parameters (ξb , ηb , ζb , αzb ). For use in SEA calculations it is sufficient to
carry out the calculations at the band centre frequencies:
(Fxp cos θp − Fyp sin θp ) + (Fx1i cos θ1 − Fy1i sin θ1 ) = 0
(5.61)
p
p
(Fxp sin θp + Fyp cos θp ) + (Fx1i sin θ1 + Fy1i cos θ1 ) = 0
−
p
Mzp +
p
ep Fyp + (e1 Fy1i − Mz1i ) = 0
p
Fzp + Fz1i = 0
(5.62)
(5.63)
(5.64)
ξp (+ξ1i if p = 1) = ξb cos θp + ηb sin θp
(5.65)
ηp (+η1i if p = 1) = −ξb sin θp + ηb cos θp + ep αzb
(5.66)
ζp (+ζ1i if p = 1) = ζb
(5.67)
αzp (+αz1i if p = 1) = αzb
(5.68)
549
S o u n d
5.2.2.2.5
I n s u l a t i o n
Transmission coefficients
The transmission coefficient for any incident wave on plate 1 that is transmitted to plate p at a
specific angle of incidence is calculated from the ratio of the wave intensities in the x-direction
using
Ixp (θi )
τ(θi ) =
(5.69)
Ix1i (θi )
Note that Eq. 5.69 is also used to calculate transmission coefficients from an incident wave to
a different wave type on the same plate, as well as to the same wave type on the same plate;
although the latter are referred to as reflection coefficients.
For a unit amplitude wave incident upon the junction, the x-direction intensities are (Cremer
et al., 1973)
3
IBx1i (θi ) = B1 ωkB1
cos θi
for an incident bending wave
(5.70)
3
for an incident quasi-longitudinal wave
(5.71)
3
for an incident transverse shear wave
(5.72)
ILx1i (θi ) = 0.5ρs1 ω kL1 cos θi
ITx1i (θi ) = 0.5ρs1 ω kT1 cos θi
For the transmitted waves on plate p, the x-direction intensities are (Cremer et al., 1973)
1
Re{Fyp (−iωηp )∗ + Mzp (iωαzp )∗ } for transmitted bending waves
2
1
ILTxp (θi ) = Re{Fxp (−iωξp )∗ + Fzp (−iωζp )∗ } for transmitted in-plane waves
2
IBxp (θi ) =
(5.73)
(5.74)
where ∗ denotes the complex conjugate.
The transmitted wave intensities can now be re-written in terms of the complex amplitudes
determined from the matrix solution as (Bosmans, 1998)
−ω
Im{C1Fyp − C1Mzp C1∗αzp }|TBNp |2
2
−ω
ILxp (θi ) =
Im{C1Fxp C1∗ξp + C1Fzp C1∗ζp }|TLp |2
2
−ω
ITxp (θi ) =
Im{C2Fxp C2∗ξp + C2Fzp C2∗ζp }|TTp |2
2
IBxp (θi ) =
(5.75)
(5.76)
(5.77)
The angular average transmission coefficients are calculated using Eq. 5.6. Transmission
coefficients usually fluctuate rapidly with incident angle; hence it is important to use a fine
angular resolution. This is in contrast to the wave approach for only bending waves where
the transmission coefficients tend to vary more slowly with incident angle. Each coupling loss
factor between the plates is calculated from Eq. 2.154.
5.2.2.2.6
Application to SEA models
To incorporate bending, quasi-longitudinal, and transverse shear waves into an SEA model;
each wave type requires its own subsystem. An example is shown in Fig. 5.7 using a T-junction.
Some transmission coefficients (and hence coupling loss factors) between different wave types
may be zero. However, it is necessary to include three subsystems for each plate because even
if wave conversion at the junction does not directly connect two subsystems together, there
can be transmission paths involving other subsystems that do connect them together. If there
is equipartition of modal energy between quasi-longitudinal and transverse shear modes it is
possible to use a single in-plane subsystem to represent both quasi-longitudinal and transverse
550
Chapter 5
Plate 3
Subsystem
T3
Subsystem
L3
Subsystem
B1
Subsystem
B3
Plate 1
Subsystem
B2
Subsystem
L1
Subsystem
T1
Subsystem
L2
Subsystem
T2
Plate 2
Figure 5.7
Nine-subsystem SEA model for three plates forming a T-junction. Subsystems for bending (B), quasi-longitudinal (L), and
transverse shear (T) wave motion are used for each plate. Double headed arrows are used to indicate all possible couplings
between subsystems; depending on the junction detail, some of the coupling loss factors will be zero.
shear waves (Craik and Thancanamootoo, 1992). As this is not usually known a priori and
requires calculation of the same transmission coefficients, this approach is not often used
unless it is beneficial in simplifying analysis of the results.
The wave approaches discussed in this section assume semi-infinite, thin plates. At this point
we want to use SEA subsystems to represent reverberant vibration fields associated with each
type of in-plane wave on finite plates. Wavelengths of in-plane waves are much longer than
those for bending waves (refer back to Fig. 2.2) and the fundamental quasi-longitudinal and
transverse shear modes tend to occur in the mid-frequency range. We can only consider use
of an SEA model where there are reverberant vibration fields; this means that inclusion of
subsystems for in-plane waves will only usually be appropriate in the mid- and high-frequency
ranges above the fundamental in-plane mode frequencies. Now it is useful to look at the thin
plate limit for bending waves on masonry/concrete walls in terms of the frequencies at which
the fundamental in-plane modes occur. Figure 5.8 shows these fundamental mode frequencies for typical wall dimensions and quasi-longitudinal phase velocities that correspond to
masonry/concrete. For 100 mm plates the thin plate limit is much higher than the fundamental
in-plane mode frequencies. For 200 mm plates we find that at frequencies where these plates
551
S o u n d
I n s u l a t i o n
Bending wave
thin plate limits
2000
h ⫽ 0.1 m, cL ⫽ 3500 m/s
Frequency (Hz)
1900
1800
1700
h ⫽ 0.1 m, cL ⫽ 3000 m/s
1600
1500
1400
Fundamental
in-plane modes
f11 (Hz)
1300
Quasi-longitudinal
cL ⫽ 3500 m/s
cL ⫽ 3000 m/s
cL ⫽ 2500 m/s
cL ⫽ 2000 m/s
1100
Transverse shear
cL ⫽ 3500 m/s
cL ⫽ 3000 m/s
cL ⫽ 2500 m/s
cL ⫽ 2000 m/s
h ⫽ 0.1 m, cL ⫽ 2500 m/s
1200
h ⫽ 0.1 m, cL ⫽ 2000 m/s
1000
h ⫽ 0.2 m, cL ⫽ 3500 m/s
900
800
h ⫽ 0.2 m, cL ⫽ 3000 m/s
700
h ⫽ 0.2 m, cL ⫽ 2500 m/s
600
h ⫽ 0.2 m, cL ⫽ 2000 m/s
500
400
300
200
2
2.5
3
3.5
4
4.5
5
Plate dimension, Lx (m)
Plate dimension, Ly ⫽ 2.4 m
Figure 5.8
Comparison of thin plate limits with fundamental in-plane mode frequencies for a plate with dimensions Lx and Ly .
can start to be considered as being multi-modal for bending waves, the confounding features of
non-diffuse in-plane wave fields need to be considered at the same time as thick plate bending
wave theory. As previously noted in practical calculations of airborne sound insulation (Section 4.3.1.4), the thin plate limit is not an absolute limit and different approaches can be taken.
The simplest approach is to continue to use thin plate bending wave theory along with the transmission coefficients derived for thin plates; comparisons with measurements often show that
this is perfectly adequate (e.g. see Kihlman, 1967). An alternative is to continue to use the thin
plate transmission coefficients but to use the group velocity for thick plates when calculating
coupling loss factors from the bending wave subsystem (Craik and Thancanamootoo, 1992).
A more complex alternative is to derive transmission coefficients that take account of shear
deformation and rotatory inertia (McCollum and Cuschieri, 1990). A purely theoretical consideration of incorporating in-plane waves into SEA models of masonry/concrete buildings tends
to be rather pessimistic of the outcome (e.g. see de Vries et al., 1981). However, comparisons
of measurements and predictions in masonry/concrete buildings indicate reasonable agreement for vibration transmission across a single junction as well as long distance transmission
across many junctions (Craik and Thancanamootoo, 1992; Kihlman, 1967; Roland, 1988).
5.2.2.3 Example: Comparison of wave approaches
It is now useful to compare the two wave approaches. We will look at a T-junction of masonry
walls that represents a solid separating wall (plate 1) rigidly connected to two solid flanking
walls (plates 2 and 3).
552
Chapter 5
0.20
Plate 3
Plate 2
0.18
Transmission coefficient (⫺)
0.16
0.14
Plate 1
0.12
0.10
B1B2
B2B1
0.08
B2B3
0.06
0.04
0.02
0.00
0
10
20
30
40
50
60
70
80
90
Angle of incidence, θ (°)
Figure 5.9
Angle-dependent transmission coefficients between three plates that form a T-junction (h1 = 0.215 m, h2 = h3 = 0.1 m,
ρ = 2000 kg/m3 , cL = 3200 m/s, ν = 0.2). These are calculated using the wave approach for bending waves only. The legend
refers to Bi Bj which corresponds to an incident bending wave on plate i that results in a transmitted bending wave on plate j.
The wave approach for only bending waves gives angle-dependent transmission coefficients
that are independent of frequency. These are shown in Fig. 5.9. For an incident wave on plate 2
that is transmitted to plate 1, kB2 > kB1 , and there is a cut-off angle, θco = 43◦ (Eq. 5.9) above
which no bending wave is transmitted and the transmission coefficient is zero. As plates 2
and 3 are the same, this cut-off angle would also occur for an incident wave on plate 3 that is
transmitted to plate 1.
The wave approach for bending and in-plane waves gives angle-dependent transmission coefficients that vary with frequency, and there are more transmission coefficients to consider. For
this reason it is simpler to present the results by grouping all the transmission and reflection
coefficients together. As an example; for an incident bending wave on plate 1, conservation of
energy ensures that
τB1B1 + τB1B2 + τB1B3 + τB1L1 + τB1L2 + τB1L3 + τB1T1 + τB1T2 + τB1T3 = 1
(5.78)
where B, L, and T indicate bending, quasi-longitudinal, and transverse shear waves respectively. Note that τB1B1 would normally be referred to as a reflection coefficient.
This grouping makes it possible to use a single graph to show the transmission and reflection
coefficients at a single frequency when one wave type, on one plate, is incident upon the
junction. Examples are shown in Fig. 5.10 for an incident bending wave. On these graphs
it is the vertical distance between the lines that gives the transmission coefficient. At 100 Hz
(Fig. 5.10 a and b), wave conversion from bending to in-plane waves mainly occurs below
10◦ , whereas above 10◦ the majority of the incident intensity is only reflected or transmitted
as bending waves. The trends are similar at 1000 Hz (Fig. 5.10 c and d) but there is wave
553
S o u n d
I n s u l a t i o n
Transmission and reflection coefficients (⫺)
(a) Incident bending wave on plate 1 at 100 Hz
Angle of incidence, θ (°)
0
10
20
30
40
1.0
B1T3
B1B3
0.9 B1L3
0.8 B1L2
50
60
70 8090
B1B2
0.7 B1B3
B1T2
0.6
B1B1
B1B2
0.5
0.4
0.3
0.2
0.1
0.0
0
0.1
0.2
0.3
0.4
0.5
sin θ
0.6
0.7
0.8
0.9
1
(b) Incident bending wave on plate 2 at 100 Hz
Angle of incidence, θ (°)
0
10
20
30
40
50
60
70 8090
Transmission and reflection coefficients (⫺)
1.0
B2T1
0.9 B2L1
B2B3
0.8
0.7
0.6
B2B2
B2B2
0.5
0.4
0.3
0.2
0.1
B2B1
B2B1
0.0
0
0.1
0.2
0.3
0.4
0.5
sin θ
0.6
0.7
0.8
0.9
1
Figure 5.10
Angle-dependent transmission coefficients between three plates that form a T-junction (h1 = 0.215 m, h2 = h3 = 0.1 m,
ρ = 2000 kg/m3 , cL = 3200 m/s, ν = 0.2). These are calculated using the wave approach for bending and in-plane waves.
The descriptors refer to BiBj / BiLj / BiTj which correspond to an incident bending wave (B) on plate i that results in a transmitted
bending wave (B) / quasi-longitudinal wave (L) / transverse shear wave (T) on plate j.
554
Chapter 5
(c) Incident bending wave on plate 1 at 1000 Hz
Angle of incidence, θ (°)
0
10
20
30
40
50
Transmission and reflection coefficients (⫺)
1.0
60
70 8090
B1B3
B1B2
0.9
B1T3
B1L3
0.8
B1B1
0.7
0.6
B1L2
B1T2
0.5
0.4
B1B3
0.3
B1B2
0.2
B1B1
0.1
0.0
0
0.1
0.2
0.3
0.4
0.5
sin θ
0.6
0.7
0.8
0.9
1
(d) Incident bending wave on plate 2 at 1000 Hz
Angle of incidence, θ (°)
0
10
20
30
40
50
60
70 8090
Transmission and reflection coefficients (⫺)
1.0
0.9
B2T1
B2B3
0.8
B2L1
0.7
0.6
B2B3
0.5
0.4
B2B2
B2B2
0.3
0.2
0.1
B2B1
B2B1
0.0
0
0.1
0.2
0.3
0.4
0.5
sin θ
0.6
0.7
0.8
0.9
1
Figure 5.10
(Continued)
555
S o u n d
I n s u l a t i o n
conversion over a slightly wider range of incident angles. At both 100 and 1000 Hz, we note
that τB2B3 has high values over a relatively narrow range of angles.
The features relating to these transmission coefficients can be considered in the context of
the equivalent angle of incidence for an individual mode (Section 2.5.3). This indicates that
with typical wall dimensions there is a limited range of angles at which waves impinge upon
the plate boundaries, particularly in the low- and mid-frequency ranges (refer back to bending
modes for a 215 mm wall in Fig. 2.30). Hence wave conversion from bending to in-plane waves
may be predicted to occur at angles of incidence that are not present on a finite size plate.
Fortunately, the uncertainty in describing wave motion on real walls and floors means that
consideration of precise angles of incidence is not usually appropriate or necessary. When
using these transmission coefficients with SEA, the assumption is made that there are diffuse
vibration fields. Therefore it is the angular average transmission coefficient that is needed.
This provides a reasonable estimate if there are a sufficient number of modes to approximately
cover the range of angles of incidence. This makes a link to a requirement often used in SEA
that there should be at least five modes in each frequency band (Ns ≥ 5) to give estimates
with low variance (Fahy and Mohammed, 1992). This requirement is not always satisfied for
masonry/concrete walls and floors in the low- and mid-frequency ranges; this is discussed
further in the examples in Section 5.2.3.
Angular average transmission coefficients are independent of frequency when the wave
approach assumes that there are only bending waves. This is a useful feature to note because
whilst this is the simpler and more convenient model, measurements on real walls and floors
usually exhibit some frequency-dependence due to the generation of in-plane waves. This
tends to be more apparent across the straight section of a T or X-junction rather than around
the corner of an L, T, or X-junction. This can be seen in Fig. 5.11 by comparing the angular average bending wave transmission coefficients for the T-junction using the two different
wave approaches; the coefficients are shown in terms of the transmission loss in decibels
using Eq. 5.2.
5.2.2.4
Other plate junctions modelled using a wave approach
This section contains a brief discussion of other idealized junctions that are relevant to buildings. Note that automotive and aerospace industries also use models of wave transmission
between connected beams and plates and some of these are applicable to lightweight building
structures.
5.2.2.4.1
Junctions of angled plates
The wave approach for bending and in-plane waves in Section 5.2.2 can also be used for rigid
junctions where the plates are orientated at angles other than right-angles to each other. It is
common for the separating wall to be perpendicular to the flanking walls to give box-shaped
rooms. However there are other room shapes in buildings such as those with hexagonal,
triangular, or octagonal floor plans. The difference between the angular average transmission
coefficients for perpendicular masonry/concrete walls in a T-junction when compared with a
junction of the same three walls with angles between 60◦ and 135◦ can be up to 10 dB or
more depending on the wall properties. Note that this does not usually result in similarly large
differences for the airborne sound insulation between adjacent rooms in terms of a singlenumber quantity such as Rw or DnT ,w .
556
Chapter 5
30
Transmission loss (dB)
25
B1B2 (Bending only)
B1B2 (Bending and in-plane)
B2B1 (Bending only)
B2B1 (Bending and in-plane)
B2B3 (Bending only)
B2B3 (Bending and in-plane)
20
15
10
5
0
50
80
125
200
315
500
800
1250
2000
3150
5000
One-third-octave-band centre frequency (Hz)
Figure 5.11
Angular average transmission coefficients (shown in decibels using the transmission loss, TLij ) for bending waves transmitted
across a T-junction (h1 = 0.215 m, h2 = h3 = 0.1 m, ρ = 2000 kg/m3 , cL = 3200 m/s, ν = 0.2). Transmission coefficients were
calculated using two wave approaches: one for bending waves only, and the other for bending and in-plane waves.
5.2.2.4.2 Resilient junctions
Examples of resilient plate junctions in buildings include a concrete floor slab resting on a
continuous resilient layer on top of the supporting walls, as well as walls, floors or floating floors
where a resilient material is used to form a vibration break or expansion joint. From Mees and
Vermeir (1993) the wave approach for bending and in-plane waves can be extended to include
an isotropic, homogeneous resilient layer at a plate junction. The resilient layer is treated as a
spring hence it must be sufficiently thin that it does not support wave motion over the frequency
range of interest.
Using the T-junction of masonry walls from Section 5.2.2.3, the transmission losses for the
rigid T-junction can be compared to two examples of resilient T-junctions in Fig. 5.12. These
indicate large changes in bending and in-plane wave transmission due to the resilient layers
at the junction. For the resilient junction in Fig. 5.12b, the bending wave transmission loss
corresponding to τB2B3 in the mid- and high-frequency ranges is 0 dB; hence there is potential
for significant flanking transmission via the two flanking walls. Both of the resilient junctions
(Fig. 5.12 a and b) indicate the need to include in-plane wave subsystems in an SEA model.
It is important to note that the insertion of resilient materials at junctions of walls and/or floors
is not a panacea for increasing the sound insulation between rooms. Due to conservation of
energy the use of resilient junctions results in a different distribution of energy between the
space and structural subsystems compared to rigid junctions; hence resilient junctions can
increase the sound insulation between some rooms in a building whilst decreasing it between
others (Osipov and Vermeir, 1996). For this reason it is useful to be able to incorporate resilient
junctions in a full SEA model of a building because calculating sound transmission between
two specific rooms along a limited number of flanking paths will sometimes be misleading.
557
(a) Rigid junction
80
75
1
70
3
2
65
60
Transmission loss (dB)
55
50
45
40
B1B2
B1L2
B1T2
B2B1
B2L1
B2T1
B2B3
B2L3
B2T3
35
30
25
20
15
10
5
0
50
80
125
200
315
500
800
1250
2000
3150
5000
One-third-octave-band centre frequency (Hz)
(b) Resilient junction (one resilient layer)
80
⫽10 mm thick resilient material (E ⫽ 3 MN/m2, ν ⫽ 0.4)
75
1
70
3
65
2
60
Transmission loss (dB)
55
50
45
40
35
30
25
20
15
B1B2
B1L2
B1T2
B2B1
B2L1
B2T1
B2B3
B2L3
B2T3
10
5
0
50
80
125
200
315
500
800
1250
2000
3150
5000
One-third-octave-band centre frequency (Hz)
Figure 5.12
Rigid and resilient T-junctions. Angular average transmission coefficients (shown in decibels using the transmission loss, TLij )
for an incident bending wave transmitted across a T-junction (h1 = 0.215 m, h2 = h3 = 0.1 m, ρ = 2000 kg/m3 , cL = 3200 m/s,
ν = 0.2). Values are calculated using a wave approach for both bending and in-plane waves. Note that B1L1, B1T1, B2L2,
and B2T2 are not shown on the graph.
