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Performance Evaluation of R134a/DMF-Based VaporAbsorption Refrigeration SystemSuresh Mariappan a & Mani Annamalai aa Department of Mechanical Engineering, Indian Institute of Technology Madras, Chennai,IndiaAccepted author version posted online: 13 Dec 2012.Published online: 13 Mar 2013.

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Heat Transfer Engineering, 34(11–12):976–984, 2013Copyright C©© Taylor and Francis Group, LLCISSN: 0145-7632 print / 1521-0537 onlineDOI: 10.1080/01457632.2012.753577

Performance Evaluationof R134a/DMF-Based VaporAbsorption Refrigeration System

SURESH MARIAPPAN and MANI ANNAMALAIDepartment of Mechanical Engineering, Indian Institute of Technology Madras, Chennai, India

This article describes an experimental investigation to measure performances of a vapor absorption refrigeration systemof 1 ton of refrigeration capacity employing tetrafluoro ethane (R134a)/dimethyl formamide (DMF). Plate heat exchangersare used as system components for evaporator, condenser, absorber, generator, and solution heat exchanger. The bubbleabsorption principle is employed in the absorber. Hot water is used as a heat source to supply heat to the generator.Effects of operating parameters such as generator, condenser, and evaporator temperatures on system performance areinvestigated. System performance was compared with theoretically simulated performance. It was found that circulationratio is lower at high generator and evaporator temperatures, whereas it is higher at higher condenser temperatures. Thecoefficient of performance is higher at high generator and evaporator temperatures, whereas it is lower at higher condensertemperatures. Experimental results indicate that with addition of a rectifier as well as improvement of vapor separation inthe generator storage tank, the R134a/DMF-based vapor absorption refrigeration system with plate heat exchangers couldbe very competitive for applications ranging from –10◦C to 10◦C, with heat source temperature in the range of 80◦C to 90◦Cand with cooling water as coolant for the absorber and condenser in a temperature range of 20◦C to 35◦C.

INTRODUCTION

The search for new working fluids to replace the well-knownworking pairs ammonia–water and lithium bromide–waterfor vapor absorption refrigeration systems (VARS), in or-der to overcome some of their limitations, is continuing.Many environmentally friendly fluid combinations have beensuggested by a number of investigators. Of these, difluo-romonochloromethane (R22)-organic solvent based absorptionrefrigeration systems have been extensively studied by Fa-touh and Srinivasa Murthy [1–4], Karthikeyan et al. [5, 6],and Sujatha et al. [7–9]. They observed that, in addition tohaving certain advantages over ammonia–water and lithiumbromide–water systems, R22-based VARS can operate usinglow-temperature heat sources in the range of about 75–95◦C.However, along with chlorofluorocarbons (CFCs), hydrochlo-rofluorocarbons (HCFCs) are also covered by Montreal andKyoto International Protocols and are being phased out. Hy-

Address correspondence to Professor Mani Annamalai, Refrigeration &Airconditioning Laboratory, Department of Mechanical Engineering, Indian In-stitute of Technology Madras, Chennai 600036, India. E-mail: mania@iitm.ac.in

drofluorocarbons (HFCs) are looked upon as alternate refriger-ants to CFCs and HCFCs. They are not destructive to the ozonelayer, though they have a slight effect on global warming. R134ais the commonly used HFC refrigerant. The Kyoto Protocol dur-ing 1997 has set R134a as one of the greenhouse gases. Howeverthere is no phase-out date for this refrigerant and it is expectedto be highly used in refrigeration industries. Thus, R134a-basedVARS are being investigated. Borde et al. [10, 11] investi-gated the possibility of using R134a as a refrigerant in com-bination with different organic absorbents and concluded thatoverall performance of R134a–tetraethylene glycol dimethylether (DMETEG) was better than that of R134a–N-methyl ∈-caprolactam (MCL) or R134a–dimethyl ethyleneurea (DMEU).Though there was a reduction in coefficient of performance(COP) and an increase in circulation ratio with R134a systemscompared to R22 systems, the R134a–DMETEG combinationcan be used as a replacement for R22-based working fluids in ab-sorption machines since R22 is being phased out. Songara et al.[12] carried out a comparative thermodynamic study of VARSworking with R134a and R22. They observed that the R22-basedsystem yields a significantly better COP than the R134a system.However, since the R134a system operates at much lower pres-sures than the R22 system, the possibility exists to improve its

