Energy Conversion and Management 46 (2005) 1703–1713 www.elsevier.com/locate/enconman Simulation studies on R134a—DMAC based half effect absorption cold storage systems S. Arivazhagan a, S.N. Murugesan b, R. Saravanan a,* , S. Renganarayanan a a b Institute for Energy Studies, Anna University, Chennai 600 025, India Department of Mechanical Engineering, Crescent Engineering College, Chennai 600 048, India Received 31 May 2004; accepted 3 October 2004 Available online 21 November 2004 Abstract This paper presents simulation studies conducted on a half effect vapour absorption cycle using R134aDMAC as the refrigerant-absorbent pair with low temperature heat sources for cold storage applications. The intermediate pressure of the cycle has been optimized for maximum COP. The effects of the temperatures of the evaporator, condenser, absorber and generator on the COP of the cycle have also been studied. It is found that the effect of the temperature of the low absorber on the performance is more pronounced than that of the high absorber. The COP for the baseline system is found to vary from 0.35 for low evaporating and high condensing temperatures to 0.46 for high evaporating and low condensing temperatures. The use of a condensate pre-cooler has resulted in an improvement of 5–15% in COP. The performance of this working fluid pair is better than that of ammonia–water for low heat source temperatures in the half effect configuration. Ó 2004 Elsevier Ltd. All rights reserved. Keywords: Half effect absorption systems; Low temperature heat sources; R134a-DMAC; Cold storage * Corresponding author. Tel.: +91 44 222 03268; fax: +91 44 222 03269. E-mail address: rsaravanan@annauniv.edu (R. Saravanan). 0196-8904/$ - see front matter Ó 2004 Elsevier Ltd. All rights reserved. doi:10.1016/j.enconman.2004.10.006 1704 S. Arivazhagan et al. / Energy Conversion and Management 46 (2005) 1703–1713 Nomenclature h m_ Q_ t T x specific enthalpy (kJ/kg) mass flow rate (kg/s) heat transfer rate (kW) temperature (°C) temperature (K) weight fraction of R134a in solution Subscripts a actual c carnot CO condenser EV evaporator LAB low absorber LGR low generator HAB high absorber HGR high generator r refrigerant ss strong solution ws weak solution 1–19 state points in the system with reference to Fig. 1. Abbreviations COP coefficient of performance CPC condensate precooler CR circulation ratio DMAC N,N-Dimethyl acetamide R134a 1,1,1,2-tetrafluoroethane SHX solution heat exchanger Greek symbols n effectiveness second law efficiency gII 1. Introduction Cold storages intended for increasing the shelf life of vegetables, fruits and food grains need to be maintained at temperatures in the range of
2 °C to 10 °C. As the refrigeration systems for this application are operated for long hours, vapour absorption systems are most economical. A single stage vapour absorption refrigeration cycle using conventional working fluids, ammonia water and water–lithium bromide, requires heat input at a minimum temperature of 100 °C in order S. Arivazhagan et al. / Energy Conversion and Management 46 (2005) 1703–1713 1705 to have a reasonable COP for a cooling temperature of around 0 °C [1]. It has been shown that the half effect cycle is feasible with heat input temperatures below that of a single stage cycle [2]. Theoretical analysis of a half effect vapour absorption cycle using water–lithium bromide reported by Ma and Deng [3] also supports this fact, but even in the half effect cycle, conventional working fluids require an input at 80 °C. Therefore, low temperature heat sources such as solar energy can not be considered as the source. In addition, conventional pairs are facing problems such as corrosion, crystallisation and vacuum operation in water–lithium bromide and rectification and incompatibility with copper in ammonia–water systems. To overcome the drawbacks of conventional working pairs, many researchers have studied the suitability of several refrigerant absorbent combinations [4–9]. Experimental investigations conducted by Muthu [10] show that a R134a-DMAC based single stage system can be operated with relatively low temperature heat sources compared to systems using conventional working fluids. In this paper, thermodynamic analysis has been conducted to obtain the variation in the performance parameters at various operating conditions for a half effect cycle using R134a-DMAC to study the feasibility of using low temperature heat sources such as solar energy. 