558
Chapter 5
(c) Resilient junction (two resilient layers)
80
1
75
3
70
65
60
Transmission loss (dB)
55
2
= 10 mm thick
resilient material
(E ⫽ 3 MN/m2, ν ⫽ 0.4)
50
45
40
35
30
25
20
B1B2
B1L2
B1T2
B2B1
B2L1
B2T1
B2B3
B2L3
B2T3
15
10
5
0
50
80
125
200
315
500
800
1250
2000
3150
5000
One-third-octave-band centre frequency (Hz)
Figure 5.12
(Continued)
Comparisons of measurements and predictions on heavyweight walls and floors with resilient
junctions show reasonable agreement when considered alongside the uncertainty in describing
the properties of the resilient layer (Craik and Osipov, 1995; Pedersen, 1995). The Young’s
modulus of a resilient layer usually shows some dependence on both frequency and static load.
In addition, the transmission loss can be highly sensitive to the Poisson’s ratio of the layer (this
is typically between 0.1 and 0.5), but less sensitive to its internal damping (Mees and Vermeir,
1993).
5.2.2.4.3 Junctions at beams/columns
The junction beam used in the wave models in Section 5.2 is not a physical beam at the junction.
However, many buildings consist of a concrete or steel frame with walls built in-between the
columns, these plate junctions can be modelled by using a junction beam with mass and
stiffness that supports wave motion (Bosmans, 1998; Steel, 1994). Transmission between the
beams (columns) and the floor slab can be predicted using an impedance approach (Craik,
1996; Ljunggren, 1985; Lyon and Eichler, 1964).
5.2.2.4.4 Hinged junctions
A hinged plate junction has negligible transfer of bending moments due to an incomplete line
connection across the cross-section of one or more of the plates along the junction line. A model
for a hinged junction is given by Mees and Vermeir (1993). This tends to be more relevant to
plates which have not been fully connected over their cross-section due to poor workmanship.
559
S o u n d
I n s u l a t i o n
Hinged junctions are less common than rigid or resilient plate junctions for masonry/concrete
load-bearing walls in buildings.
5.2.3 Finite element method
The Finite Element Method (FEM) provides an example of how numerical models can be used
to determine parameters for SEA or SEA-based models (e.g. see Simmons, 1991; Steel and
Craik, 1994). The wave approaches in Sections 5.2.1 and 5.2.2 assume a diffuse vibration
field on each plate. Whilst this assumption is convenient, it is not always appropriate for vibration transmission between finite walls and floors over the building acoustics frequency range.
Numerical models allow modelling of finite plates and complex junction connections. However, they can be computationally intensive; hence wave approaches are often considered
to be adequate for practical purposes when the junction is relatively simple. FEM is primarily used in this section to perform numerical experiments that illustrate issues relevant
to finite plates with non-diffuse vibration fields and low modal overlap. These have implications for both the prediction and measurement of vibration transmission in situ and in the
laboratory.
A brief discussion is included here on relevant features of FEM models. For a thorough overview
the reader is referred to a text such as that by Zienkiewicz (1977) or Petyt (1998). The structure
under analysis is discretized into elements that are connected at nodal points to form a mesh
of elements; this mesh may be a simple rectangular grid for a homogeneous rectangular
plate. In the dynamic analysis, the unknowns to be determined are the degrees of freedom
(displacements and rotations) of all nodes in the mesh. The general equation of motion for
linear systems under steady-state excitation by sinusoidal point forces is
[M]
*
∂2 ŵ
∂t 2
+
+ [C]
*
∂ŵ
∂t
+
+ [K]{ŵ} = {F̂}
(5.79)
where w is the nodal displacement (complex), F is the applied force (complex), [M] is the mass
matrix, [C] is the damping matrix, and [K] is the stiffness matrix.
The displacement and force vectors are given by
ŵ = |ŵ| exp(iφ) exp(iωt)
(5.80)
F̂ = |F̂| exp(iϕ) exp(iωt)
(5.81)
where φ is the phase shift due to damping, and ϕ is the applied phase shift to the input force.
Due to the harmonic time-dependence of the displacement and the force, Eq. 5.79 simplifies to
(−ω2 [M] + iω[C] + [K]){ŵ} = {F̂}
(5.82)
There are many different types of finite elements. As FEM is mainly used at frequencies
below the thin plate limit for bending waves, a plate element is usually chosen to simulate thin
plate theory. The finite element dimensions required to achieve a certain degree of accuracy
depends upon the frequency. This can be determined by successively reducing the element
size until there is satisfactory convergence towards a solution; suitable element sizes are
usually between λB /3 and λB /6.
560
Chapter 5
Damping is often introduced into the model using the fraction of critical damping, the constant
damping ratio, ζcdr . The relationship between the loss factor and the constant damping ratio is
ζcdr =
η
2
(5.83)
Whilst it is important to assign damping to the plates in a realistic way, it is not always straightforward. As an example, consider a FEM model of three walls connected to form a T-junction.
In a building these walls will be connected to other walls and floors. It is therefore the total
loss factor that is needed to calculate the constant damping ratio for the model rather than the
internal loss factor. Whilst the total loss factor can be estimated (Section 2.6.5) it is not always
possible to assign different damping to each wall in a model; fortunately the total loss factors
are sometimes similar. The total loss factor for masonry/concrete walls and floors is usually
much higher than the internal loss factor so care also needs to be taken that the damping used
in the FEM model does not cause a significant decrease in vibration with distance across each
plate (Section 2.7.7).
Around the plate boundaries there will need to be constrained nodes to account for connections
to other plates that are not included in the FEM model; this is usually done by assuming simply
supported boundaries. In a similar way to the wave approach, the nodes along the junction
line can either be unconstrained to allow generation of both bending and in-plane waves, or
simply supported so that only bending wave motion is considered.
As with laboratory measurements of the velocity level difference, excitation can be applied at
single points. One advantage of using FEM is that rain-on-the-roof excitation can be applied to
try and satisfy the SEA requirement for statistically independent excitation forces. This requires
forces with unity magnitude and random phase to be applied at all the unconstrained nodes
over the surface of the source plate. The output from the FEM model can be used to calculate
the subsystem energy and power input that apply to an SEA model of the plate junction. These
values correspond to individual frequencies whilst for practical purposes we require frequency
bands; hence the energy or power at a number of frequencies can be averaged to give a
representative value for the frequency band. As with the laboratory measurement, it is bending
wave motion that is of interest on both source and receiving plates. FEM models usually
output nodal displacements in all three coordinate directions, but it is the lateral displacement
corresponding to bending wave motion that is used in subsequent calculations. At a sinusoidal
frequency, ω, the energy associated with each plate is
2
E = mv t,s = ω
2
&
N
1
mn |ŵn |2
2
n=1
'
(5.84)
where N is the number of unconstrained nodes on the plate, ŵ is the peak lateral displacement
and m is the mass associated with the node. (Note that in Chapter 2 we used η̂ rather than ŵ
for bending wave displacement but η is now needed for loss factors.)
With single point excitation, consideration needs to be given to excluding displacement from
nodes near the excitation point where the direct field dominates over the reverberant field
(Section 2.7.5); this becomes more important when high damping is used in the FEM model.
This makes it more convenient to use rain-on-the-roof excitation over all unconstrained nodes
for which the vibration of the entire plate represents the reverberant energy stored by the plate
subsystem.
561
S o u n d
I n s u l a t i o n
The velocity at the node where the power is injected is complex and is related to the
displacement by
v̂ = iωŵ
(5.85)
For single point excitation where the applied force only has a real component, the injected
power is
1
−ω
Win = Re{F̂ v̂ ∗ } =
Re{F̂}Im{ŵ}
(5.86)
2
2
For rain-on-the-roof excitation at P nodes the forces must be complex, so the injected power is
P
ω
(Im{F̂}Re{ŵ} − Re{F̂}Im{ŵ})p
(5.87)
Win =
2
p=1
We now look at ways in which the plate energies can be used to determine parameters of more
practical use. The simplest calculation is the velocity level difference between any two of the
plates that form the junction. This is calculated from the plate energies using
mi
Ei
Dv,ij = 10 lg
(5.88)
− 10 lg
Ej
mj
where mi and mj are the mass of plates i and j respectively.
This numerical experiment with FEM has effectively simulated the physical experiment
described in Section 3.12.3.3 to determine the coupling loss factor from the measured velocity level difference. The calculations and the assumptions are identical. It therefore follows
that the vibration reduction index can also be calculated from the velocity level difference
(Section 3.12.3.4). These calculations can be considered as basic forms of experimental
SEA (ESEA).
A more involved form of ESEA makes use of the SEA power balance equations to determine the
loss factors through inversion of an energy matrix (Lyon, 1975). This makes use of a general
ESEA matrix that can be determined from the general SEA matrix (Eq. 4.10),
⎡N
⎤
η
−η
−η
·
·
·
−η
21
31
N1 ⎥
⎢n=1 1n
⎢
⎥⎡
⎤
N
⎢
⎥ E11 E12 E13 · · · E1N
⎢ −η12
⎥
η
−η
2n
32
⎥
⎢
⎥ ⎢E E E
⎢
⎥
⎥ ⎢ 21 22 23
n=1
⎢
⎥⎢
⎥
N
⎢
⎥ ⎢ E31 E32 E33
⎥
η3n
⎢ −η13 −η23
⎥⎢ .
⎥
..
⎢
⎥⎢ .
⎥
n=1
⎢
⎥⎣ .
⎦
.
⎢ ..
⎥
..
⎢ .
⎥ EN1
.
E
NN
⎢
⎥
⎣
⎦
N
−η1N
ηNn
n=1
⎡W
in(1)
⎢ ω
⎢
⎢ 0
⎢
⎢
⎢
=⎢ 0
⎢
⎢ .
⎢ .
⎢ .
⎣
0
0
0 ···
Win(2)
0
ω
Win(3)
0
ω
..
.
0
Win(N)
ω
⎤
⎥
⎥
⎥
⎥
⎥
⎥
⎥
⎥
⎥
⎥
⎥
⎦
where Eij is the energy of subsystem i when the power is input into subsystem j.
562
(5.89)
Chapter 5
Numerical experiments with FEM are carried out to fill in all the terms in the energy and power
matrices. The energy matrix should be well conditioned if the system has been partitioned into
suitable subsystems and ηij << ηii (referred to as ESEA ‘weak’ coupling); so the energy terms
on the diagonal (E11 to ENN ) should be significantly larger than off-diagonal terms. However,
ill-conditioned matrices can still occur due to errors in the subsystem energies; this can result in
negative coupling loss factors. Different approaches can be taken if this problem occurs. One
possibility is to change the matrix layout used to calculate the coupling loss factors (see Lalor,
1990), an alternative approach is to use matrix-fitting routines to persuade/force the system to
fit an SEA model (Clarkson and Ranky, 1984; Hodges et al., 1987; Woodhouse, 1981). Failure
of the matrix inversion to produce positive coupling loss factors does not prove that the system
can or cannot be modelled using SEA. Assuming that it is only errors in the energies that cause
the negative values, the system may still be suited to SEA. Fortunately, building elements such
as masonry walls are quite highly damped when they are fully connected to other walls and
floors; this tends to reduce the occurrence of negative coupling loss factors.
5.2.3.1 Introducing uncertainty
Using a deterministic model such as FEM does not allow statistical considerations to be ignored.
With any physical construction there is uncertainty in the material properties and dimensions.
For finite plates with fractional or generally low mode counts this becomes overtly obvious
when predicting vibration parameters at individual frequencies and converting the result into
one-third-octave or octave-bands. Comparisons of measurements and predictions often show
that the peaks and troughs in the response do not always occur in the same frequency bands.
To take uncertainty into account it is possible to use a Monte Carlo technique to generate an
ensemble of similar constructions. Each of these constructions can then be modelled using a
deterministic approach so that individual results can be compared, or the ensemble of results
can be analysed to give statistical parameters. The technique is based on random number
generation to determine each variable based on a chosen statistical distribution. The exact
statistical distribution is rarely known, but for physical properties it is reasonable to assume
a normal (Gaussian) distribution, N(μ, σ). Note that this distribution is not bounded and the
physical variables (e.g. dimensions, bending stiffness) can only be positive, but by confining
its use to σ/μ < 0.3 it is possible to avoid any significant errors (Keane and Manohar, 1993).
5.2.3.2 Example: Comparison of FEM with measurements
We will shortly compare SEA using wave approaches with FEM. Before this it is useful to
compare FEM with measurements to confirm that it provides a reasonable model. This makes
use of a physical experiment on five rigidly connected masonry walls that form a free-standing
H-block on a large concrete floor (Hopkins, 2003a). The H-block is essentially two T-junctions
joined by a single wall that is common to both of them. Each rectangular wall is only connected
on two or three of its four sides, so the sum of its coupling loss factors is relatively low and the
total loss factor is similar to the internal loss factor. This means that the modal fluctuations in the
velocity level differences are easier to distinguish. It also simplifies matters because the same
damping values can be used for all walls in the FEM model; note that for fully connected walls
it is more complicated to account for different damping on each plate associated with bending
and in-plane motion. The Monte Carlo approach described in Section 5.2.3.1 is used to create
an ensemble of similar constructions based on the estimated uncertainty for all dimensions
and material properties. FEM is used to model each construction in the ensemble to determine
563
S o u n d
I n s u l a t i o n
the ensemble average velocity level difference and 95% confidence interval. Finite elements
for thin plate bending wave theory are used along with rain-on-the-roof excitation of bending
wave motion on the chosen source plate. Velocity level differences represent lateral motion to
each plate surface (i.e. bending wave motion).
Figure 5.13 compares measurements with two FEM predictions, one using a simply supported
junction line and the other using an unconstrained junction line. The large modal fluctuations
in the low-frequency range show that even when trying to account for the uncertainty, the
predicted peaks and troughs can still end up in the adjacent frequency band to which they are
measured. For transmission around the corner in the mid- and high-frequency ranges, the 95%
confidence intervals for the measurements and the predictions overlap each other and either
model for the junction line could be assumed to be appropriate. However, for transmission
across the straight section in the high-frequency range, only the unconstrained junction line is
30
Transmission across the straight section
Dv,32 (Source: Plate 3)
25
20
Velocity level difference (dB)
15
10
Plate 3
5
Plate 2
0
Measured
Plate 1
FEM: Simply supported junction line
FEM: Unconstrained junction line
15
10
5
0
Transmission around the corner
Dv,13 (Source: Plate 1)
⫺5
⫺10
50
80
125
200
315
500
800
1250
2000
3150
One-third-octave-band centre frequency (Hz)
Figure 5.13
Comparison of FEM with measurements of vibration transmission between masonry walls. The H-block of five rigidly connected
masonry walls is free-standing on a wide 300 mm thick concrete floor. Plate 1 is a 215 mm separating wall (Lx = 4.5 m,
Ly = 2.5 m) and the other four plates are 100 mm flanking walls (Lx = 3.6–4.1 m, Ly = 2.5 m). All walls are built from solid
dense aggregate blocks (ρ = 2000 kg/m3 ). The thin plate limit for bending waves, fB(thin) , is in the 800 Hz band for plate 1 and
the 2000 Hz band for plates 2 and 3. Predictions and measurements are both shown with 95% confidence intervals. Measured
data from Hopkins are reproduced with permission from ODPM and BRE.
564
Chapter 5
appropriate. At frequencies above the thin plate limits there is no indication that any significant
errors occur by using only thin plate theory.
5.2.3.3 Example: Comparison of FEM with SEA (wave approaches) for isolated
junctions
In Section 4.3.1 we looked at resonant coupling between the sound field in a room and bending wave motion on a plate. Even when the plates had relatively low mode counts and low
modal overlap, there was reasonable agreement between measurements and SEA because
the plates were coupled to rooms where the mode counts and modal overlap factors were
much higher. For resonant coupling between masonry/concrete plates it is common for both
subsystems to have low mode counts and low modal overlap. This section uses FEM and
SEA to discuss statistical and deterministic views of vibration transmission with excitation of
bending waves on one plate and comparison of the velocity level difference between plates
(bending wave motion).
Here it is useful to look at FEM models for an ensemble of similar junctions where the walls
have slightly different mode frequencies. This allows an insight into issues that occur with real
masonry/concrete structures where the mode counts are low and there is uncertainty in predicting individual mode frequencies. There is more than one variable that can be altered to
generate this ensemble; we will vary the Lx dimension that lies perpendicular to the junction
line. Variation in Lx is created with a Monte Carlo approach by using random numbers drawn
from a normal distribution, N(μ, σ). The mean value, μ = Lx will either be 3, 3.5 or 4 m and the
standard deviation is chosen to be σ = 0.25 m. This gives a low degree of variation (σ/μ < 0.1)
so that the SEA prediction using the mean value for Lx is not significantly different to any individual junction with a different Lx value. The level of uncertainty introduced into this ensemble
would not be unusual at the early design stage of a building. Even when the building plans
and dimensions are finalized there will still be uncertainty in the material parameters that can
change the mode frequencies. In these examples the ensemble is quite small; the ensemble
has 10 members for 50–1000 Hz, and only 5 members for 1250–3150 Hz.
Figure 5.14a shows the velocity level difference for an L-junction where only bending waves
are considered in the SEA and FEM models. The first point to note is the high level of variation
in the individual members of the FEM ensemble; hence a single deterministic model is likely
to be of little practical use even with relatively low levels of uncertainty. Individual members of
the FEM ensemble show large fluctuations when Ns < 5 and Mav < 1. However, the arithmetic
average of the FEM ensemble shows close agreement with the SEA model, particularly under
the conditions that Mav ≥ 1 and Ns ≥ 5 (Fahy and Mohammed, 1992).
Figure 5.14b shows the same L-junction modelled with both bending and in-plane waves. The
FEM ensemble average and SEA show similar frequency-dependence but the ensemble average does not tend towards the SEA solution quite as quickly as in Fig. 5.14a. The fundamental
in-plane modes occur just as each bending wave subsystem becomes multi-modal (Ns ≥ 5)
and the frequency at which the FEM ensemble average and SEA converge is higher than in
Fig. 5.14a.
For a T-junction we will only look at an unconstrained junction line because the comparison of FEM with measurements in Section 5.2.3.2 indicates that this is a more appropriate
model for rigidly connected masonry walls in the high-frequency range. The predicted
velocity level differences are shown in Fig. 5.15. As with the L-junction, there are large variations
565
(b) Bending and in-plane waves
(a) Only bending waves
Ns1 ⫽ 0.9 1.1 1.4 1.7 2.2 2.8 3.5 4.3 5.5 7.0 8.7 11 14 17 22 28 35 43 55
Ns2 ⫽ 0.9 1.1 1.4 1.8 2.2 2.8 3.5 4.4 5.5 7.0 8.8 11 14 18 22 28 35 44 55
Mav ⫽ 0.5 0.6 0.7 0.8 0.9 1.0 1.1 1.2 1.4 1.5 1.7 1.9 2.2 2.4 2.7 3.0 3.4 3.8 4.3
30
Ns1(B) ⫽ 0.9 1.1 1.4 1.7 2.2 2.8 3.5 4.3 5.5 7.0 8.7 11
0.7 1.1
Ns1(L) ⫽
0.7 1.1 1.8 2.8
Ns1(T) ⫽
M1(B) ⫽ 0.5 0.6 0.7 0.8 0.8 1.0 1.1 1.2 1.3 1.5 1.7 1.9
0.1 0.2
M1(L) ⫽
M1(T) ⫽
0.2 0.2 0.3 0.5
14
17
22
43
18
28
4.6 7.2 11
45
71
29
0.7 1.0 1.4 2.0 2.8 3.9 5.5
Plate 2
FEM: Individual members of the ensemble
FEM: Arithmetic average of the ensemble
FEM: Arithmetic average of the ensemble
Plate 1
25
18
55
2.1 2.4 2.7 3.0 3.4 3.8 4.2
SEA (bending and in-plane waves)
FEM: Individual members of the ensemble
28
0.3 0.4 0.6 0.8 1.1 1.6 2.2
30
Plate 2
SEA (bending waves only)
35
1.8 2.9 4.5 7.3 11
Plate 1
25
Dv,21 (Source: Plate 2)
15
10
Dv,12 (Source: Plate 1)
5
Velocity level difference (dB)
Velocity level difference (dB)
Dv,21 (Source: Plate 2)
20
20
15
Dv,12 (Source: Plate 1)
10
5
0
0
50
80
125
200
315
500
800
1250
One-third-octave-band centre frequency (Hz)
2000
3150
50
80
125
200
315
500
800
1250
2000
3150
One-third-octave-band centre frequency (Hz)
Figure 5.14
Predicted velocity level differences for an L-junction modelled with (a) only bending waves and (b) bending and in-plane waves. On (a) the upper x-axis labels show the predicted statistical mode
counts for each plate and the geometric modal overlap factor in each frequency band. On (b) the upper x-axis labels only show the predicted statistical mode counts and modal overlap factor in
each frequency band for plate 1. This reduces the amount of information because the values for plate 2 are similar and there are several permutations for the geometric modal overlap factor. Note
that values are only shown above the fundamental local mode. Plate properties: Lx1 = 4.0 m, Lx2 = 3.5 m, Ly1 = Ly2 = 2.4 m, h1 = h2 = 0.1 m, ρs1 = 140 kg/m2 , ρs2 = 60 kg/m2 , cL1 = 2200 m/s,
cL2 = 1900 m/s, ν = 0.2, loss factor η = f−0.5 (This was used in the SEA and FEM models so that the losses were indicative of walls connected on all sides as they would be in situ. The modal
overlap factors are calculated using this loss factor and should therefore be treated as minimum values because the total loss factor would also include the coupling loss factors from this junction.)