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COP by resorting to two-stage operation. Songara et al. [13] alsoanalyzed thermodynamic studies on double effect VARS work-ing with R134a as refrigerant and dimethyl acetamide (DMA)as absorbent. COP yielded from these systems is comparableto single-stage R22–DMA systems, but at slightly higher heatsource temperatures. They suggested cascaded systems for sub-zero temperatures. Nezu et al. [14] examined the possibility oftesting R134a as a refrigerant in VARS with various organic sol-vents and showed that the R134a–DMA and the R134a–DMFsystems are found attractive as the working-fluid pairs for theabsorption refrigeration system than other R-134a–absorbentsystems. Yokozeki [15] studied the theoretical performance ofvarious refrigerant-absorbent pairs in a VARS cycle employingthe equations of state. R134a–DMF and R134a–DMA systemsexhibit better performance, compared to other R134a–absorbentsystems. The circulation ratio is less and COP is more forthe R134a–DMF system compared to the R134a–DMA sys-tem. Arivazhagan et al. [16] conducted simulation studies ona half effect vapor absorption cycle using R134a–DMA as therefrigerant–absorbent pair with low-temperature heat sourcesfor cold storage applications. The performance of this work-ing fluid pair is better than that of ammonia–water for lowheat source temperatures in the half effect configuration. Muthuet al. [17] conducted experimental studies on 1-kW capacityVARS using R134a–DMA as the working fluid and hot water asthe heat source. The effects of various operating parameters onthe systems performance were evaluated. The study revealed thefeasibility of using R134a–DMA as a working fluid in the futureabsorption machines using low potential heat sources. Reviewof the literature revealed that experimental studies on a single-stage VARS using R134a–DMF are scanty. The heat-exchangingcomponents used by the investigators in their experimental stud-ies on VARS were mainly shell-and-tube heat exchangers andtube-in-tube heat exchangers. Only Flamensbeck et al. [18] usedbrazed plate heat exchangers for absorber, condenser, evapora-tor, and solution heat exchanger, while testing a double effectabsorption heat pump with water and hydroxide as working pair.A better temperature control shall be established when plate heatexchangers are employed in VARS, since the temperature ap-proach in plate heat exchangers may be as low as 1◦C whereasshell-and-tube heat exchangers require an approach of 5◦C ormore. For the same amount of heat exchanged, the size of theplate heat exchanger is smaller, because of the large heat trans-fer area afforded by the plates. Plate heat exchangers have somelimitations also. In situations where there is an extreme temper-ature difference between two fluids, it is generally more cost-efficient to use a shell-and-tube exchanger. There can be highpressure losses in a plate heat exchanger due to the large amountof turbulence created by the narrow flow channels. Shell-and-tube exchangers are considered for applications that require lowpressure losses. Despite these limitations, plate heat exchangersare the most efficient choice for a wide variety of applications.Due to their compactness, better temperature control, high effi-ciency, flexibility, low cost, and ease of maintenance, plate heatexchangers have been used in the present investigation. The ob-

jective of this work is to study performance characteristics andthe effects of generator, evaporator, and condenser temperatureson circulation ratio and coefficient of performance in a 1-TRcapacity vapor absorption refrigeration system, using plate heatexchangers as system components and employing R134a/DMFas the working fluid.

EXPERIMENTAL SETUP

The schematic diagram of experimental setup has been shownin Figure 1. The setup consists of VARS, hot water simulator,cooling load simulator, cooling water simulator, instruments,valves and control devices. Table 1 lists the specifications ofplate heat exchangers used as VARS components, namely, theevaporator, condenser, absorber, generator, and solution heat ex-changer. The VARS comprises a refrigerant circuit and a solutioncircuit. In the refrigerant circuit, R134a vapor coming from thegenerator storage tank is condensed in the condenser and accu-mulated in the receiver. The heat of condensation is removed bythe cooling water. Liquid R134a from the receiver is expandedthrough the throttle valve/capillary tube and evaporated in theevaporator by absorbing the heat. Chilled water acts as the heatload. In solution circuit, R134a vapor from the evaporator isabsorbed in the absorber by weak DMF solution. The heat ofmixing is removed by cooling water. Strong solution collectedin the absorber storage tank is pumped by a hermetically sealedmultistage centrifugal pump through a solution heat exchangerto the generator. R134a vapor boiled off in the generator is sep-arated in the generator storage tank. Hot water is supplied asheat source for the generator. The weak solution remaining inthe tank is transferred through the solution heat exchanger andthe pressure-reducing valve to the absorber for absorption. Thesolution heat exchanger is provided to cool the weak solutionbefore entering the absorber and preheat the strong solution be-fore entering the generator. The hot water simulator consistsof a hot water tank insulated with glass wool, electric heaters,pump, flowmeter, PT100 sensor (uncertainty of ±0.25◦C), PIDtemperature controller, contactor, piping, and valves. It sup-plies hot water to the generator. The cooling water simulatorconsists of two circuits of R22-based vapor compression re-frigeration (VCR) system with 3.4 TR capacity each, a coolingwater tank insulated with expanded polyethylene (EPE) sheets,electric heaters, pump, flowmeter, PT100 sensor, PID tempera-ture controller, contactor, piping, and valves. The VCR circuitconsists of a hermetically sealed reciprocating compressor, air-cooled condenser, thermostatic expansion valve, and coolingcoil. The cooling-water thermostat supplies cooling water to theabsorber and the condenser, which are connected in series andin parallel by cooling-water pipelines. The cooling load simu-lator consists of a chilled water tank insulated with expandedpolyethylene (EPE) sheets, electric heaters, pump, flowmeter,PT100 sensor, PID temperature controller, contactor, piping,and valves. It supplies water as heat load to the evaporator and