2. Description of the system A half effect cycle takes heat input at two pressure levels. The condenser and evaporator operate at high and low pressures. The vapour generated at an intermediate pressure is fed to a second absorber, which feeds the generator at high pressure. Both generators can be supplied with heat at the same temperature. Fig. 1(a) represents the schematic arrangement of the half effect vapour absorption refrigeration system. The system consists of two solution circuits, each consisting of an absorber, generator, solution pump, solution heat exchanger and an expansion device. The low absorber is at evaporator pressure pe. The low generator and high absorber are at intermediate pressure pi. The high generator is at condenser pressure pc. The order of these pressures is pc > pi > pe ð1Þ The refrigerant vapour from the evaporator is absorbed by the weak solution in the low absorber. The strong solution from the low absorber is pumped to the low generator through the low solution heat exchanger. The weak solution in the low generator is returned to the low absorber through the low solution heat exchanger. The refrigerant vapour from the low generator is absorbed by the weak solution in the high absorber. The strong solution from the high absorber is pumped to the high generator through the high solution heat exchanger. The weak solution in the high generator is returned to the high absorber through the high solution heat exchanger. In both heat exchangers, the strong solution from the absorber is heated by the weak solution from the generator. The refrigerant is boiled out of the solution in the high generator and circulated to the condenser. The liquid refrigerant from the condenser is returned to the evaporator through an expansion valve. Both generators are supplied with heat at the same temperature. Fig. 1(b) shows the half effect vapour absorption refrigeration cycle on a log p-T diagram. The cycle 1–3–4–6 represents solution circuit 1, and the cycle 8–10–11–13 represents solution circuit 2. 1706 S. Arivazhagan et al. / Energy Conversion and Management 46 (2005) 1703–1713 14 HIGH GENERATOR CONDENCER 10 11 SOLUTION CIRCUIT 2 15 9 12 PRESSURE REDUCING VALVE SOLUTION PUMP 13 16 8 HIGH ABSORBER EXPANSION DEVICE 7 LOW GENERATOR 3 4 17 SOLUTION CIRCUIT 1 5 2 PRESSURE REDUCING VALVE SOLUTION PUMP 1 18 6 19 EVAPORATOR LOW ABSORBER (a) (b) Fig. 1. Half effect vapour absorption cycle (a) schematic (b) on log p-T coordinates. 3. Simulation study Simulation was performed to evaluate the half effect vapour absorption cycle with the following assumptions: 1. The flow through all the components is under steady state. 2. The pressure drop due to friction within the cycle can be neglected, except through the expansion valve. 3. The fluid streams in the piping between the components and the heat exchangers are adiabatic. 4. The strong solution at the outlet of the absorbers and the weak solution at the outlet of the generators are saturated. 5. The effectiveness of the solution heat exchangers is 0.7. 6. Refrigerant vapour alone leaves the generators, that is, it does not contain any traces of the absorbent. 7. Pump work is negligible. S. Arivazhagan et al. / Energy Conversion and Management 46 (2005) 1703–1713 1707 The properties of the pure refrigerant and the refrigerant-absorbent mixture were evaluated using correlations proposed by various researchers [4,9,11]. For a given set of generator, evaporator, condenser, low absorber temperature and high absorber temperature, the intermediate pressure was optimized to get maximum COP. Equations based on mass, energy and concentration balance were used in the performance analysis. Equations of mass balance are the following: m_ 1 ¼ m_ 2 ¼ m_ 3 ð2Þ m_ 4 ¼ m_ 5 ¼ m_ 6 ð3Þ m_ 8 ¼ m_ 9 ¼ m_ 10 ð4Þ m_ 11 ¼ m_ 12 ¼ m_ 13 ð5Þ m_ 14 ¼ m_ 15 ¼ m_ 16 ¼ m_ 17 ¼ m_ 18 ¼ m_ 19 ð6Þ m_ 1 x1 ¼ m_ 6 x6 þ m_ 19 ð7Þ m_ 3 x3 ¼ m_ 4 x4 þ m_ 7 ð8Þ m_ 8 x8 ¼ m_ 12 x12 þ m_ 7 ð9Þ m_ 10 x10 ¼ m_ 11 x11 þ m_ 14 ð10Þ Equations for the energy balance are the following: Q_ EV ¼ m_ 17 ðh18