Chapter 5
between individual members of the FEM ensemble. Below the fundamental in-plane modes
where Mav < 1 and Ns < 5 for bending wave motion, the largest differences between ensemble average FEM values and SEA occur across the straight section of the T-junction. With
low modal overlap and low mode counts there is a tendency for the SEA wave approach to
overestimate the strength of the coupling loss factors. However, the differences between the
ensemble average FEM values and SEA are reasonably low and the latter can be used for most
practical purposes. Above the fundamental in-plane modes there are many permutations of Mav
between subsystems of different wave types and this simple descriptor becomes less useful.
5.2.3.4 Example: Statistical distributions of coupling parameters
In the previous example there were large differences in vibration transmission between junctions with similar walls when the mode counts and the modal overlap were low. Information
on the statistical distribution of coupling parameters is now needed to calculate the ensemble average and the standard deviation when such parameters are calculated from numerical
experiments or measured on a set of similar junctions. Any probability density function is likely
to be specific to certain systems, as well as specific to the choice of coupling parameter, the
method by which the ensemble is created, type of excitation, type of subsystems, and the
prediction or measurement method. Numerical experiments used to study various different
dynamical systems indicate that probability density functions are not always described by a
normal distribution; it has been found that coupled plates or beams often have right-skewed
distributions if modal overlap factors are less than unity (Fahy and Mohammed, 1992; Hodges
and Woodhouse, 1989; Manohar and Keane, 1994; Wester and Mace, 1999).
We now look at statistical distributions for an ensemble of junctions by taking the approach
of Fahy and Mohammed (1992). An ensemble is created using a Monte Carlo approach with
random numbers drawn from a normal distribution. An ensemble of 30 similar junctions of
masonry walls is created in exactly the same way as in Section 5.2.3.3 by varying the Lx
dimension of each plate and using a FEM model for each junction. We will focus on the lowand mid-frequency ranges where modal overlap is lowest and it is reasonable to model only
bending wave motion by using a simply supported junction line with both L and T-junctions.
Coupling loss factors are determined using ESEA as described in Section 5.2.3. These are
then used to produce normal quantile plots. This type of plot indicates normal distributions
when the majority of values in the ensemble lie along a straight line; any outliers will occur far
from this line. If there are two distinctly different gradients for the ensemble, this indicates a
skewed distribution. Figure 5.16 shows examples for an ensemble of random numbers drawn
from a normal distribution; the lines are approximately straight but there are some fluctuations.
These are intended for visual comparison with the normal quantile plots for the coupling loss
factors shown in Fig. 5.17 for an L and a T-junction (Hopkins, 2002). When the modal overlap
factor is less than unity for the coupled plates, the normal quantile plots for the linear coupling
loss factors have distinctly non-normal, right-skewed distributions, and tend to have the widest
spread of values. However, when plotted in decibels they tend towards straight lines; hence a
practical simplification is to describe the linear coupling loss factor by a log-normal distribution.
The mean, standard deviation and confidence intervals can then be calculated from the values
in decibels. Upper and lower confidence intervals can subsequently be used in path analysis
with SEA or SEA-based models to give an indication of the range of possible values.
The SEA matrix solution requires linear coupling loss factors, so a reverse transformation is
needed to convert the mean and standard deviation from logarithmic to linear values. The
567
S o u n d
I n s u l a t i o n
Ns1(B) ⫽
Ns1(L) ⫽
0.4 0.6 0.7 0.9 1.1 1.4 1.8 2.2 2.8 3.5 4.4 5.6 6.9 8.9 11
18
14
0.9 1.4 2.1 3.5 5.4 8.5 13
0.8 1.3 2.2 3.4 5.3 8.7 14
Ns1(T) ⫽
Ns2(B) ⫽ 0.5 0.7 0.8 1.0 1.3 1.7 2.1 2.6 3.3 4.2 5.2 6.6 8.4 10 13 17 21
M1(B) ⫽
M1(L) ⫽
21
34
26
33
0.2 0.2 0.3 0.3 0.3 0.4 0.4 0.5 0.5 0.6 0.7 0.8 0.9 1.0 1.1 1.2 1.4
0.1 0.2 0.3 0.4 0.5 0.7 1.0
M1(T) ⫽
0.2 0.2 0.3 0.5 0.7 0.9 1.3 1.8 2.6
M2(B) ⫽ 0.3 0.4 0.4 0.5 0.5 0.6 0.6 0.7 0.8 0.9 1.0 1.1 1.3 1.4 1.6 1.8 2.0 2.3 2.6
45
SEA (bending and in-plane waves)
FEM: Individual members of the ensemble
40
FEM: Arithmetic average of the ensemble
Plate 3
Plate 2
35
30
Plate 1
25
Dv,23 (Source: Plate 2)
20
Velocity level difference (dB)
15
10
35
30
Dv,21 (Source: Plate 2)
25
20
15
20
Dv,12 (Source: Plate 1)
15
10
5
0
50
80
125
200
315
500
800
1250
One-third-octave-band centre frequency (Hz)
568
2000
3150
Chapter 5
3.0
2.5
2.0
Normal distribution, z
1.5
1.0
0.5
0.0
⫺0.5
⫺1.0
⫺1.5
⫺2.0
⫺2.5
⫺3.0
0
2
4
6
8
10
12
Arbitrary values
14
16
18
20
Figure 5.16
Normal quantile plot for 15 sets of 30 random numbers drawn from a normal distribution.
normal distribution N(μ, σ) has two parameters, μ and σ, and is referred to as a two-parameter
distribution. The log-normal distribution is generally a three-parameter distribution (τ, μ , σ )
where τ is the lowest possible value or threshold of the distribution. Assuming that the linear
coupling loss factor can potentially be zero or infinitesimally small, then τ = 0, so the twoparameter log-normal distribution (μ , σ ) can be used. The log-normal distribution of linear
ηij is now denoted by (μ , σ ), and the normal distribution of lg(ηij ) by N(μ, σ). This gives the
geometric mean, μ = 10μ and the geometric standard deviation, σ = 10σ . The SEA matrix
solution can now be used with μ representing the ensemble average coupling loss factor.
Similar issues with non-normal distributions occur with the vibration reduction index because
it is directly related to the coupling loss factor (see Section 5.4.1.1). When the modal overlap
factor is greater than unity the spread of the coupling loss factors is much narrower and the
normal distribution can be used to describe either the linear values or those in decibels.
Non-normal distributions with modal overlap factors less than unity were found by Fahy and
Mohammed (1992) in numerical experiments using thin steel coupled plates. To date, this
<
Figure 5.15
Predicted velocity level differences for a T-junction modelled with bending and in-plane waves. The upper x-axis labels show
the predicted statistical mode counts and modal overlap factor in each frequency band for plate 1 (all three wave types) and
for plate 2 (bending waves). Note that these are only shown above the fundamental local mode. Values for plate 3 (bending
waves) are similar to plate 2 (bending waves) and values for plates 2 and 3 (in-plane waves) are similar to plate 1 (in-plane
waves). Plate properties: Lx1 = 4.0 m, Lx2 = 3.5 m, Lx3 = 3.0 m, Ly1 = Ly2 = Ly3 = 2.4 m, h1 = 0.215 m, h2 = h3 = 0.1 m,
ρs1 = 430 kg/m2 , ρs2 = ρs3 = 200 kg/m2 , cL = 3200 m/s, ν = 0.2, loss factor η = f−0.5 (This was used in the SEA and FEM
models so that the losses were indicative of walls connected on all sides as they would be in situ. The modal overlap factors
are calculated using this loss factor and should therefore be treated as minimum values because the total loss factor would
also include the coupling loss factors from this junction.)
569
Normal distribution, z
S o u n d
I n s u l a t i o n
3.0
3.0
2.5
2.5
2.0
2.0
1.5
1.5
1.0
1.0
0.5
0.5
0.0
0.0
⫺0.5
⫺0.5
Plate 2
⫺1.0
⫺1.0
⫺1.5
⫺1.5
⫺2.0
⫺2.0
Plate 1
⫺2.5
⫺2.5
⫺3.0
⫺3.0
0
20
40
60
80
100
120
92 94 96 98 100 102 104 106 108 110 112
Normal distribution, z
η12 (linear ⫻ 10⫺3)
η12 (dB)
3.0
3.0
2.5
2.5
2.0
2.0
1.5
1.5
1.0
1.0
0.5
0.5
0.0
0.0
⫺0.5
⫺0.5
Plate 2
⫺1.0
⫺1.0
⫺1.5
⫺1.5
⫺2.0
⫺2.0
Plate 1
⫺2.5
⫺2.5
⫺3.0
⫺3.0
0
1
2
3
4
η21 (linear ⫻ 10⫺3)
5
6
70 72 74 76 78 80 82 84 86 88 90 92 94 96 98
η21 (dB)
Figure 5.17
Normal quantile plots for the coupling loss factor (linear values and decibels) calculated using FEM (bending waves only)
and ESEA. Fourteen curves are shown, each corresponding to a one-third-octave-band between 50 and 1000 Hz. Each
curve is formed from an ensemble of 30 members. Curves denoted by –x– indicate that the modal overlap factor in the
frequency band is less than unity for plate i and/or plate j for the relevant ηij . L-junction plate properties: Lx1 = 4.0 m,
Lx2 = 3.5 m, Ly1 = Ly2 = 2.4 m, h1 = h2 = 0.1 m, ρs1 = 140 kg/m2 , ρs2 = 60 kg/m2 , cL1 = 2200 m/s, cL2 = 1900 m/s, ν = 0.2,
loss factor η = f−0.5 . T-junction plate properties: Lx1 = 4.0 m, Lx2 = 3.5 m, Lx3 = 3.0 m, Ly1 = Ly2 = Ly3 = 2.4 m, h1 = 0.215 m,
h2 = h3 = 0.1 m, ρs1 = 430 kg/m2 , ρs2 = ρs3 = 60 kg/m2 , cL1 = 3200 m/s, cL2 = cL3 = 1900 m/s, ν = 0.2, loss factor η = f−0.5 .
feature has therefore been identified with lightly damped plates that have relatively high modal
densities, as well as with highly damped plates with relatively low modal densities. Note that
these findings apply to ensembles created using parameters drawn from a normal distribution. It is not possible to draw wide-ranging conclusions relating to all numerical or physical
experiments. However, it emphasizes that more cautious conclusions should be drawn when
using results from a single deterministic model or a single physical experiment when the modal
570
Normal distribution, z
Chapter 5
3.0
3.0
2.5
2.5
2.0
2.0
1.5
1.5
1.0
1.0
0.5
0.5
0.0
0.0
⫺0.5
⫺0.5
Plate 2
⫺1.0
Plate 3
⫺1.0
⫺1.5
⫺1.5
⫺2.0
⫺2.0
⫺2.5
⫺2.5
⫺3.0
⫺3.0
0
0.05
0.1
0.15
0.2
η23 (linear ⫻ 10⫺3)
0.25
0.3
42 46 50 54 58 62 66 70 74 78 82 86
η23 (dB)
Figure 5.17
(Continued)
overlap factor of any plate in the junction is less than unity. This has implications for laboratory measurements used to determine coupling loss factors or the vibration reduction index
with junctions of masonry/concrete walls (Hopkins, 1996). There are potential advantages in
building free-standing wall constructions so that no laboratory facility is needed and the number
of flanking transmission paths can be minimized. However, when these plates are unconnected
on some sides they will have lower total loss factors, and hence lower modal overlap factors
than in situ. Therefore the resulting coupling parameters measured in the laboratory are likely
to be different in situ. Similar issues with low modal overlap occur when the walls and floors
are connected by a resilient material at the junction.
Most masonry/concrete walls and floors are rigidly connected and it is more convenient to
use the wave approach (bending waves only) to predict the coupling loss factor than to carry
out numerical or physical experiments. SEA predictions using the wave approach tend to
give reasonable estimates with low variance when the plates that form the junction satisfy the
empirical conditions, Mav ≥ 1 and Ns ≥ 5 (Fahy and Mohammed, 1992) where Mav is defined in
Eq. 4.7. For masonry/concrete plates these conditions are rarely satisfied in the low-frequency
range, and sometimes in the mid-frequency range too. When they are not satisfied the wave
approach tends to overestimate the coupling loss factor; this can be inferred from the examples
in Section 5.2.3.3 where the velocity level difference from SEA is generally lower than from
the FEM model. The modal overlap factor provides a simple, practical variable with which to
define these conditions, but it is not necessarily the key variable. Some numerical experiments
show it to be an inadequate indicator for rectangular plates (Wester and Mace, 1996). Using the
same ensemble for the L-junction as in the previous numerical experiments, Fig. 5.18 indicates
that there is no clear cut-off point and that for practical purposes the empirical conditions for
masonry/concrete walls could be relaxed. By allowing for errors in an SEA prediction that are
similar to those encountered from variation due to workmanship, Craik et al. (1991) propose
more lenient conditions when using the wave approach (bending waves only) to calculate
coupling loss factors in SEA; for a 5 dB error limit these are M > 0.25 and Ns > 0.3 for each plate.
571
S o u n d
I n s u l a t i o n
Difference between predicted coupling loss factors
Wave approach – Deterministic approach (dB)
10
N⬍5
Nⱖ5
5
0
⫺5
⫺10
0.0
0.5
1.0
1.5
2.0
Modal overlap factor (Geometric mean), Mav
2.5
3.0
Figure 5.18
Comparison of coupling loss factors determined from the wave approach with those from the deterministic approach using
FEM and ESEA. The same L-junction is used as described in Fig. 5.17. Each point corresponds to η12 or η21 for one of the
30 members of the ensemble and for one of the one-third-octave-bands between 50 and 1000 Hz.
5.2.3.5 Example: Walls with openings (e.g. windows, doors)
Flanking walls that are connected to a separating wall or floor usually contain openings such
as windows and doors. If an opening is introduced into a wall close to a junction it can alter
vibration transmission between the plates. The wave approach is well-suited to the calculation
of transmission coefficients between plates without openings where a wave is incident upon
the junction at a certain angle of incidence. It is less suited to plates with an opening close to
the junction because simple assumptions about angles of incidence and transmission are no
longer valid and any constraints on the displacement of the junction beam will vary along its
length. It is therefore easier to take a modal view of vibration transmission.
Two aspects require consideration when modelling walls with an opening: the first is whether
the wall can still be modelled as a single plate, and the second is how the vibration transmission
is changed by the presence of an opening. Modelling of doors and windows needs to be considered on a case-by-case basis. Here we will focus on masonry/concrete walls with windows
because this type of external flanking wall often forms part of important flanking paths.
The local mode shapes of a plate with an opening depend upon the boundary conditions of the
opening. At this point we can refer back to Fig. 2.33 where a window opening was modelled with
simply supported or free boundaries. For windows and doors, most permutations of these two
boundary conditions are possible. For windows in external cavity masonry walls, the boundaries
are not usually connected in a way that would justify the simply supported model; partly to avoid
thermal bridging and ingress of moisture. An exception for doors and windows is a lintel that
rigidly connects the inner to the outer leaf. When most or all boundaries of an opening are free
(i.e. unconstrained), it is usually possible to model the wall as a single plate. If most or all of
its boundaries are simply supported then the wall may act as one or more smaller plates that
572
Chapter 5
LAp
Lij
LAp
Lij
Figure 5.19
Window openings in flanking walls.
are separated by the opening. However, this will be limited to low frequencies where strips of
wall between the wall boundaries and the boundaries of the opening cannot support vibration.
Over the building acoustics frequency range, many masonry/concrete walls with openings can
simply be modelled as a single plate with a reduced surface area due to the opening. Note that
below the critical frequency, the boundary conditions of the opening can affect the radiation
efficiency of the wall. As a rule-of-thumb, the frequency-average radiation efficiency still gives
a reasonable estimate without any modification if the ratio of wall area to opening area is at
least two and the boundaries of the opening are not simply supported.
We now consider the effect of an opening on the power flow between connected walls; this was
illustrated with structural intensity measurements in Section 3.12.3.1.4. For a diffuse incidence
vibration field on a junction, an opening close to the junction can be seen as reducing the length
of the junction line (see Fig. 5.19); hence a simple estimate for the reduction in the coupling
loss factor (in decibels) due to an opening is given by
10 lg
Lij − LAp
Lij
(5.90)
where LAp is the opening dimension that lies closest to and parallel to the junction line, Lij .
(Note that when using the vibration reduction index this estimate would be an increase instead
of a reduction.)
As we have already established that the vibration field is not usually diffuse, we will use the
results from the L-junction in Section 5.2.3.3 to assess the estimate in Eq. 5.90. Additional
calculations with FEM and Monte Carlo methods are used to model an ensemble of similar
573
S o u n d
I n s u l a t i o n
8
LJ ⫽ 0.3 m
LJ ⫽ 0.4 m
6
η12
LJ ⫽ 0.2 m
LJ ⫽ 0.3 m
4
LJ ⫽ 0.4 m
3
LJ
LAp⫽1.1 m
LJ ⫽ 0.5 m
5
Change in the coupling loss factor
due to a window opening (dB)
Plate 2 with opening
LJ ⫽ 0.2 m
7
0.8 m
η21
Lij = 2.4 m
0.5 m
Plate 2
LJ ⫽ 0.5 m
Estimate
2
Plate 1
1
0
⫺1
⫺2
⫺3
⫺4
⫺5
⫺6
⫺7
⫺8
50
80
125
200
315
500
800
1250
2000
3150
One-third-octave-band centre frequency (Hz)
Figure 5.20
Predicted change in the coupling loss factor due to window openings in an L-junction. Plate properties: Lx1 = 4.0 m, Lx2 = 3.5 m,
Ly1 = Ly2 = 2.4 m, h1 = h2 = 0.1 m, ρs1 = 140 kg/m2 , ρs2 = 60 kg/m2 , cL1 = 2200 m/s, cL2 = 1900 m/s, ν = 0.2, loss factor
η = f−0.5 . Opening: Top boundary is simply supported. All other boundaries are free.
L-junctions with a small window opening at different positions in the flanking wall. The effect of
the opening is given by the difference between the ensemble average coupling loss factor with
and without the opening. It is reasonable to assume that the effect of openings on vibration
transmission will be most apparent when they are very close to the junction. Therefore the
distance, LJ , between the edge of the opening and the junction line is varied from 0.2 to 0.5 m.