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Figure 1 Schematic diagram of R134a–DMF-based vapor absorption refrigeration system.

maintains a constant desired value of chilled water temperatureat the evaporator inlet. The locations of various temperaturesensors, pressure sensors, flowmeters, and valves are shown inFigure 1. All these measuring instruments are precalibrated.Copper–constantan thermocouples of 22 numbers are used astemperature sensors with a measurement uncertainty up to±0.5◦C. The piezoelectric-type pressure transducers (14 num-bers) are used as pressure sensors with a measurement uncer-tainty up to ±1.2%. Metal tube rotameters are used to measureflow of liquid refrigerant, weak solution, and hot water witha measurement uncertainty up to ±4.6%. Glass rotameters areused to measure flow of cooling water and chilled water witha measurement uncertainty up to ±2.5%. An online densitymeter is used to measure the density of strong and weak solu-tions with a measurement uncertainty of ±0.1%. The concentra-tions of strong and weak solutions are evaluated from measureddensity values using the HBT (Hankinson–Brobst–Thomson)equation used by Reid and others in their book [19]. Readingsfrom all these instruments and sensors are monitored contin-uously by connecting them to a data-acquisition system and acomputer.

EXPERIMENTAL PROCEDURES

Initially the refrigerant and solution circuit are separated byclosing the valve between (1) the generator storage tank andcondenser and (2) the evaporator and absorber. Absorbentcharge is calculated based on system volume in solution circuitand charged at the absorber storage tank (59 kg). Refrigerantcharge is calculated based on system volume in both refriger-ant and solution circuits (28 kg). Thus, the initial concentrationof refrigerant in the charging solution is 32% (by mass). It ischarged and mixed with absorbent at the absorber storage tankby circulating the solution through the solution circuit usinga solution pump. Heat of absorption is allowed to be trans-ferred from the solution circuit components to ambient air. Thesolution pump is switched off after charging. The hot watersimulator and cooling water simulator are started. Hot water iscirculated through the generator at a temperature higher than thatto be maintained in the generator. Cooling water is circulatedthrough absorber and condenser in parallel or series circuit at atemperature lower than that to be maintained in both the com-ponents. The cooling load simulator is operated by circulating

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Table 1 Specifications of plate heat exchangers used in VARS

Name of the equipment Quantity Design specifications

Evaporator 1 no. Type : Brazed plate type heat exchangerFluid circuit : Cold side: R134a; hot side: waterFluid direction : CountercurrentHeat transfer area, m2 : 0.6Number of plates : 24Number of effective plates : 22Number of passes : 1

Condenser 1 no. Type : Brazed plate type heat exchangerFluid circuit : Cold side: water; hot side: R134aFluid direction : CountercurrentHeat transfer area, m2 : 0.7Number of plates : 30Number of effective plates : 28Number of passes : 1

Absorber 1 no. Type : Brazed plate type heat exchangerFluid circuit : Cold side: water;

hot side: R134a/DMF solutionFluid direction : CountercurrentHeat transfer area, m2 : 0.6Number of plates : 24Number of effective plates : 22Number of passes : 1

Generator 1 no. Type : Brazed plate type heat exchangerFluid circuit : Cold side: R134a/DMF solution;

hot side: waterFluid direction : CountercurrentHeat transfer area, m2 : 0.6Number of plates : 24Number of effective plates : 22Number of passes : 1