h17 Þ ð11Þ Q_ CO ¼ m_ 14 ðh14
h15 Þ ð12Þ Q_ LAB ¼ m_ 6 h6 þ m_ 19 h19
m_ 1 h1 ð13Þ Q_ LGR ¼ m_ 7 h7 þ m_ 4 h4
m_ 3 h3 ð14Þ Q_ HAB ¼ m_ 13 h13 þ m_ 7 h7
m_ 8 h8 ð15Þ Q_ HGR ¼ m_ 11 h11 þ m_ 14 h14
m_ 10 h10 ð16Þ The circulation ratio (CR) is defined as the ratio of the mass flow rate of the strong solution to the mass flow rate of the refrigerant. The circulation ratio of the low pressure stage can be obtained using the refrigerant mass balance for the low absorber. The circulation ratio was obtained from the equation   1
x4 ð17Þ CRL ¼ x1
x4 1708 S. Arivazhagan et al. / Energy Conversion and Management 46 (2005) 1703–1713 Similarly, the circulation ratio for the high pressure stage is given by   1
x11 CRH ¼ x8
x11 The coefficient of performance (COP) of the cycle is   QEV COP ¼ QLGR þ QHGR The second law efficiency is   COPa gII ¼ COPc ð18Þ ð19Þ ð20Þ 4. Results and discussion The performance of the half effect vapour absorption refrigeration cycle using R134a and DMAC as the working fluid pair was evaluated for evaporator temperatures varying from
5 °C to 10 °C in steps of 5 °C. The absorber outlet temperatures were varied from 25 °C to 40 °C in steps of 5 °C. The condenser temperatures were varied from 20 °C to 30 °C in steps of 5 °C. The generator temperatures were varied from 50 °C to 70 °C in steps of 10 °C. Fig. 2 shows the variation in circulation ratio of the high and low pressure stages for different high and low absorber temperatures keeping the temperatures of all other components constant. For a given high absorber temperature, the circulation ratio is high for the lower evaporator temperature as the degassing width decreases when the evaporator temperature is lowered. For a constant evaporating temperature when the high absorber temperature is increased, the circulation ratio of both stages increases, but above a certain high absorber temperature, the increase in high circulation ratio becomes higher for solution circuit 2 than that of solution circuit 1. For a given low absorber temperature, the circulation ratio of solution circuit 1 is higher than that of solution circuit 2. Also, the increase in circulation ratio for a given increase in low absorber temperature is higher for solution circuit 1 than that for solution circuit 2. The reason for this is attributed to the relative p-t-x variation of the working fluid at high and low pressures. It can be concluded that from the circulation point of view, lowering the low absorber temperature is preferred relative to lowering the high absorber temperature. Fig. 3 shows the variation in COP with both low and high absorber temperatures at different evaporator temperatures. The COP has been computed by keeping the temperature of the high absorber constant and varying the temperature of the low absorber and vice versa for the same generator, condenser and evaporator temperatures. It is observed that for given low and high absorber temperatures, the higher is the evaporator temperature, the higher is the COP obtained, as the temperature limit between which the heat is pumped is less at higher evaporator temperatures. For a given condenser, evaporator, generator and low absorber temperatures, an increase in the low absorber temperature results in a lower degassing width of the low pressure stage, which, in turn, results in a lower COP. This decrease is negligible at higher evaporating temperatures, but at 1709 S. Arivazhagan et al. / Energy Conversion and Management 46 (2005) 1703–1713 7 ξ cpc = 0.8; t c = 30oC; t g = 70 oC; t la = 30 oC 6 High Absorber o te = - 5 C