Comparisons with laboratory measurements indicate that for an L-junction without an opening,
a simply supported junction line is appropriate in the FEM model (Hopkins, 2003). However,
when there is an opening close to the junction, there is better agreement with measurements
when the junction line is unconstrained to allow generation of bending and in-plane waves. The
L-junction used in this numerical example is kept within bounds that are similar to the construction validated with measurements (Hopkins, 2003b). For each member of the ensemble, ESEA
is used to determine coupling loss factors by assuming that the L-junction can be represented
by only two bending wave subsystems (i.e. one subsystem for each plate).
The predicted change in the coupling loss factors is shown in Fig. 5.20. These show large
fluctuations about the estimated value in the low- and mid-frequency ranges. Measurements
on similar walls with windows also show these large fluctuations; these are shown in Fig. 5.21
574
Chapter 5
8
7
6
5
4
3
Change in the coupling loss factor
due to a window opening (dB)
η12
Plate 2
Plate 1
2
1
η21
Estimate
0
⫺1
⫺2
⫺3
⫺4
⫺5
⫺6
⫺7
⫺8
⫺9
⫺10
LJ
LAp ≈ 1.2 m
Plate 2
Lij ⫽ 2.4 m
with opening
0.8–1 m 0.5–1.2 m
LJ ⫽ 0.25–0.5 m
≈4m
⫺11
⫺12
⫺13
⫺14
50
80
125
200
315
500
800
1250
One-third-octave-band centre frequency (Hz)
2000
3150
Figure 5.21
Measured change in the coupling loss factor due to window openings from five different L-junctions. Plate properties are
similar to those in Fig. 5.20. Measured data from Hopkins are reproduced with permission from ODPM and BRE.
where the window openings were cut out of each flanking wall after they were built; this allowed
measurements to initially be taken on the walls without an opening. In the mid- and highfrequency ranges where LJ > λB /4 the measurements indicate a larger reduction than given
by the simple estimate or the FEM model. This is partly due to rain-on-the-roof excitation
over the entire wall in FEM compared to measurements with point excitation at distances
well away from the junction. Numerical and physical experiments on T-junctions also show
large fluctuations from which it is difficult to assess whether the simple estimate is always
appropriate.
As it is impractical and not particularly informative to calculate coupling loss factors using FEM
for every combination of window and junction we need to consider how the simple estimate can
be used with the wave approach. It is fortunate that there are many transmission paths between
two rooms in a building. Therefore as LAp is typically half of Lij , the additional uncertainty of
±3 dB in this coupling loss factor can be tolerated as it will often have a small effect on the
overall sound insulation. To assess the effect of the opening with an SEA model the wave
approach can be used to calculate coupling loss factors with and without the simple estimate,
or the openings can be completely ignored if they are small. Any decision to exclude a small
wall from the model (because it contains an opening) should primarily be based around its
575
S o u n d
I n s u l a t i o n
inability to support bending modes. Care is needed to ensure that this will not exclude important
flanking paths that cross more than one junction; the effect of excluding flanking walls or floors
from an SEA model is shown with an example in Section 5.3.2.1. For most masonry/concrete
constructions, the predicted change in the single-number quantity due to an opening will be a
small increase in the sound insulation up to a few decibels. This small but beneficial aspect of
windows in flanking walls has been identified in statistical analysis of many field airborne sound
insulation measurements in flats (Sewell and Savage, 1987). If the windows are openable,
consideration also needs to be given to the flanking path involving airborne sound transmitted
between open windows in flanking walls on either side of the separating wall (Fothergill and
Hargreaves, 1992; Kawai et al., 2004).
5.2.3.6 Example: Using FEM, ESEA, and SEA with combinations of junctions
Numerical methods have been used in the previous examples to model isolated junctions consisting of plates with low mode counts and low modal overlap. For these isolated junctions,
the combination of FEM, ESEA, and the Monte Carlo technique can be used to give ensemble
average coupling loss factors. These are now incorporated in an SEA model of a larger construction formed from several different junctions. For this example we will look at seven coupled
plates formed by connecting T-junctions and L-junctions as shown in Fig. 5.22. In this example
we are getting closer to a situation of more practical interest, namely two adjacent rooms. The
focus here is on the low-frequency range where Mav < 1 and Ns < 5 and only bending waves
need consideration. Two SEA models are used for the construction; one using coupling loss
factors from the wave approach and the other using ensemble average coupling loss factors
from isolated junctions. The velocity level difference from the SEA models is compared with an
ensemble of 30 similar constructions modelled with FEM. This ensemble is created by varying
the Lx dimension of the plates in the same way as for the isolated junctions (Section 5.2.3.3);
note that the Lx values for the three plates of one T-junction are applied to the other plates so
that the perpendicular plates remain perpendicular to each other in each construction.
Predicted Dv,ij values are shown in Fig. 5.23 (Hopkins, 2002). Figure 5.23a shows Dv,ij between
plates that share the same junction line; these are relevant to the measurement of coupling
parameters (i.e. ηij or Kij ) as well as to SEA-based models that only consider vibration transmission across one plate junction (see Section 5.4). Figure 5.23b shows Dv,ij between plates
that do not share the same junction line. The values are generally higher but act as a reminder
that when predicting airborne or impact sound insulation between two adjacent rooms we will
need to consider the role of all the walls and floors that form each room.
From Fig. 5.23 the SEA model using the wave approach tends to give a lower Dv,ij than individual members of the ensemble; this was previously seen with the isolated junctions in Section
5.2.3.3. There is no clear advantage in using the ensemble average coupling loss factors; they
2
4
5
3
1
7
6
Figure 5.22
Seven coupled plates. Plate properties: Lx1 = Lx4 = Lx7 = 4.0 m, Lx2 = Lx5 = 3.5 m, Lx3 = Lx6 = 3.0 m, Ly = 2.4 m,
h1 = h7 = 0.215 m, h2 = h3 = h4 = h5 = h6 = 0.1 m, ρs1 = ρs7 = 430 kg/m2 , ρs4 = 140 kg/m2 , ρs2 = ρs3 = ρs5 = ρs6 =
60 kg/m2 , cL1 = cL7 = 3200 m/s, cL4 = 2200 m/s, cL2 = cL3 = cL5 = cL6 = 1900 m/s, ν = 0.2, loss factor η = f−0.5 .
576
Chapter 5
(a) Source and receiving plates connected together at a junction
SEA (using a wave approach: bending waves only)
SEA (ensemble average coupling loss factors from FEM and ESEA)
FEM: Individual members of the ensemble
55
Dv,23 (Source: Plate 2)
50
45
40
35
Velocity level difference (dB)
30
50
Dv,21 (Source: Plate 2)
45
40
35
30
25
20
Dv,13 (Source: Plate 1)
15
10
5
0
20
Dv,12 (Source: Plate 1)
15
10
5
0
50
63
80
100
125
160
One-third-octave-band centre frequency (Hz)
200
Figure 5.23
Predicted velocity level differences for the seven coupled plates.
can give a worse, similar, or better estimate than the wave approach. Note that these plates
have low mode counts (local modes) and a numerical method has been used to model vibration transmission based on global modes of the isolated junctions. The results have then been
used to derive coupling loss factors assuming that the transmission can be modelled on a local
mode basis. The isolated junctions have then been combined to make a larger construction of
seven coupled plates which has different global modes. The conclusion for masonry/concrete
plates with low mode counts and low modal overlap is that SEA models incorporating coupling
loss factors from either a single deterministic model or a single laboratory measurement on one
junction, need to be treated with some caution. If the statistics of the ensemble are of primary
importance, these can be calculated using numerical methods and Monte Carlo techniques.
However, the time needed to create and compute the numerical models tends to be prohibitive.
A simpler and more pragmatic approach to predictions for buildings is to use SEA with the
wave approach and then make an allowance for uncertainty when interpreting the results and
applying them to a single construction. For laboratory measurement of coupling parameters,
577
S o u n d
I n s u l a t i o n
(b) Source and receiving plates that are not directly connected together at a junction
40
35
30
Dv,25 (Source: Plate 2)
25
20
15
10
Dv,27
(Source: Plate 2)
Velocity level difference (dB)
80
75
70
65
60
55
50
45
Dv,17 (Source: Plate 1)
40
35
30
25
30
25
20
15
Dv,14 (Source: Plate 1)
10
5
50
63
80
100
125
160
One-third-octave-band centre frequency (Hz)
200
Figure 5.23
(Continued)
it is costly to build and measure an ensemble of similar junctions. However it is important to
note that field measurements confirm the high levels of variation in the low-frequency range
for junctions of masonry/concrete walls and floors (Craik and Evans, 1989).
5.2.4 Foundation coupling (Wave approach: bending waves only)
An important transmission path for masonry/concrete cavity walls occurs via the foundations
(Sections 4.3.5.2.2 and 4.3.5.4.3). The following model from Wilson and Craik (1995) considers
bending wave transmission between two plates on a strip foundation and highlights the role
of soil stiffness in vibration transmission. In this model, bending waves on the source wall are
assumed to cause rotation and displacement of the strip foundation; this is then resisted by the
soil stiffness and the inertia of the foundation. As motion of the foundation causes both shearing
and compression of the soil, the two relevant soil parameters are the shear stiffness per unit
′
area, Gsoil , and the compression stiffness per unit area, ssoil
. It is necessary to incorporate
578
Chapter 5
Plate i
Plate j
d1
d2
Figure 5.24
Cavity wall on a strip foundation.
′
damping in the compression stiffness using the loss factor, hence ssoil
is used in the form,
′
ssoil
(1 + iηsoil ). The shear stiffness per unit area can then be calculated from
Gsoil =
′
ssoil
(1 + iηsoil )
2(1 + νsoil )
(5.91)
where νsoil is Poisson’s ratio for soils (if unknown this can usually be assumed to be 0.25).
Two plates, i and j, are connected to a strip foundation as shown in Fig. 5.24. It is assumed that
the strip foundation has a rectangular cross-section, a mass per unit length, ρl , and does not
support bending or twisting motion. The incident bending wave on plate i has unit amplitude
and is incident upon the strip foundation at an angle of incidence, θi , from a line on the plate
that is perpendicular to the junction. Interaction of the incident wave with the foundation gives
rise to a reflected bending wave on plate i with an amplitude, RB , and a nearfield on plate i with
amplitude, Rn . The angle of transmission is found from Snell’s law (Section 5.2.1.2). When
(kB,i /kB, j ) sin θi ≤ 1, a bending wave is transmitted on plate j with an amplitude, TB , at an angle,
θj , as well as a nearfield with amplitude Tn . Otherwise there is no transmitted bending wave,
only a nearfield on plate j.
The following four equations can be solved in matrix form to give R, Rn , T , and Tn , for a specific
angle of incidence, although it is only T that is needed to calculate the transmission coefficient
(Wilson and Craik, 1995)
−R − Rn + T + Tn = 1
ikB,i cos θi R + kn,i Rn − ikB,j cos θj T − kn,j Tn = ikB,i cos θi
2
[Bp,i kB,i
( cos2 θi + νi sin2 θi ) − ik B,i cos θi (Kφ − ω2 Ir )]R
2
2
sin2 θi )]Rn
− νi kB,i
− [kn,i (Kφ − ω2 Ir ) + Bp,i (kn,i
2
2
2
( cos2 θj + νj sin2 θj )T − Bp,j (kn,j
+ Bp,j kB,j
− νj kB,j
sin2 θj )Tn
2
( cos2 θi + νi sin2 θi )
= −ik B,i cos θi (Kφ − ω2 Ir ) − Bp,i kB,i
3
[iBp,i kB,i
2
2
(5.92)
2
cos θi ( cos θi + (2 − νi ) sin θi ) − (Kf − ω ρl )]R
2
2
sin2 θi ) + (Kf − ω2 ρl )]Rn
− (2 + νi )kB,i
− [Bp,i kn,i (kn,i
3
+ iBp,j kB,j
cos θj ( cos2 θj + (2 − νj ) sin2 θj )T
2
2
− Bp,j kn,j (kn,j
− (2 + νj )kB,j
sin2 θj )Tn
3
= Kf − ω2 ρl + iBp,i kB,i
cos θi ( cos2 θi + (2 − νi ) sin2 θi )
579
S o u n d
I n s u l a t i o n
where the nearfield wavenumber on each plate, kn , is
kn = kB2 (1 + sin2 θ)
(5.93)
and the effective stiffness, Kf , acting on a metre length of foundation due to compression of
the soil along the two edges of the foundation, and shearing of the soil under the foundation is
′
Kf = 2d1 ssoil
(1 + iηsoil ) + d2 Gsoil
(5.94)
(where the foundation dimensions, d1 and d2 are shown in Fig. 5.24)
and Kφ is the stiffness of the soil resisting rotation of the foundation,
Kφ =
Gsoil d1 d22
d 3 s′ (1 + iηsoil )
+ 2 soil
2
12
(5.95)
and for the rectangular cross-section of the strip foundation, the mass moment of inertia, Ir , is
d2
ρl d1
+
(5.96)
Ir =
12 d2
d1
The transmission coefficient at a specific angle of incidence is given by
τij (θ) =
ρs,j cB,j cos θj
|T |2
ρs,i cB,i cos θi
(5.97)
The above calculations are repeated to give the angular average transmission coefficient, τij ,
using Eq. 5.6. The coupling loss factor, ηij , is calculated from τij , using Eq. 2.154; for bending
waves this is
cB,i Lij τij
(5.98)
ηij =
π2 fSi
where Lij is the length of the line connection between plate subsystems i and j.
From Wilson and Craik (1995), the compression stiffness and damping can be measured using
the method of Briaud and Cepert (1990). This measurement is based on the assumption that
the soil simply acts as a spring when a lump mass on top of the soil is excited by a vertical
force. A trench is dug in the ground to the required depth with a relatively flat surface at
its base. A concrete cuboid (≈100 kg) is then cast onto this surface to form the lump mass.
Measurement of the driving-point mobility of this cuboid can then be used to identify the mass–
spring resonance frequency; this measurement is similar to that used for the dynamic stiffness
of resilient materials (Section 3.11.3.1). The spring representing the soil can be highly damped
and it is necessary to estimate this damping from measurements. For linear springs the ratio of
the 3 dB bandwidth of the resonance peak to the resonance frequency equals the loss factor
when η < 0.3. When η ≥ 0.3 these values can only be treated as rough estimates because the
ratio is no longer linearly related to the loss factor. Example data from three different soils are
shown in Table 5.1. Note that the stiffness may vary significantly with depth below ground level.
5.3 Statistical energy analysis
The fundamental aspects of SEA that are needed to predict the combination of direct and flanking transmission have been covered in Section 4.2. Calculation of transmission coefficients for
vibration transmission between idealized plate junctions has been discussed in Section 5.2; and
580
Chapter 5
Table 5.1. Measured soil stiffness data (UK)
Type of soil
Stiff clay with large stones∗
Lower green sand
overlying sandstone#
London clay – wet#
Depth below
ground level (m)
Compression stiffness
′
per unit area, ssoil
(N/m3 )
Loss factor,
ηsoil (−)
0.45
0.9
1.96 × 109
1.63 × 108
0.96
0.2
1.5
8.67 × 107
0.4
Measurements from ∗ Wilson and Craik, and # Hopkins courtesy of BRE, ODPM, and BRE Trust.
these can be used to calculate the coupling loss factors. This section looks at how measured
data can be included within the SEA framework and then uses some examples to illustrate the
prediction of direct and flanking transmission.
5.3.1 Inclusion of measured data
The SEA framework is well-suited to the inclusion of laboratory measurements in cases where
the internal, coupling, or total loss factor of a subsystem cannot be accurately predicted or
where the airborne sound insulation of a building element is too complex to model.
5.3.1.1 Airborne sound insulation
To calculate transmission of airborne sound between adjacent rooms, laboratory measurements of airborne sound insulation can be included within the SEA framework. This is useful
when direct transmission across an element is too complicated to model and the element itself
does not affect any important flanking transmission paths. One example of this is a lightweight
separating wall between adjacent rooms where a flanking path is formed by a continuous concrete floor. The SEA framework can be used to predict structure-borne sound transmission
via the concrete floor and to incorporate the measured sound reduction index for direct sound
transmission across the separating wall. Another example occurs with a building element such
as a doorset, window or vent in a solid separating wall between adjacent rooms. It can often be
assumed that any transmission of vibration between the separating wall and this type of element will have negligible effect on the overall sound insulation. Therefore the solid separating
wall can simply be modelled, and laboratory measurements can be incorporated in the SEA
model for the doorset, window or vent.
To use the measured sound reduction index of a separating element to calculate airborne
sound transmission between two rooms, the element is not included as a subsystem in the
SEA model. We deliberately ignore the mechanisms by which sound is transmitted across
the element, such as non-resonant transmission, resonant transmission, mass–spring–mass
resonances, dilatational wave motion; these are treated as a black box of unknowns. The room
on each side of the separating element is included in the SEA model as a space subsystem
and a non-resonant transmission mechanism is used to transmit sound between them. Hence
we can convert the sound reduction index, R, into a coupling loss factor (Eq. 4.26) between
two subsystems i and j using
ηij =
c0 S
10−R/10
4ωVi
(5.99)
where S is the area of the separating element.
581
S o u n d
I n s u l a t i o n
Sound reduction index (dB)
Apparent sound reduction index (dB)
95
90
1→5
85
1→3→5
80
1→2→3→4→5
75
Matrix solution
70
65
60
55
50
45
Void (cavity)
40
3
4
35
2
30
Receiving
room (5)
25
Source
room (1)
20
Lightweight
separating wall
15
10
50
80
125
200
315
500
800
1250
2000
3150
5000
One-third-octave-band centre frequency (Hz)
Figure 5.25
Predicted airborne sound insulation between two adjacent rooms with flanking transmission via the ceiling void. SEA model
incorporates the measured sound reduction index for the lightweight separating wall. Rooms: Lx = 4 m, Ly = 3 m, Lz = 2.5 m,
T = 0.5 s. Ceiling void: Lx = 4 m, Ly = 6 m, Lz = 0.5 m, T = 0.5 s. Ceiling: 12.5 mm plasterboard, Lx = 4 m, Ly = 3 m,
ρs = 10.8 kg/m2 , cL = 1490 m/s, ηint = 0.0141.
Similarly, the coupling loss factor for the element-normalized level difference is
ηij =
c0
10−(Dn,e −10 dB)/10
4ωVi
(5.100)
The consistency relationship can be used to calculate the coupling loss factor in the reverse
direction.
5.3.1.1.1
Example
This example is based on a lightweight separating wall with flanking transmission via a ceiling
void as shown in Fig. 5.25. A laboratory measurement of the sound reduction index is used
for the lightweight separating wall (e.g. plasterboard on light steel frame); hence this wall is
not assigned to a subsystem. A five-subsystem model is used to determine the combination of
direct and flanking transmission. Note that we are effectively using a plate–cavity–plate system
as a flanking construction rather than a separating construction as in Section 4.3.5. Each plate
582
Chapter 5
for the ceiling is modelled as a single sheet of plasterboard (ignoring its supporting frame) and
it is assumed to be uncoupled from the separating wall and the ceiling in the adjacent room.
The radiation efficiency for the plasterboard is ideally calculated using method no. 2 to give an
indication of lower and upper limits near the critical frequency; however, to simplify this example
only the lower limit is used. The direct path, 1 → 5 corresponds to the sound reduction index of
the separating wall, and the matrix solution corresponds to the apparent sound reduction index.
Significant flanking transmission occurs via the non-resonant path 1 → 3 → 5 below the critical
frequency of the plasterboard, and via the resonant path 1 → 2 → 3 → 4 → 5 above the critical
frequency. For this example, summing the paths 1 → 5, 1 → 3 → 5, and 1 → 2 → 3 → 4 → 5
gives approximately the same result as the matrix solution.
This model can be adapted to other situations in buildings. For example, it could be used to
model sound transmission between two adjacent rooms with flanking transmission via a short
external corridor instead of a ceiling void. Plates 2 and 4 would then form the flanking walls
into the corridor and because the sound reduction indices for the doorsets in these walls can
be awkward to predict it would be convenient to incorporate laboratory measurements.
5.3.1.2 Coupling loss factors
Some junctions between plates, or between plates and beams are too complex to model and
it may be necessary to measure the coupling loss factor as described in Section 3.12.3.3. This
approach effectively treats the junction as a black box of unknowns and there are three points
that need to be considered.
The first point concerns the situation where an SEA model of an existing structure is being
created to gain an insight into the sound transmission paths. This usually means that the
measurement will be taken in situ rather than in the laboratory where there is some control
over the flanking paths. Depending on the measurement technique, the measured coupling
loss factor may quantify more than one transmission path. These other paths may be purely
structural but there may also be paths that involve sound radiation into spaces such as cavities;
some examples are shown in Fig. 5.26. Care is then needed to make sure that transmission
paths are not included more than once in the model. If the coupling loss factor is found to play
a role in an important transmission path, more measurements may be necessary to isolate the
individual paths. The second point concerns the omission of subsystems from the SEA model.
A common example is a plate junction where only bending waves are excited and only bending
wave motion is measured on the source and receiving plates. We recall from Section 5.2.2
that in-plane waves can also be generated at the junction. Using these measurements in
an SEA model that considers only bending waves will not usually cause significant errors
when predicting vibration transmission across one junction between adjacent rooms. (The
assumption here is that the total loss factor associated with the in-plane wave subsystems
does not vary significantly from one situation to another.) However, the measured coupling
loss factor is unlikely to be suitable when modelling transmission between more distant rooms
with transmission paths involving several junctions. The third point concerns plates with low
mode counts and low modal overlap because a single physical measurement on one junction
may be significantly different to similar junctions (refer back to the examples in Section 5.2.3).
5.3.1.3 Total loss factors
For some subsystems it is difficult to predict the total loss factor accurately. However, the total
loss factor can be critical in accurately quantifying the overall sound insulation; this occurs with
583
S o u n d
I n s u l a t i o n
(a) Lightweight construction: Junction of walls and floors
i
j
(b) Heavyweight construction: Junction of solid separating wall and external cavity wall with wall ties
j
i
Figure 5.26
Examples of two complex junctions where measurement of the structural coupling loss factor, ηij , could include more than
one transmission path.
some cavities in walls and floors, and with some plates that form walls, floors, or ground floor
slabs. Before including measurements of total loss factors in a model it is important to consider
which part of the total loss factor is difficult to predict; the internal or the coupling loss factors.
Internal loss factors are almost always determined from measurements, so this is rarely an
issue. If it is a coupling loss factor then it is necessary to consider whether the coupling losses
have been adequately dealt with in the model.
5.3.2 Models for direct and flanking transmission
Direct transmission across solid plates in the laboratory was modelled in Chapter 4 using a total
loss factor for each plate that represented the losses when coupled to the rest of the laboratory
structure. This assumed that each plate subsystem would lose energy into the laboratory
584
Chapter 5
structure but there would be negligible flow of energy from this structure into the plate. Now
we want to assess flanking transmission it is necessary to create an SEA model of the building
that will allow energy to flow in and out of each plate. A complete model will allow the total
loss factor of each subsystem to be calculated from the internal loss factor and the sum of
its coupling loss factors. However, most buildings contain some walls, floors and junctions
that are awkward or not possible to include in the model. This section uses some examples
to illustrate factors that need to be considered when restricting the model to only part of the
flanking construction and when restricting calculations to a limited number of flanking paths.
5.3.2.1 Example: SEA model of adjacent rooms
To model direct and flanking transmission between adjacent rooms we start by looking at a
T-junction. For this we can use the T-junction from Section 5.2.3.3 where the wave approach
was used to calculate vibration transmission between rigidly connected masonry walls. We will
shortly connect the same T-junction to other plates to form two rooms; hence in this example
we will set the total loss factor of each plate subsystem to the values they will have when
fully connected to these other plates. The SEA model consists of the source room, receiving
room, and the three plates that form the T-junction. The predicted airborne sound insulation is
shown in Fig. 5.27 from two SEA models; one with only bending wave motion, and the other
95
Matrix solution (Bending waves only)
90
Apparent sound reduction index (dB)
85
80
75
70
Matrix solution (Bending and in-plane waves)
S→1 →R
S→1 →3 →R
Bending waves only
S→2 →1 →R
S→2 →3 →R
S→1 →R
S→1 →3 →R
Bending and
in-plane waves
S→2 →1 →R
S→2 →3 →R
65
60
Plate 2
Plate 3
55
50
45
Source
room (S)
Receiving
room (R)
Plate 1
40
100 125 160 200 250 315 400 500 630 800 1000 1250 1600 2000 2500 3150
One-third-octave-band centre frequency (Hz)
Figure 5.27
Predicted sound insulation between two adjacent rooms where all sound transmission occurs across a T-junction of masonry
walls. The matrix solution and path analysis are shown for two SEA models (1) bending waves only and (2) bending and
in-plane waves. In both models the plate subsystems used in the path analysis are the bending wave subsystems. Plate
properties: Lx1 = 4.0 m, Lx2 = 3.5 m, Lx3 = 3.0 m, Ly1 = Ly2 = Ly3 = 2.4 m, h1 = 0.215 m, h2 = h3 = 0.1 m, ρs1 = 430 kg/m2 ,
ρs2 = ρs3 = 200 kg/m2 , cL = 3200 m/s, ν = 0.2, total loss factors correspond to those in the fully connected system of nine
plates in Fig. 5.28.
585
S o u n d
I n s u l a t i o n
with bending and in-plane wave motion. Flanking transmission is evident from the fact that the
sound insulation predicted with the matrix solution is lower than the direct path S → 1 → R.
The bending and in-plane wave model is only shown in the high-frequency range where there
are both quasi-longitudinal and transverse shear modes. Between 1000 and 2000 Hz the matrix
solutions for the two models are approximately equal; it is only above 2500 Hz that they differ by
>1 dB. However, the three flanking paths involving transmission between bending wave subsystems across one junction (S → 1 → 3 → R, S → 2 → 1 → R, and S → 2 → 3 → R) indicate
that whilst the resulting sound insulation is the same, the strengths of the individual transmission paths are quite different for the two models. Flanking path S → 2 → 3 → R becomes more
important when both bending and in-plane waves are included. For adjacent rooms it is often
the low- and mid-frequency ranges that are of most interest so it is simplest to consider only
bending wave transmission. For masonry walls in the high-frequency range there are other
factors that affect the accuracy such as thick plate theory, bonded surface finishes starting to
vibrate independently of the base wall, and a decrease in vibration with distance. Although
the coupling between plates and rooms can usually be estimated with thin plate theory up
to 4fB(thin) , these other factors mean that there is no significant increase in accuracy with the
bending and in-plane wave model for transmission between adjacent rooms. For both models,
the matrix solution is within 0.5 dB of the sum of the direct path and the three flanking paths.
This is noted because it indicates that with some constructions it may be possible to use path
analysis to estimate the overall sound insulation using paths that only cross one junction; this
approach is discussed in Section 5.4 for an SEA-based model. It is now instructive to look at
a more realistic construction where there are several flanking walls and floors.
We now incorporate the T-junction (plates 1, 2, and 3) as part of a construction forming
two adjacent rooms as shown in Fig. 5.28. The full construction has nine rigidly connected
masonry/concrete plates for which the wave approach (bending waves only, thin plate theory)
is used to calculate the transmission coefficients. The space outside the rooms is not included
in the model so we only allow sound to be transmitted between the two rooms via the walls and
floors. The concrete plates that form the ground floor and ceiling are identical except for the
high internal damping given to the ground floor to simulate losses into the ground. Using the full
construction we can now create smaller SEA models by removing certain plates, but ensuring
that the remaining plates have the same total loss factor as when they are fully connected.
Using this approach gives the predicted sound insulation for the separating wall, the separating
wall with various flanking walls and floors, and for the full construction. These are shown in
4
5
2
8
3
1
6
Source
room (S)
9
7
Receiving
room (R)
Figure 5.28
Two adjacent rooms formed from nine masonry/concrete plate subsystems. Source room: 4 × 3.5 × 2.4 m. Receiving room:
4 × 3 × 2.4 m. Plate properties: h1 = 0.215 m, h2 = h3 = h8 = h9 = 0.1 m, h4 = h5 = h6 = h7 = 0.15 m. Plates 1, 2, 3,
8, and 9: ρ = 2000 kg/m3 , cL = 3200 m/s, ν = 0.2, ηint = 0.01. Plates 4, 5, 6, and 7: ρ = 2200 kg/m3 , cL = 3800 m/s,
ν = 0.2, ηint = 0.005 for plates 4 and 5, ηint = 3f−0.5 for plates 6 and 7 to simulate a highly damped ground floor. The total
loss factor for each plate (apart from the highly damped ground floor plates) is within 2 dB of ηint + f−0.5 .
586
Chapter 5
Fig. 5.29. The sound reduction index for the separating wall alone is higher than would usually
be measured in the laboratory because it is rigidly connected to six other walls and therefore
has a relatively high total loss factor. The sound insulation decreases as flanking walls and
floors are added to the separating wall and more transmission paths are formed in models (b)
to (f ). Models (b), (c), (d), and (e) indicate the presence of important flanking transmission but
do not give an adequate estimate of the sound insulation with model (f). This is particularly
noticeable with this construction because flanking paths dominate over direct transmission via
the separating wall. In this example, the difference between models (e) and (f) shows that the
distribution of energy between the subsystems is significantly altered by including the back
walls. This indicates that making a priori decisions on which flanking walls and floors can be
excluded from a model is not always straightforward; in most cases it is better to include more,
rather than less, of the construction.
95
(c)
(b)
(a)
90
85
(a) Sound reduction index (dB)
(b,c,d,e,f ) Apparent sound reduction index (dB)
80
(d)
(e)
75
70
65
(f )
60
55
50
45
40
35
30
100 125 160 200 250 315 400 500 630 800 1000 1250 1600 2000 2500 3150
One-third-octave-band centre frequency (Hz)
Figure 5.29
Sound insulation between two adjacent rooms predicted with the SEA matrix solution (bending waves only) when the separating
wall is modelled as an isolated element in (a) and when connected to different combinations of flanking walls and floors in (b),
(c), (d), (e), and (f).
587
S o u n d
I n s u l a t i o n
For the isolated T-junctions in models (b) or (c), the matrix solution is within 1 dB of the sum of
the direct path and the flanking paths that only cross one junction. With models (d), (e), and (f)
there are many more flanking paths that cross more than one junction. Without incurring errors
of at least a few decibels it is no longer possible to determine the same values as the matrix
solution by summing the direct path and flanking paths that only cross one junction. There are
usually a great number of paths that determine the sound insulation between adjacent rooms,
as well as rooms that are far apart in a building; hence analysis of individual paths tends to
become of limited use (Craik, 1996).
5.3.2.2 Example: Comparison of SEA with measurements
Validating models of direct and flanking transmission with measurements is usually easier in a
flanking laboratory where there is more control over the quality of construction as well as the
effect of the laboratory walls and floors on the flanking transmission. Two examples are used
in this section: a solid masonry separating wall and a cavity masonry separating wall, both of
which have cavity masonry flanking walls. In both cases the SEA model included both the test
construction and the flanking laboratory.
Figure 5.30a shows an example with a solid masonry separating wall construction. The inner
and outer leaves of the external cavity wall are modelled as separate subsystems. However
the wall ties that connect these leaves have a high dynamic stiffness; so this is only appropriate above the mass–spring–mass resonance frequency (i.e. inner leaf – wall ties – outer
leaf). The ties can then be treated as point connections to calculate the coupling loss factors
(Section 4.3.5.4.1). Figure 5.30a shows that there is good agreement between measurements
and the SEA models. Note that an alternative SEA model could potentially be used below the
mass–spring–mass resonance by assuming that the inner and outer leaves act as a single
plate, but this is not pursued here.
Figure 5.30b shows some of the short transmission paths involving the separating and flanking walls using the bending wave model. This indicates the importance of the flanking path
via the wall ties in the external cavity wall, S → 2 → 4 → 3 → R compared with the direct path
S → 1 → R. Paths via the external cavities are relatively unimportant due to the cavity closer
at the junction. Summing the direct path, path S → 2 → 4 → 3 → R, and the flanking paths that
only cross one junction overestimates the standardized level difference from the matrix solution
by ≈3 dB in each frequency band.
Figure 5.30c shows the difference between measured and predicted velocity level differences
with structure-borne excitation of walls 1, 2, and 4 using a shaker. Prediction of velocity level
differences using SEA (wave approach) gives smooth curves that are in contrast to the fluctuating curves expected from masonry walls with low mode counts and low modal overlap.
In the context of the measurement uncertainty and the assumptions made in using the wave
approach, Fig. 5.30c shows good agreement between measurements and the SEA model.
Figure 5.31 shows an example with a cavity masonry separating wall on a split foundation. Only
bending waves were considered in the SEA model. In Section 4.3.5.2.2, structural coupling
via the foundations was shown to play an important role in direct transmission across these
walls. The coupling loss factor for the foundation coupling, and reverberation times for the
separating and flanking wall cavities were measured and incorporated in the model because
they were difficult to predict accurately (Hopkins, 1997). The wall ties in the external cavity wall
have a low dynamic stiffness giving a resonance frequency below 50 Hz. Comparing the path
588
Chapter 5
(a) Comparison of an SEA model with airborne sound insulation measurements in a flanking laboratory
100
95
Receiving room (R)
(Ground floor)
90
85
80
Standardized level difference, DnT (dB)
75
70
65
Source room (S)
(Ground floor)
1
3
Plan view:
Test construction and
flanking laboratory
walls
2
4
Test construction:
Solid masonry separating wall
Cavity masonry flanking walls
Ceiling: Plasterboard on battens fixed to concrete plank floor
Flanking cavities: Empty, separated by a mineral fibre cavity closer along
the edge of the separating wall,with wall ties (high dynamic stiffness)
Measured
60
55
50
SEA matrix solution
(bending waves only)
SEA matrix solution
(bending and in-plane
waves)
45
40
35
30
25
SEA model used
above the cavity
wall resonance due
to the wall ties
fmsm ⫽ 162 Hz
Flanking laboratory.
Test construction
forms two rooms on
the ground floor.
Below ground
20
100 125 160 200 250 315 400 500 630 800 1000 1250 1600 2000 2500 3150
One-third-octave-band centre frequency (Hz)
Figure 5.30
Direct and flanking transmission across a solid masonry wall construction. Separating wall: One leaf of 215 mm, 430 kg/m2
solid masonry (plaster finish). Flanking cavity wall: Inner leaves of 100 mm, 150 kg/m2 solid masonry (plaster finish). Outer
leaf of 100 mm, 170 kg/m2 solid brick. 100 mm cavity (2.5 wall ties/m2 , s100mm = 43.4 MN/m). Ceiling and separating floor:
9.5 mm plasterboard fixed to 45 × 45 mm timber battens fixed to 150 mm, 300 kg/m2 concrete slabs. Flanking laboratory:
560 mm, 900 kg/m2 solid brick walls, 125 mm cast concrete ground floor slabs on hardcore. Measured data from Hopkins are
reproduced with permission from ODPM and BRE.
S → 1 → 2 → R across the split foundations with the matrix solution in Fig. 5.31 indicates that
this path is relatively unimportant. There are too many paths to make path analysis a practical
option to determine the overall sound insulation. Hence it is the matrix solution that is needed
and this gives a good estimate of the measured sound insulation.
589
S o u n d
I n s u l a t i o n
(b) SEA path analysis between the source and receiving rooms (bending waves only)
100
S→1→R
S → 1 → 3 → R (= S → 2 → 1 → R)
S→2→3→R
S→2→4→3→R
Matrix solution
Standardized level difference, DnT (dB)
95
90
85
80
75
70
65
60
55
50
45
40
35
30
100 125 160 200 250 315 400 500 630 800 1000 1250 1600 2000 2500 3150
One-third-octave-band centre frequency (Hz)
(c) Comparison of SEA model with measurements of velocity level differences (bending wave vibration)
between walls. This is shown in terms of the difference between measured and SEA velocity level
differences, Dv,ij, where i is the source subsystem. SEA model: 100–800 Hz (bending waves only), 1000–
3150 Hz (bending and in-plane waves). The 95% confidence intervals correspond to the measurements.
10
Difference between Dv,ij
Measured – SEA (dB)
5
0
⫺5
Dv,12
Dv,23
Dv,24
Dv,21
Dv,32
Dv,42
⫺10
100 125 160 200 250 315 400 500 630 800 1000 1250 1600 2000 2500 3150
One-third-octave-band centre frequency (Hz)
Figure 5.30
(Continued)
As with transmission suite measurements, the structure of the flanking laboratory also plays a
role in determining the measured sound insulation. For the cavity separating wall this is obvious from the way that the foundations form part of the laboratory, but even the heavy walls
and floors of a laboratory can participate in flanking transmission. Two options are usually
590
Chapter 5
100
Receiving room (R)
(Ground floor)
95
2
90
Plan view:
Test construction and
flanking laboratory
walls
Source room (S)
(Ground floor)
1
85
Test construction:
Cavity masonry separating wall on split foundation
Cavity masonry flanking walls
Timber joist separating floors
Separating cavity: Empty, no wall ties.
Flanking cavity: Empty, with wall ties (low dynamic stiffness)
80
Standardized level difference, DnT (dB)
75
70
65
60
55
Measured
50
SEA matrix solution
(bending waves only)
SEA path S →1→2 →R
(via foundation)
45
40
35
Flanking laboratory.
Test construction forms
four rooms.
Values above correspond
to sound insulation between
ground floor rooms.
30
25
Below ground
20
50
63
80
100
125
160
200
250
315
400
500
630
800
1000
One-third-octave-band centre frequency (Hz)
Figure 5.31
Direct and flanking transmission across a cavity masonry wall construction. Comparison of SEA model with airborne sound
insulation measurements in a flanking laboratory. (Note: The frequency range is restricted to between 50 and 1000 Hz where
measurements of the coupling loss factor were available for structural coupling between the separating cavity wall leaves via
the foundations.) Separating cavity wall: Two leaves of 100 mm, 166 kg/m2 solid masonry (plaster finish). 60 mm empty cavity
(no wall ties). Flanking cavity wall: Inner leaves of 100 mm, 70 kg/m2 solid masonry (plaster finish). Outer leaf of 100 mm,
154 kg/m2 solid brick. 75 mm cavity (2.5 wall ties/m2 , s75mm = 1.7 MN/m). Measured data from Hopkins are reproduced with
permission from ODPM and BRE.
considered when building a test construction into a flanking laboratory: the first is to isolate the
test construction from the laboratory using resilient layers, the second is to rigidly connect the
test construction to the laboratory. The first option reduces the coupling losses from each wall
or floor; if this causes low modal overlap then the coupling loss factors can be expected to vary
591
S o u n d
I n s u l a t i o n
significantly between similar constructions (see examples in Section 5.2.3). Note that for structural reasons it is not always possible to completely isolate the test construction on resilient
layers. The second option means that the total loss factors and modal overlap factors are more
representative of in situ; but there is potential for the laboratory structure to participate in the
flanking transmission. In these examples the test construction was rigidly connected to the laboratory. The SEA models were used to estimate the effect of including the laboratory walls and
floors on the measured sound insulation. The predicted effect on the sound insulation varied
with frequency; the range was only 0.1–1 dB for the cavity wall, but 1–4 dB for the solid wall. This
reinforces the point that it is better to include as much of the construction in the model as
possible.
The examples in this section indicate that variation in the sound insulation performance of
nominally identical separating walls or floors will not only be due to workmanship, but also
due to different flanking constructions. This needs to be placed in the context of guidance
documents that describe building constructions with the potential to achieve certain values
of sound insulation. To avoid completely dictating the building design it is common to focus
on the separating wall or floor, and the flanking walls or floors that are connected to it (e.g.
see DIN 4109 1989; Homb et al., 1983). Such guidance is usually based on field sound
insulation measurements, so it is possible to make some statistical inference about the likelihood of achieving a certain value of sound insulation. Some account is therefore taken
of workmanship and of walls and floors that are not directly connected to the separating
element.