Circulation Ratio 5 te = 0 oC 4 3 te = 5 oC 2 1 Solution Circuit 1 Solution Circuit 2 0 10 Low Absorber 9 ξcpc = 0.8; tc = 30ºC; tg = 70ºC; tha=30ºC 8 te=-5ºC Circulation Ratio 7 6 5 te=0ºC 4 te=5ºC 3 2 1 Solution Circuit 1 Solution Circuit 2 0 20 25 30 35 Absorber Temperature (ºC) 40 45 Fig. 2. Variation in circulation ratio with absorber temperature at different evaporating temperatures. lower evaporating temperatures, the decrease is appreciable. The reason is that at low evaporating temperature, the degassing width of the LP stage is more sensitive to the absorber temperatures. The same trend is obtained by keeping the temperature of the low absorber constant and varying the temperature of the high absorber, but the decrease is less when compared to the former condition. 4.1. Comparison with NH3–H2O as working pair Simulation has been conducted with and without the condensate pre-cooler. Fig. 4 shows the effect of the condensate pre-cooler on the performance of the cycle. Use of a condensate pre-cooler with an effectiveness of 0.8 is found to give an improvement of about 13.5% in COP for a R134a-DMAC cycle for different absorber temperatures when the temperatures of all the 1710 S. Arivazhagan et al. / Energy Conversion and Management 46 (2005) 1703–1713 0.5 0.48 ξcpc = 0.8; tc = 30ºC; tg = 70ºC te = 5ºC 0.46 te = 0ºC 0.44 COP 0.42 te = -5ºC 0.4 0.38 0.36 0.34 tla varied at tha = 30ºC 0.32 tha varied at tla = 30ºC 0.3 20 25 30 35 Absorber Temperature (ºC) 40 45 Fig. 3. Variation in COP with respect to absorber temperature at different evaporating temperatures. 0.5 R134a-DMAC With CPC 0.45 COP 0.4 R134a-DMAC without CPC NH3-H2O With CPC 0.35 0.3 NH3-H2O Without CPC 0.25 tla varied at tha = 30ºC tha varied at tla = 30ºC ξcpc = 0.8; tc = 30ºC; tg = 70ºC 0.2 20 25 30 35 Absorber Temperature, (ºC) 40 45 Fig. 4. Effect of condensate pre-cooler on COP. other components are kept constant. For the ammonia–water cycle, the improvement in COP is only 5.6% under the same conditions. The COP of the R134a–DMAC cycle is 25% higher than that of the ammonia–water cycle without the condensate pre-cooler and 35% higher with the condensate pre-cooler. Further, half effect ammonia–water cycles are not feasible with low absorber temperatures above 30 °C. 1711 S. Arivazhagan et al. / Energy Conversion and Management 46 (2005) 1703–1713 0.5 ξcpc = 0.8; te = 0ºC; tla = 30ºC; tg = 70ºC 0.48 0.46 0.44 COP R134a-DMAC tha=25ºC 0.42 tha=30ºC 0.4 tha=35ºC 0.38 tha=40ºC 0.36 tha=25-40ºC 0.34 0.32 NH3-H2O 0.3 15 20 25 Condenser Temperature (ºC) 30 35 Fig. 5. Variation in COP with condenser temperature at different high absorber temperature. Fig. 5 shows the effect of condenser temperature on the performance of the cycle for both R134a–DMAC and ammonia–water at different high absorber temperatures. As the condenser temperature increases, the COP decreases as it increases the overall temperature of heat rejection. This decrease per degree is high at higher values of the high absorber temperatures. As discussed earlier, for a given condenser temperature, the higher is the absorber temperature, the lower is the COP obtained. Fig. 6 shows the effect of generator temperature on the performance of the cycle for both R134a–DMAC and ammonia–water at different high absorber temperatures. At low generator temperature, the absorber temperature is found to be more significant, and its effect becomes 0. 5 tla =30 oC, t c =30 oC, t e =0 o C, ξ CPC = 0. 8 R134aDMAC 0.45 tha =25 oC tha =30 oC 0. 4 COP tha =35 oC tha =40 oC 0.35 tha = 25 - 40 o C NH3 -H2 O 0. 3 0.25 50 60 70 80 90 100 o Generator Temperature ( C) Fig. 6. Variation in COP with generator temperature at different high absorber temperature. 