5.4 SEA-based model
For practical purposes it is very useful to have a model for direct and flanking transmission
that can incorporate laboratory sound insulation measurements. This approach is taken in the
model developed by Gerretsen (1979, 1986, 1994, and 1996) and subsequently implemented
in the Standard EN 12354. The model for airborne and impact sound insulation can either
be derived directly from classical diffuse field theories of sound and vibration, or by using
SEA to provide the framework for the same classical theories (Gerretsen, 1979). The resulting
model is affected by the same limitations that affect SEA; and this is made explicit here by
referring to it as an SEA-based method. By deriving the model from an SEA perspective it is
easier to see the links between SEA loss factors and laboratory measurements of the sound
reduction index, structural reverberation times, and the vibration reduction index. The SEA
framework also simplifies discussion of the roles of resonant and non-resonant transmission.
The model is essentially the same as SEA path analysis between adjacent rooms, but the
flanking transmission paths are restricted in their length. Each flanking path is restricted to
vibration transmission across no more than one plate junction as shown in Fig. 5.32 (Gerretsen,
1979). This approach is quite commonly taken in documents giving construction guidance for
sound insulation in buildings.
5.4.1 Airborne sound insulation
The aim is to determine the apparent sound reduction index; this is easily converted to other
in situ sound insulation descriptors (e.g. Dn , DnT ) at the end of the calculation. Using SEA
592
Chapter 5
Source room
Direct path Dd
Receiving room
Flanking path Ff
Flanking path Fd
Flanking path Df
Source room
Source room
Source room
i
i
i
j
j
j
Receiving room
Receiving room
Receiving room
Figure 5.32
Direct and flanking transmission paths between two adjacent rooms (only one junction is shown). The length of each flanking
path is restricted to vibration transmission across no more than one junction. Direct transmission via the separating element
is indicated by D in the source room, and d in the receiving room. Flanking surfaces are indicated by F in the source room and
f in the receiving room. Flanking paths Ff, Fd, and Df apply to airborne sound insulation. Flanking path Df applies to impact
sound insulation.
path analysis, the apparent sound reduction index for P transmission paths is determined from
Eq. 4.18 as
VR
ES
S
+ 10 lg
(5.101)
+ 10 lg
R ′ = 10 lg
ER Due to
VS
A
P paths
where ES and ER are the energies for the source and receiving rooms respectively.
The transmission paths need to be combined outside the framework of SEA so it is useful to
define a flanking sound reduction index, Rij (Gerretsen, 1979). This is based on the ratio of the
sound power, W1 , incident on a reference area, SS , in the source room to the sound power,
Wij , radiated by flanking plate j due to sound power incident on plate i in the source room. This
is given in decibels by
W1
Rij = 10 lg
(5.102)
Wij
Note that Rij has not been written in terms of a transmission coefficient, τij . This is because τij
has already been defined in Eq. 5.1 to describe bending wave transmission between plates i
and j and it will shortly be needed again.
We will temporarily assume that the total loss factor of each plate connected in situ is the same
as when its sound reduction index is measured in the laboratory; this assumption will shortly
593
S o u n d
I n s u l a t i o n
be removed when generalizing the model. By setting the reference area, SS , to equal the area
of the separating plate the apparent sound reduction index is determined using
⎛
⎞
P
−R
/10
−R
/10
′
10 ij ⎠
R = −10 lg ⎝10 Dd +
(5.103)
p=1
where RDd is the sound reduction index of the separating plate; a combination of non-resonant
and resonant transmission.
It is assumed that the separating and flanking plates support a reverberant bending wave field
with no significant decrease in vibration with distance across their surface. Using SEA path
analysis, the flanking sound reduction index is
ηi ηj ηR
ES
VR
SS
V R SS
Rij = 10 lg
= 10 lg
+ 10 lg
(5.104)
+ 10 lg
+ 10 lg
ER
VS
A
ηSi ηij ηjR
VS A
where the subscripts are: S for the source room, R for the receiving room, i for plate i in the
source room, and j for plate j in the receiving room.
There are two types of coupling loss factor in Eq. 5.104. Both of these involve a reverberant
bending wave field on the plates; ηSi and ηjR involve coupling between a plate and a room, and
ηij involves vibration transmission resulting in a bending wave field on each of the two coupled
plates. The former can be written in terms of the resonant sound reduction indices and the
latter can be written in terms of a direction-averaged velocity level difference. Hence we can
start to remove any direct reference to the loss factors in Eq. 5.104.
The resonant sound reduction index is given in terms of loss factors by Eq. 4.22. The plates are
defined so that one side of plate i will face into the source room and one side of plate j will face
into the receiving room. Hence regardless of whether the other side of plate i faces into the
receiving room, or the other side of plate j faces into the source room it is useful to determine
the resonant sound reduction indices using the same source and receiving room volumes and
the same receiving room absorption as the adjacent rooms. This gives the resonant sound
reduction indices for plates i and j as
ηj ηR VR Sj
ηi ηR VR Si
RResonant,i = 10 lg
and RResonant,j = 10 lg
(5.105)
ηSi ηiR VS A
ηSj ηjR VS A
Both of the resonant sound reduction indices can now be rewritten purely in terms of ηSi and ηjR
by using the consistency relationship (Eq. 4.2) and the fact that ηiS = ηiR and ηjS = ηjR ; this gives
ηj ηR nS VR Sj
ηi ηR ni VR Si
and RResonant,j = 10 lg
(5.106)
RResonant,i = 10 lg
η2Si nS VS A
η2jR nj VS A
We now need to combine the two resonant sound reduction indices in such a way that Eq. 5.104
can be rewritten in terms of the resonant sound reduction indices. This is done by combining
them as follows,
RResonant,j
ηi ηj Si Sj ni
RResonant,i
ηR
VR
+
= 10 lg
+ 5 lg
(5.107)
+ 10 lg
2
2
ηSi ηjR
VS A
nj
This allows Eq. 5.104 to be given in terms of the resonant sound reduction indices whilst leaving
vibration transmission between the two plates in terms of a coupling loss factor,
RResonant,j
η i η j nj
1
RResonant,i
SS
Rij =
+
+ 10 lg
(5.108)
+ 5 lg
+ 10 lg
2
2
ηij
ni
Si Sj
594
Chapter 5
The next step is to look at vibration transmission between plates i and j for which we wish to
remove ηij and use a velocity level difference instead. To do this we take a blinkered view of
the entire construction and focus on the two coupled plates to create a two-subsystem SEA
model for plate subsystems i and j. Under the assumption that with excitation of subsystem i
there is negligible power flow back from subsystem j to i, the coupling loss factor can be written
in terms of the velocity level difference, Dv,ij (Section 3.12.3.3)
ηj
mi
− 10 lg
(5.109)
Dv,ij = 10 lg
ηij
mj
Equations 5.108 and 5.109 indicate that if we also calculate the velocity level difference, Dv,ji ,
and write it in terms of ηij using the consistency relationship we will have all the relevant
terms that we need to rewrite Eq. 5.108 in terms of velocity level differences. For excitation of
subsystem j, Dv,ji , is calculated in a similar way to Dv,ij giving
ηi nj
mi
mi
ηi
Dv,ji = 10 lg
(5.110)
+ 10 lg
= 10 lg
+ 10 lg
ηji
mj
ηij ni
mj
Dv,ij and Dv,ji can now be combined to give the direction-averaged velocity level difference,
Dv,ij , as
Dv,ij + Dv,ji
ηi ηj nj
1
= 10 lg
Dv,ij =
(5.111)
+ 5 lg
2
ηij
ni
Substituting Eq. 5.111 in Eq. 5.108, we can now rewrite Eq. 5.104 purely in terms of the
resonant sound reduction indices and the direction-averaged velocity level difference,
RResonant,j
RResonant,i
SS
Rij =
+
+ Dv,ij + 10 lg
(5.112)
2
2
Si Sj
5.4.1.1
Generalizing the model for in situ
Now it is useful to remove the assumption that the total loss factor of each plate in situ is
the same as when the sound reduction index is measured in the laboratory. This requires the
structural reverberation time of each plate to be measured whilst it is installed in the laboratory for the sound reduction index measurement. It is then necessary to predict the structural
reverberation time for each plate when installed in situ. The resonant sound reduction index of
each plate measured in the laboratory is converted to the in situ value using
Ts,situ
Rsitu = RResonant − 10 lg
(5.113)
Ts,lab
where the structural reverberation times in situ and in the laboratory are Ts,situ and Ts,lab respectively. If the resonant sound reduction index has been predicted then Ts,lab corresponds to the
value used in the prediction model.
For the direct path, RDd,situ needs to include both non-resonant and resonant transmission as
measured in the laboratory; however the conversion in Eq. 5.113 only applies to the resonant
component. This may require estimating the non-resonant component of the sound reduction
index so that it can be removed. The conversion is then carried out on the resonant component;
and after the conversion the non-resonant component is re-introduced.
Laboratory measurements of vibration transmission between plates also need to be converted
to a direction-averaged velocity level difference that corresponds to the in situ situation. This is
595
S o u n d
I n s u l a t i o n
achieved by defining a vibration reduction index, Kij (Gerretsen, 1996) given by Eq. 3.252. This
can either be measured in the laboratory or predicted. From Eq. 3.252, the direction-averaged
velocity level difference for the in situ situation is given by
Lij
Dv,ij,situ = Kij − 10 lg √
(5.114)
ai,situ aj,situ
where ai,situ and aj,situ are the absorption lengths for plates i and j in situ, and Lij is the junction
length between elements i and j.
The in situ structural reverberation time for Eq. 5.113 and the in situ absorption lengths for Eq.
5.114 are related to the in situ total loss factor. For masonry/concrete plates that are rigidly
connected on all sides, estimates for the in situ total loss factor are discussed in Section 2.6.5.
Various aspects relating to the measurement of Kij are discussed in Section 3.12.3. To calculate
Kij it is useful to note the relationship between the vibration reduction index and the coupling loss
factor. This can be determined from Eq. 5.111. By assuming that i and j are solid homogenous
plates, Kij can be written in terms of the critical frequencies as
c02 L2ij fc,j
1
+ 5 lg
(5.115)
Kij = 10 lg
ηij
π4 Si2 fc,i fref f
from which the relationship to the bending wave transmission coefficient (Eq. 5.1) is
fc,j
1
+ 5 lg
Kij = 10 lg
τij
fref
(5.116)
where fref is a reference frequency of 1000 Hz.
Hence Kij can be calculated using the wave approach (bending waves only) or by using a
numerical method as discussed in Section 5.2.
As most walls and floors have linings on one or both sides it is useful to incorporate laboratory
measurements of the sound reduction improvement index. We recall that the resonant sound
reduction index is used to replace the coupling loss factor between a plate and a room. To use
laboratory measurements of the sound reduction improvement index for a lining; we need to
use the resonant sound reduction improvement index (Sections 3.5.1.2.2 and 4.3.8.2). With
reference to the coupling loss factor (Eq. 4.21) the lining can be viewed as modifying the
radiation efficiency of the plate. In practice there may be other effects due to interaction between
the lining and the base wall or floor (Section 4.3.8.2). One of these concerns the mass–spring–
mass resonance frequency which means that measurements of RResonant can have highly
negative values near the resonance. A practical solution is to use R instead of RResonant
and accept that it sometimes gives an overestimate; hence the in situ value is given by
(5.117)
Rsitu = RResonant ≈ R
The in situ sound reduction index of the separating plate is now given by
(5.118)
RDd = RDd,situ + RD,situ + Rd,situ
and the in situ flanking sound reduction index for each path is
Rj,situ
SS
Ri,situ
+ Rj,situ + Dv,ij,situ + 10 lg
+ Ri,situ +
Rij =
2
2
Si Sj
(5.119)
The apparent sound reduction index is calculated using Eqs 5.118 and 5.119 in Eq. 5.103.
596
Chapter 5
5.4.2
Impact sound insulation
The model for impact sound insulation can be derived from SEA path analysis in a very similar
way to airborne sound insulation. It is assumed that the plates support a reverberant bending
wave field with no significant decrease in vibration with distance across their surface. As with
airborne sound insulation, the resonant sound reduction indices and the direction-averaged
velocity level difference are used to replace the coupling loss factors. The main difference is
the inclusion of the power input from the ISO tapping machine; this was previously discussed
in Sections 3.6.3 and 4.4.1.
The aim is to determine the normalized impact sound pressure level. In order to describe
the different transmission paths a flanking normalized impact sound pressure level, Ln,ij , is
defined as the normalized impact sound pressure level due to sound radiated by flanking plate
j (wall/floor) in the receiving room with excitation of floor plate i by the ISO tapping machine.
The normalized impact sound pressure level in situ can then be given by
⎛
⎞
L
/10
′
L
/10
Ln = 10 lg ⎝10 n,d +
(5.120)
10 n,ij ⎠
j
where Ln,d is the in situ normalized impact sound pressure level for direct transmission via the
separating floor. Note that Ln,d is not relevant to horizontally adjacent rooms.
The normalized impact sound pressure level for the floor measured in the laboratory is
converted to the in situ value using
Ts,situ
Ln,situ = Ln + 10 lg
(5.121)
Ts,lab
where the structural reverberation times in situ and in the laboratory are Ts,situ and Ts,lab respectively. Note that if Ln has been predicted then Ts,lab corresponds to the value used in the
prediction model.
The in situ normalized impact sound pressure level for direct transmission via the separating
floor is
Ln,d = Ln,situ − Lsitu − RResonant
and the in situ flanking normalized impact sound pressure level for each path is
RResonant,j
Si
RResonant,i
−
− Rj,situ − Dv,ij,situ − 5 lg
Ln,ij = Ln,situ − Lsitu +
2
2
Sj
(5.122)
(5.123)
where Lsitu applies to a floor covering (Lsitu = L for concrete floors), RResonant applies to
a ceiling and Rj,situ (Eq. 5.117) applies to a lining facing into the receiving room on plate j.
5.4.3
Application
The aim is to use the SEA-based model to make a link between sound insulation measured in
the laboratory and in situ. This link is made via the resonant sound reduction index. The sound
reduction index that is measured in the laboratory represents a combination of non-resonant
and resonant transmission below the critical frequency; it only represents resonant transmission at and above the critical frequency. For this reason it is useful to write the equations
597
S o u n d
I n s u l a t i o n
in terms of RResonant to avoid giving the impression that it applies to all measured R values
at all frequencies. When using measured R values in the model, the valid frequency range
depends on the critical frequencies of the plates. To allow predictions over a wide frequency
range the model is well-suited to plates with critical frequencies in the low-frequency range;
this applies to many masonry/concrete plates for which the measured R can often be used
as a reasonable estimate for RResonant below the critical frequency (see examples in Section
4.3.1.3 for solid masonry walls). This is in marked contrast to lightweight plates with high critical frequencies where non-resonant transmission dominates over the majority of the building
acoustics frequency range (see examples in Section 4.3.1.3 for glass and plasterboard). For
some lightweight walls and floors it may be possible to predict the non-resonant component and
extract it from the measured sound reduction index; this assumes that there is no significant
decrease in the bending wave vibration level across their surface. This complication does not
apply to solid homogenous plates (lightweight or heavyweight) for which the resonant sound
reduction index can be predicted (Section 4.3.1.1).
We now consider the link with the resonant sound reduction index in terms of the assumption
that each plate has a reverberant bending wave field. For thick plates (Section 4.3.1.4), plates
formed from certain types of hollow bricks/blocks (Section 4.3.2.3) and sandwich panels (Section 4.3.6), there are thickness resonances and dilatational waves that give pronounced dips in
the measured sound reduction index. The SEA-based model uses the resonant sound reduction index to determine bending wave vibration on plate i that is transmitted across a junction
to give a reverberant bending wave field on plate j. Hence at frequencies where plate motion
other than bending wave motion determines the measured sound reduction index, the measured R no longer represents RResonant . Therefore the valid frequency range for plates which
are not solid and homogeneous is limited to frequencies where non-resonant transmission is
negligible and where sound radiation is only due to bending wave motion. Some external flanking walls are formed from a number of rigid and resilient layers for thermal purposes (Lang,
1993). These are not usually referred to as sandwich plates or identified as a base plate with
a lining. However, they exhibit mass–spring–mass resonances which also give rise to dips in
the measured sound reduction index where the measured R does not represent RResonant .
The limitations discussed above stem from the fact that different sound transmission mechanisms are ‘contained within’ the measured sound reduction index. A plate–cavity–plate system
not only ‘contains’ different transmission mechanisms but also different transmission paths
(Section 4.3.5). Any extension of the SEA-based model to lightweight or heavyweight cavity
walls or floors requires knowledge of the transmission paths both in the laboratory and in situ
because the structural coupling between the plates may be different in each case. In addition,
there may be important flanking paths that cross more than one junction, or involve cavities in
the flanking construction. If an SEA model is suitable for a plate–cavity–plate system then it
is simpler to work within the SEA framework (see example in Section 5.3.2.2). However, SEA
models are not always available for direct and flanking transmission across lightweight walls
and floors; particularly those that are built from layers of boards and resilient components.
Lightweight walls and floors sometimes show a significant decrease in vibration across the
element when they are excited along a junction line or by the ISO tapping machine; in such
cases it may be reasonable to only consider short flanking paths because the long paths are
less important. Ongoing research shows that there is the potential to modify the SEA-based
model for lightweight constructions (e.g. see Guigou-Carter et al., 2006; Nightingale, 1995;
Schumacher and Sass, 1999; Villot and Guigou-Carter, 2006).
598
Chapter 5
In discussing limitations of SEA and the SEA-based model it is important to note that there are
many flanking paths and the ones that cannot be easily modelled are sometimes unimportant.
As an example we can consider a construction in which a plate–cavity–plate system is conveniently incorporated into the SEA-based model. Consider airborne sound insulation across a
lightweight separating floor, such as a timber joist floor, where masonry/concrete walls i and
j form the flanking path Ff (shown on Fig. 5.32). Direct transmission across the timber joist
separating floor may be rather complex to predict, so we use the laboratory measurement of
the sound reduction index to give RDd for the model. Depending on joist orientation and the way
they are fixed to the walls it may be reasonable to assume that the joists will have negligible
effect on flanking path Ff; hence plates i and j can be treated as an in-line junction. In some
cases it can also be assumed that vibration transmission between the joists and the flanking
walls is negligible so that flanking paths Fd and Df can be ignored.
In the above example the number of paths was reduced in order to make a partial assessment
of the flanking transmission. With the SEA-based model it is assumed that transmission paths
involving more than one junction will have negligible effect on the overall sound insulation. With
SEA this assumption can be checked by comparing path analysis with the matrix solution; this
was previously done in Section 5.3.2. Flanking paths that cross more than one junction may
be relatively weak, but as there are so many of them between adjacent rooms in a building
they can be significant when they are combined together. Hence without the matrix solution
there is uncertainty over the importance of the longer flanking paths. For buildings comprising
only rigidly connected solid masonry/concrete walls and floors, SEA models indicate that if
the airborne sound insulation for the direct path is X dB, and for all transmission paths it is
Y dB (SEA matrix solution), then the sound insulation predicted using just the direct path
and the flanking paths involving one junction is approximately (X + Y )/2 dB in any frequency
band over the building acoustics frequency range (Craik, 2001). So if X − Y is 10 dB, the
SEA-based model will overestimate the airborne sound insulation by 5 dB. Whilst this indicates
the potential importance of paths that cross more than one junction, any bias error will differ
between different types of construction.
It is important to note that bias errors have not been found in some comparisons of measured
and predicted single-number quantities for airborne and impact sound insulation (Gerretsen,
1979; Pedersen, 1999); other comparisons do indicate a bias error but this could be attributed
to the input data (Metzen, 1999). Good agreement between the SEA-based model and
measurements has also been shown for the sound insulation spectrum (Gerretsen, 1994).
There are some confounding factors that complicate the comparison of SEA and the SEAbased model in real buildings, particularly dwellings. Firstly, these do not usually have large
flanking walls that are devoid of openings such as doors and windows; for some types of
construction, including or excluding these walls in a model will make a significant difference.
Secondly, masonry/concrete walls and floors can have low mode counts and low modal overlap
in the low and mid-frequency ranges. This can result in large differences in vibration transmission between similar junctions of different size walls and floors; in addition there is a tendency
for the wave approach to overestimate the strength of vibration transmission (Section 5.2.3.3).
These two issues are not avoided by measuring vibration reduction indices or coupling loss
factors on junctions in the laboratory. This is because the inclusion of coupling parameters
from isolated junctions into larger SEA models does not always improve their accuracy when
the plates that form each junction have low mode counts and low modal overlap (see example
in Section 5.2.3.6). An additional complication occurs when these measurements are made on
599
S o u n d
I n s u l a t i o n
a junction in situ because there may be unwanted vibration transmission via paths involving
more than one junction (refer back to Fig. 3.96).
Both SEA and the SEA-based models have their limitations; it is a case of working with their
strengths and trying to work around their weaknesses. There are many walls, floors, and linings
that require laboratory measurements because they cannot be modelled easily or accurately
with a complete SEA model. Therefore any insight gained into flanking transmission with the
SEA-based model is useful; an allowance can often be made for neglecting the longer flanking
paths. A partial assessment of the flanking transmission that identifies one important flanking
path can still be used to make design decisions and to help find solutions to existing sound
insulation problems. An example of this is given in Section 5.4.4.
The SEA-based model is not restricted to one type of construction, but just as with SEA, it is
often necessary to restrict its use to certain frequency ranges. Use of the SEA-based model
can be considered in two separate categories: the first is to provide insight that helps solve
a sound insulation problem, the second is to predict single-number quantities for airborne or
impact sound insulation. The former is the easiest because there is plenty of flexibility; the
SEA-based model can be combined with SEA path analysis, other prediction models, and
other measurements. Sound insulation problems are often solved by looking at specific parts
of the frequency range. It is therefore of less concern that we do not have a model which covers
the entire building acoustics frequency range; it is possible to use a mixture of measurements,
predictions and empiricism. This is not the case when a single model is required to calculate
single-number quantities for regulatory purposes. In most countries there are preferred or
‘traditional’ types of wall and floor construction for which there are a limited number of practical
or feasible combinations. By using a database of field and laboratory measurements it is often
possible to make simplifications or empirical corrections to the SEA-based model that will give
a reasonable estimate of the in situ sound insulation in terms of the average single-number
quantity. As an example of this we consider the fact that the measured Kij can be frequencydependent with large fluctuations in the low-frequency range that depend on the dimensions
and how the plates are connected. However, to simplify calculations a frequency-average Kij
value is often used in the SEA-based model. Using such an approach is a pragmatic solution
to an awkward problem; but care needs to be taken to ensure that it does not give a reasonable
estimate of the average single-number quantity for the wrong reason. The desire to have one
model for all types of construction means that the SEA-based model tends to evolve into a
calculation procedure. This procedure can be a step removed from the original model and
so the links to the original assumptions may be changed or severed. This often suits the
intended purpose once the procedure is validated and kept within well-defined constraints. It
also provides a pragmatic, simple solution for predicting sound insulation in buildings which
can be quite complex for even the simplest constructions.
5.4.4 Example: Flanking transmission past non-homogeneous separating walls or
floors
This example concerns non-homogenous separating walls or floors that have a significant
decrease in vibration across their surface; these are awkward to include in SEA or SEA-based
models. However, a partial assessment of the combined direct and flanking transmission may
be possible when the flanking elements are homogenous and support a reverberant bending
wave field. An example of this occurs with airborne sound insulation of a beam and block
600
Chapter 5
100 mm solid masonry
flanking walls i and j
(Lx ⫽ 4 m, Ly ⫽ 2.4 m,
ρs ⫽ 142 kg/m2, cL ⫽ 2220 m/s)
Measured Kij
Separating floor:
150 mm beam and block floor
(Lx ⫽ 4 m, Ly ⫽ 4 m, ρs ⫽ 313 kg/m2)
Vibration reduction index, Kij (dB)
i
i
j
Separating
floor
j
300 mm
Predicted Kij using wave approach
(rigidly connected plates – bending waves only)
Separating floor:
150 mm cast in-situ concrete slab (330 kg/m2)
20
15
10
5
0
100 125 160 200 250 315 400 500 630 800 1000 1250 1600 2000 2500 3150
One-third-octave-band centre frequency (Hz)
Figure 5.33
Vibration reduction index (path Ff ) for a beam and block floor with masonry flanking walls (measured) compared with a solid
homogenous isotropic concrete plate (predicted). Measured data from Hopkins are reproduced with permission from ODPM
and BRE.
separating floor with masonry flanking walls (Hopkins, 2004). The decrease in vibration with
distance across this beam and block floor was previously shown in Fig. 2.42.
In this example there are two aspects that are difficult to model, direct transmission across the
separating floor and vibration transmission across the junction. This makes it more convenient to use the SEA-based model and incorporate laboratory measurements. The measured
vibration reduction index is shown in Fig. 5.33 for path Ff across a T-junction; note that we are
ignoring paths Fd and Df which involve the non-homogenous floor. A prediction is also shown
for the same masonry flanking walls but with a homogeneous concrete separating floor; this
provides an example of a similar floor that can support a reverberant bending wave field without
a significant decrease in vibration. The measured Kij is frequency-dependent and shows no
indication that it can be modelled as a junction of rigidly connected plates using either of the
wave approaches in Section 5.2. In the low-frequency range this measurement may not be
representative of an ensemble of similar junctions because Mav < 1 and Ns < 5. However, in
this case there are large differences between the measured Kij with the beam and block floor
601
S o u n d
I n s u l a t i o n
and the predicted Kij with a concrete floor. This acts as a useful reminder of the errors that can
be incurred by making gross assumptions about walls, floors and junctions.
We can now calculate the apparent sound reduction index for the combination of the direct
path Dd with flanking paths Ff. This is only a partial assessment of the flanking transmission
because the overall sound insulation will potentially be reduced by flanking paths Fd and Df as
well as many other flanking paths involving more than one junction. For this example we will
assume that flanking path Ff is identical on all four sides of the floor; in practice Kij will differ
slightly on two sides because of the beam orientation. The in situ beam and block floor will
usually have a lining on each side so we can use a laboratory measurement of the floor with a
lightweight floating floor and ceiling treatment to give RDd . It is reasonable to assume that these
lightweight linings will not change the measured Kij . The masonry flanking walls are assumed to
have a bonded surface finish (e.g. plaster). The resonant sound reduction index can therefore
be predicted because the walls can be treated as homogeneous isotropic plates; the in situ
total loss factor for each wall is estimated to be 0.01 + 0.5f −0.5 , and the radiation efficiency
is calculated using method no. 3 (Section 2.9.4.3). The various sound reduction indices are
shown in Fig. 5.34. R ′ is significantly lower than RDd due to the four flanking transmission
paths (Ff). This provides an adequate basis on which to investigate various changes to the
design that could increase the sound insulation, such as using wall linings on the flanking
90
85
RDd for a 150 mm beam and block floor with lightweight
floating floor and ceiling (laboratory measurement)
80
Sound reduction index (dB)
75
Rij predicted using Kij measurement
70
65
R⬘ predicted using four
flanking paths (Rij)
60
55
50
45
40
35
30
25
100 125 160 200 250 315 400 500 630 800 1000 1250 1600 2000 2500 3150
One-third-octave-band centre frequency (Hz)
Figure 5.34
Sound reduction indices for a beam and block floor: direct airborne sound insulation (RDd ) and predicted sound insulation
in situ (R′ ) due to flanking path Ff with masonry flanking walls on all four sides of the floor.
602
Chapter 5
walls. The fact that we have been able to identify significant flanking transmission indicates
the benefit in using SEA or an SEA-based model even when all the paths cannot be included.
There are many different types of beam and block floor; Kij measurements tend to indicate
significant differences compared to junctions of rigidly connected, solid, homogeneous plates
(e.g. see Cortés et al., 2002).
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Appendix: Material properties
T
his appendix contains indicative material properties for use with the equations in the
main text; manufacturers should be consulted for the material properties of any specific
product. Note that the properties in Tables A2, A3, and A4 will not always allow a
definitive comparison of similar products because of the variation that exists between different
manufacturers in different countries.
Table A1 Properties of gases
Gas
Ratio of specific Molar mass,
heats,γ
M(kg/mol)
Phase velocity at Density, ρ at
NTP* (m/s)
NTP* (kg/m3 )
Air (dry)
Argon
Carbon dioxide
Nitrogen
Oxygen
Sulphur hexafluoride
1.41
1.67
1.33
1.41
1.41
1.33
345
319
271
350
328
149
0.02895
0.040
0.044
0.028
0.032
0.146
1.205
1.662
1.842
1.165
1.331
6.2
*Normal temperature and pressure (20◦ C, 1.013 × 105 Pa).
607
Table A2 Material properties
Material name
Aircrete/Autoclaved Aerated Concrete
(AAC) blocks (solid) connected with
mortar or thin joint compound (Hopkins)
Aluminium (Heckl, 1981)
Density, ρ
(kg/m3 )
400–800
2700
Quasi-longitudinal
phase velocitya ,
cL (m/s)
Poisson’s
ratio, ν
Internal loss
factor (bending
waves),ηint
h.f c (m.Hz)
(assuming
c 0 = 343 m/s)
1900
(Typical range:
1600–2300)
0.2b
0.0125
34.1
5100
0.34
≤ 0.001
12.7
b
0.01
b
24.0
Bricks (solid) connected with mortar
(Hopkins)
1500–2000
2700
0.2
Calcium-silicate blocks (solid) connected
with thin joint compound (Schmitz et al.,
1999)
1800
2500
0.2b
0.01
25.9
760
2200
0.3b
0.01b
29.5
b
b
35.1
29.5
Chipboard (Hopkins)
Clinker concrete blocks (solid)
connected with mortar (Rindel, 1994)
1030
1720
1850
2200
0.2
0.2b
0.01
0.01b
Clinker concrete slabs (Rindel, 1994)
1725
1910
0.2b
0.01b
b
0.005
34.0
b
17.1
Concrete – cast in situ (Hopkins)
2200
3800
0.2
Dense aggregate blocks (solid)
connected with mortar (Hopkins)
2000
3200
0.2b
0.01b
20.3
Expanded clay blocks (solid)
connected with mortar (Hopkins)
800
2300
0.2b
0.007b
28.2
2500
5200
0.24
0.003–0.006
12.5
b
Glass (Hopkins)
Lightweight aggregate blocks (solid)
connected with mortar (Hopkins)
1400
2200
0.2
Medium Density Fibreboard (MDF)
(Hopkins)
760
2560
0.3b
b
29.5
0.01b
25.3
0.01
Mortar (Maysenhölder and Horvatic, 1998)
1600
2450
0.2
0.013
26.5
Oriented Strand Board (OSB)
(Hopkins)
590
2570c
0.3b
0.01b
25.2
Perspex, plexiglass (Hopkins)
1250
2350
0.3b
–
27.6
Plaster – gypsum based (Hopkins)
650
1610
0.2b
0.012b
40.3
Plasterboard – natural gypsum
(Hopkins, 1999)
860
1490
0.3b
0.0141
43.5
Plasterboard – combination of flue
gas gypsum and natural
gypsum (Hopkins, 1999)
680
1810
0.3b
0.0125
35.8
Plasterboard – gypsum with glass
fibre and other additives (Hopkins)
800
2010
0.3b
–
32.3
Plywood (Birch) (Hopkins)
710
3850
0.3b
0.016
16.8
b
Sand-cement screed (Hopkins)
2000
3250
0.2
Steel
7800
5270 (Fahy, 1985)
0.28 (Fahy, 1985)
5000
0.3b
Timber (soft wood) used for joists, studs
or battens (Hopkins)
a
b
440
0.01
20.0
≤0.0001 (Heckl, 1981)
12.3
–
Values can be used as estimates for beams or plates.
Estimate.
c
This material is usually orthotropic with values between 2200 and 3500 m/s depending on the direction. The value quoted here is cL,eff .
13.0
S o u n d
I n s u l a t i o n
Table A3 Dynamic stiffness per unit area of resilient materials measured according to ISO
9052-1
Material name
Closed-cell polyethylene
foam (Hopkins)
Expanded polystyrene
(Hopkins)
Expanded polystyrene –
pre-compressed (Hopkins)
Mineral wool – rock
(Hopkins)
Mineral wool – glass
(Hopkins)
Density (kg/m3 )
Dynamic stiffness
per unit area,
s′ (MN/m3 )
45
5
115
14
50
78
10
50
68
60
80
100
140
36
30
30
30
30
13
25
25
40
15
20
25
15
10
11
14
19
28
11
12
7
12
9
7
16
75
Rebond foam (reconstituted
open cell foam)
(Hopkins and Hall, 2006)
Nominal
uncompressed
thickness (mm)
64
96
Table A4 Dynamic stiffness of wall ties
Wall tie
Butterfly tie
(Hopkins et al., 1999)
(described in BS 1243:1978)
Double-triangle tie
(Hopkins et al., 1999)
(described in BS 1243:1978)
Vertical-twist tie
(Hopkins et al., 1999)
(described in BS 1243:1978)
Vertical-twist tie
(proprietary) (Hall et al., 2001)
Cavity width,
X (mm)
Dynamic stiffness,
sX mm (MN/m)
50
1.7
50
16.1
50
94.0
100
43.4
References
Fahy, F.J. (1985). Sound and structural vibration. Radiation, transmission and response, London:
Academic Press ISBN: 0122476700.
Hall, R. and Hopkins, C. (2001). Dynamic stiffness of wall ties used in masonry cavity walls: measurement
procedure, BRE Information Paper IP3/01, BRE, Watford, England ISBN: 1860814611.
610
Appendix
Hall, R., Hopkins, C. and Turner, P. (2001). The effect of wall ties in external cavity walls on the airborne
sound insulation of solid separating walls, Proceedings of ICA 2001, Rome, Italy.
Heckl, M. (1981). The tenth Sir Richard Fairey Memorial lecture: sound transmission in buildings, Journal
of Sound and Vibration, 77 (2), 165–189.
Hopkins, C. (1999). Building acoustics measurements, Proceedings of the Institute of Acoustics, 21 (3),
9–16.
Hopkins, C. and Hall, R. (2006). Impact sound insulation using timber platform floating floors on a concrete
floor base, Building Acoustics, 13 (4), 273–284.
Hopkins, C., Wilson, R. and Craik, R.J.M. (1999). Dynamic stiffness as an acoustic specification parameter
for wall ties used in masonry cavity walls, Applied Acoustics, 58, 51–68.
Hopkins, C. Courtesy of ODPM and BRE.
Maysenhölder, W. and Horvatic, B. (1998). Determination of elastodynamic properties of building materials,
Proceedings of Euronoise 98, Munich, Germany, 421–424.
Rindel, J.H. (1994). Dispersion and absorption of structure-borne sound in acoustically thick plates, Applied
Acoustics, 41, 97–111.
Schmitz, A., Meier, A. and Raabe, G. (1999). Inter-laboratory test of sound insulation measurements on
heavy walls. Part 1 – Preliminary test, Building Acoustics, 6 (3/4), 159–169.
611
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Standards
T
he following list of acoustic Standards relate to the measurement and prediction of sound
and vibration; these were current at the time of writing and are referenced in the text.
The reader is advised that other Standards exist and that the Standards listed below will
be updated over time.
ISO 140-1:1997 Acoustics – Measurement of sound insulation in buildings and of building
elements – Part 1: Requirements for laboratory test facilities with suppressed flanking
transmission, International Organization for Standardization.
ISO 140-2:1991 Acoustics – Measurement of sound insulation in buildings and of building
elements – Part 2: Determination, verification and application of precision data, International
Organization for Standardization.
ISO 140-3:1995 Acoustics – Measurement of sound insulation in buildings and of building
elements – Part 3: Laboratory measurements of airborne sound insulation of building elements,
International Organization for Standardization.
ISO 140-4:1998 Acoustics – Measurement of sound insulation in buildings and of building elements – Part 4: Field measurements of airborne sound insulation between rooms, International
Organization for Standardization.
ISO 140-5:1998 Acoustics – Measurement of sound insulation in buildings and of building
elements – Part 5: Field measurements of airborne sound insulation of facade elements and
facades, International Organization for Standardization.
ISO 140-6:1998 Acoustics – Measurement of sound insulation in buildings and of building
elements – Part 6: Laboratory measurements of impact sound insulation of floors, International
Organization for Standardization.
ISO 140-7:1998 Acoustics – Measurement of sound insulation in buildings and of building
elements – Part 7: Field measurements of impact sound insulation of floors, International
Organization for Standardization.
ISO 140-8:1997 Acoustics – Measurement of sound insulation in buildings and of building
elements – Part 8: Laboratory measurements of the reduction of transmitted impact noise by
floor coverings on a heavyweight standard floor, International Organization for Standardization.
ISO 140-9:1985 Acoustics – Measurement of sound insulation in buildings and of building
elements – Part 9: Laboratory measurement of room-to-room airborne sound insulation of a
suspended ceiling with a plenum above it, International Organization for Standardization.
ISO 140-10:1991 Acoustics – Measurement of sound insulation in buildings and of building
elements – Part 10: Laboratory measurement of airborne sound insulation of small building
elements, International Organization for Standardization.
613
S o u n d
I n s u l a t i o n
ISO 140-11:2005 Acoustics – Measurement of sound insulation in buildings and of building
elements – Part 11: Laboratory measurements of the reduction of transmitted impact sound by
floor coverings on lightweight reference floors, International Organization for Standardization.
ISO 140-12:2000 Acoustics – Measurement of sound insulation in buildings and of building
elements – Part 12: Laboratory measurement of room-to-room airborne and impact sound
insulation of an access floor, International Organization for Standardization.
ISO 140-14:2004 Acoustics – Measurement of sound insulation in buildings and of building
elements – Part 14: Guidelines for special situations in the field, International Organization for
Standardization.
ISO 140-16:2006 Acoustics – Measurement of sound insulation in buildings and of building
elements – Part 16: Laboratory measurement of the sound reduction index improvement by
additional lining, International Organization for Standardization.
ISO 140-11:2005 Acoustics – Measurement of sound insulation in buildings and of building
elements – Part 11: Laboratory measurements of the reduction of transmitted impact sound by
floor coverings on lightweight reference floors, International Organization for Standardization.
ISO 140-18:2006 Acoustics – Measurement of sound insulation in buildings and of building elements – Part 18: Laboratory measurement of sound generated by rainfall on building elements,
International Organization for Standardization.
ISO 717-1:1996 Acoustics – Rating of sound insulation in buildings and of building elements –
Part 1: Airborne sound insulation, International Organization for Standardization.
ISO 717-2:1996 Acoustics – Rating of sound insulation in buildings and of building elements –
Part 2: Impact sound insulation, International Organization for Standardization.
ISO 15186-1:2000 Acoustics – Measurement of sound insulation in buildings and of building
elements using sound intensity – Part 1: Laboratory measurements, International Organization
for Standardization.
ISO 15186-2:2003 Acoustics – Measurement of sound insulation in buildings and of building
elements using sound intensity – Part 2: Field measurements, International Organization for
Standardization.
ISO 15186-3:2002 Acoustics – Measurement of sound insulation in buildings and of building elements using sound intensity – Part 3: Laboratory measurements at low frequencies,
International Organization for Standardization.
ISO 9614-1:1993 Acoustics – Determination of sound power levels of noise sources using
sound intensity – Part 1: Measurement at discrete points, International Organization for
Standardization.
ISO 10848-1:2006 Acoustics – Laboratory measurement of the flanking transmission of airborne and impact sound between adjoining rooms – Part 1: Frame document, International
Organization for Standardization.
614
Standards
ISO 10848-2:2006 Acoustics – Laboratory measurement of the flanking transmission of airborne and impact sound between adjoining rooms – Part 2: Application to light elements when
the junction has a small influence, International Organization for Standardization.
ISO 10848-3:2006 Acoustics – Laboratory measurement of the flanking transmission of airborne and impact sound between adjoining rooms – Part 3: Application to light elements when
the junction has a substantial influence, International Organization for Standardization.