1712 S. Arivazhagan et al. / Energy Conversion and Management 46 (2005) 1703–1713 0.5 0.45 Second Law efficiency 0.4 te=-5ºC 0.35 te=0ºC te=5ºC 0.3 0.25 0.2 0.15 0.1 0.05 NH3 –H2O tha = 30ºC; tla = 30ºC; tc = 0ºC; ξcpc = 0.8 R134a-DMAC 0 60 70 80 90 Generator Temperature (ºC) Fig. 7. Variation in second law efficiency with generator temperature at different evaporator temperature. Table 1 Comparison of R134a-DMAC and NH3–H2O in half effect cycles Parameter R134a-DMAC NH3-H2O COP Second law efficiency Sink temperature Generator temperature 0.35–0.46 28–44% Less than 30 °C 60–90 °C 0.30–0.34 24–30% Less than 30 °C 70–90 °C negligible at high temperature. It can be concluded that the ammonia–water system cannot be operated with source temperatures below 70 °C under these operating conditions. Fig. 7 shows the effect of generator temperature on the second law efficiency of the cycle for both R134a–DMAC and ammonia–water at different evaporator temperatures. As the generator temperature increases, the second law efficiency decreases because of the lower usage of the higher availability at high temperatures. At the higher evaporating temperature, also the second law efficiency is lower. The comparison of the performance of R134a-DMAC and ammonia–water for half effect systems for all the operating conditions is shown in Table 1. It is found that the R134a-DMAC pair is better for half effect systems from the viewpoints of COP, second law efficiency and source temperature for solar energy based cold storage systems. 5. Conclusions Simulation studies on a half effect vapour absorption refrigeration cycle for solar energy based cold storage systems using R134a-DMAC as working fluids have been conducted. The COP of this cycle is found to be about 0.35–0.46 for an evaporating temperature of
5 to 5 °C with a heat input at 70 °C with a condensing temperature of 20–30 °C and absorber temperatures at 25 °C. Improvement in COP up to 13% has been found with the use of a condensate pre-cooler. When S. Arivazhagan et al. / Energy Conversion and Management 46 (2005) 1703–1713 1713 compared to ammonia–water, R134a-DMAC gives a marginally higher COP in the half effect cycle at low heat source temperatures. An average increase in COP of 33% can be obtained within the evaporator temperature range of
5 °C to 5 °C for a typical operating condition. From these, it is evident that the R134a-DMAC refrigerant absorbent combination may be considered as one of the most favorable working fluids when a half effect system is to be operated with low temperature heat sources. Acknowledgement The authors wish to thank DST-SERC, Government of India, for providing financial assistance to conduct the project. References [1] Alefield, Heat conversion systems, CRC Press. [2] Keith Herold, Readermacher, Absorption chillers and heat pump, CRC Press. [3] Ma WB, Deng SM. Theoretical analysis of low temperature hot source driven two stage LiBr–H2O absorption refrigeration system. Int J Refrig 1996;19(2):141–6. [4] Borde I, Jelinek M, Daltrophe NC. Refrigerant-absorbent mixtures based on the refrigerant R134A. In: Proc XVIII Int Conf Refrig, Montreal, Canada, 1991;653–8. [5] Borde I, Jelinek M, Daltrophe NC. Absorption system based on the refrigerant R134a. Int J Refrig 1995;18(6): 387–94. [6] Borde I, Jelinek M, Daltrophe NC. Working substance for absorption heat pumps based on R32. Proc nineteenth int congress refrig 1995:81–7. [7] Jelinek M, Borde I. Working fluids for absorption heat pumps based on R125 (pentafluroethane) and organic absorbents. In: Proc int sorption Heat pump conf, 1999; pp. 205–8. [8] Vikas Patnaik, Kenneth J. Schultz, Peter Xla. A thermodynamic investigation of the R22-DMF, R134a-DMETEG and NH3–H2O working pairs for enhanced vapour compression via a solution circuit in air conditioning applications, In: Proc int Sorption Heat Pump Conf, 2002; pp. 221–6. [9] Borde I, Jelinek M, Daltrophe NC. Working fluids for an absorption system based on R124–Chloro–1, 1, 1, 2–tetrafluoroethane and organic absorbents. Int J Refrig 1997;20(4):256–66. [10] Muthu V. (2003), Studies on vapour absorption refrigeration system using R134a-DMAC, a Ph.D. thesis, Department of Mechanical Engineering, Anna University, Chennai. [11] Clelend AC. Polynomial curve–fits for refrigerant thermodynamic properties: Extension to include R134a. Rev Int Froid 1994;17(4):245–9.