ISO 18233:2006 Acoustics – Application of new measurement methods in building and room
acoustics, International Organization for Standardization.
Technical specifications for measurement equipment
IEC 1043:1993 Electroacoustics – Instruments for the measurement of sound intensity –
Measurement with pairs of pressure sensing microphones, International Electrotechnical
Commission.
IEC 61260:1995 Electroacoustics – Octave-band and fractional-octave-band filters, International Electrotechnical Commission.
IEC 61672-1:2002 Electroacoustics – Sound level meters. Part 1: Specifications, International
Electrotechnical Commission.
ISO 5348:1998 Mechanical vibration and shock – Mechanical mounting of accelerometers,
International Organization for Standardization.
Reverberation time and absorption
ISO 3382:1997 Acoustics – Measurement of the reverberation time of rooms with reference to
other acoustical parameters, International Organization for Standardization.
ISO 354:2003 Acoustics – Measurement of sound absorption in a reverberation room,
International Organization for Standardization.
ISO 10534-1:1996 Acoustics – Determination of sound absorption coefficient and impedance
in impedance tubes. Part 1: Method using standing wave ratio, International Organization for
Standardization.
ISO 10534-2:1998 Acoustics – Determination of sound absorption coefficient and impedance
in impedance tubes. Part 2: Transfer-function method, International Organization for
Standardization.
Prediction
EN 12354-1:2000 Building acoustics – Estimation of acoustic performance of buildings from
the performance of elements – Part 1: Airborne sound insulation between rooms, European
Committee for Standardization.
615
S o u n d
I n s u l a t i o n
EN 12354-2:2000 Building acoustics – Estimation of acoustic performance of buildings from
the performance of elements – Part 2: Impact sound insulation between rooms, European
Committee for Standardization.
EN 12354-3:2000 Building acoustics – Estimation of acoustic performance of buildings from the
performance of elements – Part 3: Airborne sound insulation against outdoor sound, European
Committee for Standardization.
EN 12354-4:2000 Building acoustics – Estimation of acoustic performance of buildings from
the performance of elements – Part 4: Transmission of indoor sound to the outside, European
Committee for Standardization.
Material properties
ISO 9052-1:1989 Acoustics – Method for the determination of dynamic stiffness – Part 1:
Materials used under floating floors in dwellings, International Organization for Standardization.
ISO 9053:1991 Acoustics – Materials for acoustical applications – Determination of airflow
resistance, International Organization for Standardization.
ISO/PAS 16940:2004 Glass in building – Glazing and airborne sound insulation – Measurement of the mechanical impedance of laminated glass, International Organization for
Standardization.
Other Standards
ISO 1683:1983 Acoustics – Preferred reference quantities for acoustic levels, International
Organization for Standardization.
ISO 266:1997 Acoustics – Preferred frequencies, International Organization for Standardization.
ISO 1996-1:1982 Description and measurement of environmental noise. Part 1: Guide to
quantities and procedures, International Organization for Standardization.
ISO 9613-1:1993 Acoustics – Attenuation of sound during propagation outdoors. Part 1:
Calculation of the absorption of sound by the atmosphere, International Organization for
Standardization.
ISO 7626-2:1990 Method for experimental determination of mechanical mobility. Part 2:
Measurements using single-point translation excitation with an attached vibration exciter,
International Organization for Standardization.
ISO 7626-5:1994 Method for experimental determination of mechanical mobility. Part 5:
Measurement using impact excitation with an exciter which is not attached to the structure,
International Organization for Standardization.
EN 12758:2002 Glass in building – Glazing and airborne sound insulation – Product
descriptions and determination of properties, European Committee for Standardization.
616
Standards
IEC 60721-2-2:1988 Classification of environmental conditions – Part 2: Environmental
conditions appearing in nature – Precipitation and wind, International Electrotechnical
Commission.
British Standards
British Standards can be obtained from BSI Customer Services, 389 Chiswick High Road,
London W4 4AL. Tel: +44 (0)20 8996 9001. Email: cservices@bsi-global.com
617
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Index
Absorption (sound)
air, 26, 29, 60, 92, 104
area, 29, 33, 42, 61, 242, 260
coefficient, 27–29, 93–96
measurement, 40, 354–355
normal incidence sound absorption coefficient, 28,
37, 94, 354
statistical sound absorption coefficient, 28–29, 93,
94–96, 354, 355
Absorption length (vibration), 156
Accelerometers, 222–223
axis of sensitivity, 223, 391
mass loading, 223
mounting, 223, 371–372, 380, 391
Admittance
specific acoustic, 8, 28
Air
density, 1–2
gas compressibility, 5
properties of dry air, 607
Airborne sound insulation descriptors
apparent intensity sound reduction index, 261, 264
apparent sound reduction index, 259, 262, 263, 264,
266–267
element-normalized level difference, 242
intensity element-normalized level difference, 243,
244, 261, 265
intensity normalized level difference, 261
intensity sound reduction index, 242, 244, 261, 270
level difference, 242
modified intensity sound reduction index, 270
normalized flanking level difference, 377, 378
normalized level difference, 259, 268
sound reduction improvement index, 245, 246
sound reduction index (transmission loss), 240, 451
standardized level difference, 259–260, 268
standardized sound exposure level difference, 268
Airflow resistance, 80, 81, 353–354
Airflow resistivity, 80, 81, 82, 95, 360, 492
Air paths, 493–494
circular aperture, 494, 497–498
identification, 271
slit-shaped aperture, 494–497
Angular frequency, 3
Attenuation coefficient (air), 29
A-weighting, 9–10
Background noise, 268–270, 314, 317
Bandwidth
3 dB modal (half-power bandwidth), 42, 181, 367
effective, 228
ideal filter, 227
octave, 227
one-third-octave, 227
statistical (filter), 228
Bang machine, 298, 299, 300, 516
Bending stiffness
beam, 126
effective (orthotropic plate), 135
measurement, 371
plate, 128
Building elements/materials
access floors, 375–377
beam and block floors, 193, 448, 510–511, 600–603
blockwork (porous), 491–493
cast in situ concrete, 148, 152, 183, 187, 190, 215,
437–438, 446
doors, 440, 480, 483, 494, 507–508
glass (float), 157, 255, 258, 311–312, 427–429,
502–503, 608
glass (laminated), 313, 367–368, 503–504
ground floor slab, 214
insulating glass unit, 311–312, 460, 502, 504–506
masonry walls, 134, 153, 155, 159, 249–252, 271,
374–375, 429, 431–433, 435, 446–449, 476–478,
484, 491–493, 508, 563–576, 585–592
mineral wool, 79, 81, 85, 360, 361, 490–491, 522,
610
plasterboard, 134, 151, 152, 154, 194, 210, 211,
212, 213, 214, 366, 429, 440, 442, 459, 479, 609
resilient materials under floating floors, 187–188,
356–361, 516–524, 610
suspended ceilings, 375–377, 512
timber joist floors, 79, 89–91, 93, 148, 177, 293,
469–471, 512, 599
wall ties, 361–364, 467–469, 473–474, 588–590, 610
windows, 104, 263, 264–265, 271, 381, 462–467,
499, 502, 576
Cavities, 77–97, 454–480, 506–507, 512, 582–583,
588–592
Coincidence or trace-matching, 198, 255, 437
Complex exponential notation, 3–4
Consistency relationship, 410, 541
Corridors, 62–65, 583
Coupling loss factor, 26, 410
between plates and/or beams, 158–159, 473–476,
583
between plates and spaces, 420
between plates – across foundations, 476–478, 580
between plates – statistical distributions, 567–571
between spaces – non-resonant transmission,
421–422, 471–472, 493, 501, 581–582
between spaces – using the element-normalized
level difference, 582
between spaces – using the sound reduction index,
581
determined using Experimental SEA, 562–563
measurement, 400–402
Critical frequency, 197–198
effective, 198
Damping
see absorption, internal loss factor, total loss factor,
reverberation time
3dB modal bandwidth, 42, 181, 367
constant damping ratio, 561
Dd and Df, 247, 593
Decay curve, 29–31, 313
diffuse field, 30, 32
effect of detector, 318–321
effect of filters, 321–327
ensemble average, 314
619
I n d e x
Decay curve (continued)
evaluation of, 327
evaluation range, 30
measurement method – integrated impulse
response, 314–317
measurement method – interrupted noise, 313–314
non-diffuse field, 34–41, 156
Decibels – reference values
energy, 8, 113
impedance, 113
loss factors, 8, 113
mobility, 113
sound, 8
vibration, 113
Delany and Bazley equations, 85
Diffuse field, 11–12, 140, 167
Dirac delta function, 10, 226, 314, 315, 335
Dynamic stiffness, 355–356
resilient materials, 356–361, 458, 522, 610
wall ties, 361–364, 474, 610
Eigenfrequency, 19
Eigenfunction, 53
Eigenvalue, 19
Energy, 75–76, 217, 409–410
Energy density, 59–62, 76–77, 171
Energy doubling, 100, 267
Equipartition of modal energy, 410–411
Equivalent continuous level, 230
Experimental SEA – see Statistical Energy Analysis
Façades
effect of diffraction, 101–102
energy doubling, 100, 105, 267
pressure doubling, 52, 99, 100, 262
surface sound pressure level, 102, 104, 262, 266,
267
Facade sound insulation, 262–268
aircraft noise, 266–268
loudspeaker method, 262–265, 425
propagation paths, 97–98
railway noise, 266–268
road traffic noise, 266, 267–268
Fd and Ff, 247, 593
Fibrous materials, 80–82, 84, 85, 353, 459, 490–491
Filters, 227–230, 321–327
Finite Element Methods (FEM), 141, 160, 560–563
Finite plates
non-resonant transmission, 424–427
sound radiation, 202–213
Floating floors, 149, 177, 187, 188, 194, 195, 273–275,
303, 516–525
Footsteps, 272–273, 285–300
Foundations, 467–469, 476–478, 578–580, 591
Frequency range, low, mid and high – definition, viii
Gas
bulk compression modulus, 83, 360, 457
compressibility, 5, 78, 83
properties, 607
Group velocity
bending wave (thin beam), 128
bending wave (thick plate), 134–135
bending wave (thin plate), 129
quasi-longitudinal wave (beam), 117
quasi-longitudinal wave (plate), 117
sound in air, 22
torsional wave, 120
transverse shear wave, 122
620
Heavyweight constructions, viii
Helmholtz resonator, 460–461
Image sources, 15–17, 49–50, 59, 97–98
Impact sound insulation descriptors
improvement of impact sound insulation, 273–275
normalized flanking impact sound pressure level,
377, 378
normalized impact sound pressure level, 273
standardized impact sound pressure level, 275
Impedance (sound)
acoustic, 7
characteristic, 5
normal acoustic surface, 8
specific acoustic, 5, 8, 28
Impedance (vibration)
driving-point, 178, 185–197
Impulse response, 10–11
Infinite beams
driving-point impedance, 186, 189
Infinite plates
airborne sound insulation (isotropic), 435–440
airborne sound insulation (orthotropic), 443–445
driving-point impedance, 185, 189
non-resonant transmission, 422–424
sound radiation, 198–202
Intensity measurement
sound intensity, 342–353
structural intensity, 384–396
Internal loss factor, 26, 42, 92–94, 157–158
absorptive surfaces, 26
air absorption, 26
air pumping, 158
edge damping, 157
frequency-dependence, 157, 366
measurement, 365–367
temperature-dependence, 157, 367–368
ISO rubber ball, 274–275, 298–300, 515–516, 525
ISO tapping machine, 272–275, 275–298, 301–305,
509–526
Junction – beams
rigid (bending waves only), 543
Junction – plates
angled plates, 556
at a beam/column, 559
beam, 539, 542, 544, 545
foundations (bending waves only), 578–580
hinged, 559
resilient (bending and in-plane waves), 557–559
rigid (bending and in-plane waves), 543–550
rigid (bending waves only), 541–543
k-space, 20
Lamé constants, 434
Leaks – see Air paths
Lightweight cavity walls, 475, 478
Lightweight constructions, viii
Lightweight floors, 274, 300, 512, 515, 524–525
Low-frequency airborne sound insulation, 451–454
Mass law
diffuse incidence, 424
empirical, 507–509
field incidence, 424, 426–427
normal incidence, 424
single angle of incidence, 424
Mass-spring systems, 180–181, 278–280, 289,
355–359, 361–364, 457–461, 479, 523
Index
Material properties
indicative values, 607–611
measurement, 353–375
Maximum Length Sequence (MLS), 333–342
Maximum time-weighted level, 232
Mean free path, 12–15, 92, 140–141
Membranes, 454
Microphones, 221–222
Mobility (driving-point), 179
closely connected plates, 194
effect of static load, 188
envelope of peaks, 188
measurement, 371–373
plates formed from beams, 191–193
plates with attached beams, 194–197
plates with complex cross-sections, 189–190
ribbed plate, 190–191
single degree-of-freedom system, 180–181
Mobility (transfer), 181
Modal density
beams, 145–148
cavities, 87–89
measurement, 371–373
plates, 150–152
rooms, 19–22
Modal overlap factor, 42–44
geometric mean, 413
Modes
axial, tangential, oblique, 19
bending (beam), 142–144
bending (plate), 149
count, 22–23
degenerate, 187
equivalent angle, 24, 89, 152–154
excitation of, 56–57, 181–182, 182–184
global, 182–184
local, 17–19, 141–145
shape (plate), 160
shape (room), 53
spacing, 23
torsional (beam), 144–145
transverse shear (plate), 149–150
quasi-longitudinal (beam), 145
quasi-longitudinal (plate), 150
Multi-modal subsystems, 413
Nearfield
radiation, 214–217
vibration, 128, 161–167
Niche effect, 253–258
Non-resonant transmission, 419
air paths, 493–502
finite plate, 424–427
infinite plate, 422–424
mass law, 421–427
porous materials, 488–493
Openings (e.g. windows) in walls
coupling loss factor, 572–576
modes (bending), 163
structural intensity, 394–395
Orthotropic plates, 135–140, 442–451
Oscillators (SEA), 409–410
Particle velocity, 4–7, 197, 202, 343
Phase velocity
bending wave (thin beam), 128
bending wave (thick plate), 135
bending wave (thin plate), 129
quasi-longitudinal wave (beam), 117
quasi-longitudinal wave (plate), 117
sound in air, 2, 22
sound in gases, 78
sound in porous material, 84
torsional wave, 120
transverse shear wave, 122
Pink noise, 226
Piston (sound radiation), 197
Plane waves (sound), 4–6
Plates
closely connected, 158, 194, 440–442
corrugated/profiled, 136–139, 449–451
laminated, 141, 157, 212, 503–504
orthotropic, 135–140, 442–451
porous, 484, 491–493
ribbed, 139–140, 190–191
spatially periodic, 174–177
thick, 117–118, 132, 133–135, 433–435, 551–552
thin, 114–117, 121–122, 123–133
Plate-cavity-plate systems, 454–480
mass-spring-mass resonance, 457–460
structural coupling, 473–478
Point excitation (structural), 168–172
Point source (sound), 6
Porosity, 79–80
Power (sound)
measurement, 346
point source, 7
radiated by nearfield, 214–217
Power (structural)
absorbed by plate boundaries, 156
incident upon junction (beam), 158
incident upon junction (plate), 159
input (point force), 179
Pressure doubling, 52, 99–100
Radiation efficiency (plates), 197
corner, edge and surface radiators, 205–208
frequency-average, 209
individual modes, 202–208
into porous material, 213–214
measurement, 373–375
method No. 1–4, 209–213
plate connected to a frame, 213
Rain noise
measurement, 305–313
power input, 305–310
radiated sound, 311–313, 526–527
raindrop diameters, 307–310
Rain-on-the-roof excitation, 411
Ratio of specific heats, 78
Reflection
coefficient (sound), 27
coefficient (vibration), 164–165, 550, 553
diffuse, 11–12, 140
specular, 11, 15
Resilient materials
air stiffness (porous materials), 360
at plate junctions, 249, 446–447, 557–559
dynamic stiffness, 356–361
in floating floors, 516–525
Reverberation distance, 62, 172
Reverberation time, 29–31
see decay curve
cavities, 92
evaluation range, 30
Eyring’s equation, 33
621
I n d e x
Reverberation time (continued)
measurement, 313–333
rooms, 29–31
Sabine’s equation, 33
structural, 154–156
Rooms, 1–77
furnished, 260, 324
typical, viii
Root-mean-square (rms), 6
Sandwich panel, 480–482
dilatational resonance, 481
Schroeder cut-off frequency, 43
SEA-based model, 592–603
Snell’s law, 541
Soft floor coverings, 273–275, 280, 285–286, 301–302,
303, 513–516
Soil stiffness, 581
Sound exposure level, 232
Sound intensity, 5, 7
dynamic capability index, 351
error analysis, 348–353
measurement, 342–353
p-p probe, 344–345
phase mismatch, 348–352
pressure-residual intensity index, 351
residual intensity, 350
Sound power, 346–348
from intensity measurements, 380–384
from vibration measurements, 380
ranking, 380
Sound pressure level
definition, 8
spatial variation, 44–75
Spatial averaging, 234–238
Spatial correlation coefficient, 235–236
Spatial sampling
accelerometer positions, 400
continuously moving microphone, 237–238
microphone attached to façade surface, 102,
104
stationary microphone positions, 235–237
uncorrelated samples, 235–238
Specific acoustic admittance, 8, 28
Specific airflow resistance, 80
Speed of sound – see Phase velocity
Standard deviation and variance
filtered Gaussian white noise, 232
plate vibration (spatial variation), 172–173
reverberation time (diffuse sound fields), 329–333
sound pressure (spatial variation), 71–75
temporal and spatial averaging, 234
temporal variation, 232–234
velocity level difference (spatial variation), 399
Static pressure, 78
Statistical Energy Analysis (SEA), 409–418
consistency relationship, 410
coupled oscillators, 409
equipartition of modal energy,
410–411
experimental SEA (ESEA), 562–563
matrix solution, 414–415
path analysis, 416–418
power balance equations, 414
statistically independent excitation forces, 410–411
subsystem definition, 411–412
thermal analogy, 410
weak coupling, 411
622
Structural coupling
around the perimeter, 478
foundations, 467–469, 476–478, 578–580
line connections, 475–476
plate-cavity-plate systems, 473–478
point connections, 473–475
Structural intensity, 384–388
a-a probe, 388–389
error analysis, 390–391
measurement, 384–396
Surface finish
bonded, 483–484
linings, 245–246, 485–488
Surfaces
locally reacting, 27, 38–41
surfaces of extended reaction, 38, 40
Thickness resonances, 433–435
Thin plate limit
bending waves, 134, 246, 433–435, 551–552
quasi-longitudinal waves, 117–118
Time constant
Fast, 232
Impulse, 232
Slow, 232
Time-weighted level, 231
Total loss factor, 26, 42, 159
with transmission suite measurements, 249–253
estimates for rigidly connected masonry/concrete
plates, 159
Trace matching – see Coincidence
Transmission coefficients
apertures, 494, 497, 499–502
plates, 421, 423–424, 425, 437, 443
sound, 240, 258–259
vibration, 158, 538, 540, 541, 550
Transmission suites, 239, 246–258, 274
Universal gas constant, 78
Variance – see Standard deviation
Velocity level difference, 378–379, 396–400
accelerometer positions, 400
direction-averaged, 402
excitation positions, 400
Vibration level
definitions, 113
spatial variation, 160–178
Vibration reduction index, 402, 596
Vibration transmission
Finite Element Methods, 560–578
impedance models, 473–476
wave approach, 538–560, 578–580
Volume velocity, 7
Waterhouse correction, 76–77, 241, 270, 381–382
Wavelength, 2–3, 86, 114
Wavenumber, 3, 84, 117, 120, 122, 127, 129
Waves
bending, 123–135
dilatational, 434
dispersive, 128
plane, 4–6
quasi-longitudinal, 114–118
spherical, 6–7
torsional, 118–120
transverse shear, 121–122
White noise, 225–226