Solution heat exchanger 1 no. Type : Brazed plate type heat exchangerFluid circuit : Cold side: R134a/DMF solution;

hot side: R134a/DMF solutionFluid direction : CountercurrentHeat transfer area, m2 : 1.9Number of plates : 40Number of effective plates : 38Number of passes : 1

water through the evaporator. Water temperature in the chilledwater tank is maintained constant by switching on heaters equiv-alent to the cooling capacity of system. The solution pump isthen started to circulate strong solution through the generator.The level of weak solution collected in the generator storagetank, level of strong solution in the absorber storage tank, andpressure in each component of solution circuit are monitoredcontinuously. When pressure in the generator storage tank be-comes greater than that in the condenser, the valve between thegenerator storage tank and the condenser is opened to allowrefrigerant vapor to enter the condenser. The level of liquid re-frigerant collected in the liquid refrigerant receiver is monitoredcontinuously. When a sufficient amount of refrigerant is stored,it is admitted though expansion devices to enter the evapora-tor. The valve between the evaporator and absorber is openedto allow refrigerant vapor to enter the absorber. Flow rates ofweak solution and liquid refrigerant are regulated to maintaina steady flow in the system. The system is run continuously by

monitoring the pressure transducer, thermocouple, flowmeter,and level gauge readings at various locations. When all thesereadings remain constant over a period of time, it indicates thatsystem has attained steady-state operating conditions and allthese readings are recorded in the computer. Water flow ratesin the hot water simulator, cooling water simulator, and cool-ing load simulator are maintained constant at the desired value.While shutting down the system after recording the readings, thesolution circuit and refrigerant circuit are isolated by closing thevalve between the evaporator and the absorber and then closingthe valve between the generator storage tank and the condenser.Then solution pump is switched off.

RESULTS AND DISCUSSION

Experimentation was conducted with a cooling capacity of 2to 5 kW by varying the operating parameters, namely, the liquid

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Figure 2 Effect of generator temperature on mass of refrigerant entering theevaporator.

Figure 3 Effect of generator temperature on strong and weak solution con-centration difference.

Figure 4 Effect of generator temperature on circulation ratio.

refrigerant flow rate from 0.03 m3 h−1 to 0.09 m3 h−1, solutionflow rate from 0.16 m3 h−1 to 1.6 m3 h−1, hot water tempera-ture from 67 to 95◦C, and cooling water temperature from 15to 30◦C. Water flow rates were maintained constant at 2.45 m3

h−1 for hot water, 2 m3 h−1 for cooling water in series flow path,1.25 m3 h−1 for parallel flow path, and 0.65 m3 h−1 for chilledwater. The parametric studies were carried out from the follow-ing temperature range of operating parameters: generator tem-perature 63–90◦C, condenser temperature 17–32◦C, absorbertemperature 15–30◦C, and evaporator temperature –2.5–10◦C.Cooling water can be supplied to absorber and condenser inparallel as well as series flow arrangements as in Figure 1. Thecalculated quantities and parameters to estimate performancefrom the experimental observations made for a particular op-erating condition of VARS are listed in the appendix. Thoughparallel flow of cooling water supply gives better performance,it requires a large quantity of water, which is not met by thecooling water pump in the system. Hence the system is run atlower capacities with parallel flow and at higher capacities withseries flow. In the series flow arrangement, cooling water firstgoes to the absorber and then to the condenser. Figure 2 showsthat the mass of refrigerant entering the evaporator increasesas the generator temperature increases, which is due to gener-ation of more refrigerant vapor at high generator temperatures.Figure 3 shows that the concentration difference between thestrong solution and weak solution increases as the generator andevaporator temperature increase. At higher generator tempera-tures, the weak solution concentration decreases, and at higherevaporator temperatures, the strong solution concentration in-creases. Both lead to an increase in concentration differencebetween strong and weak solutions. It is observed from Figure 3that theoretical values of concentration difference are higherthan the experimental values. The reason is that theoretical val-ues of strong and weak solution concentrations are calculatedat equilibrium conditions assuming absorber and generator ef-fectiveness as 100%. Experimental values of strong and weaksolution concentrations are calculated from solution densitiesmeasured by online density meter. Figure 4 shows the effectof generator temperature on circulation ratio (CR) at differentevaporator temperatures. CR decreases as the generator tem-perature increases. Since the concentration difference is moreat high generator temperatures, the required solution circulationrate is less to generate the unit mass of refrigerant vapor. The CRis lower at higher evaporator temperatures due to the increasein strong solution concentration. Theoretical values of CR arelower than those of experimental values in Figure 4. The reasonis that the theoretical circulation ratio has been calculated byconsidering equilibrium conditions at generator and absorberoutlets. The experimental circulation ratio has been calculatedfrom actual values of strong and weak solution concentrations.In practice, the actual strong solution concentration is lowerthan the equilibrium value and the actual weak solution con-centration is higher than the equilibrium value. Therefore, theactual difference between strong and weak solution is less thanthat of equilibrium values. This results in higher values of the

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Figure 5 Effect of generator temperature on coefficient of performance.

Figure 6 Effect of condenser temperature on strong and weak solution con-centration difference.

Figure 7 Effect of condenser temperature on circulation ratio.

Figure 8 Effect of condenser temperature on coefficient of performance.

Figure 9 Effect of evaporator temperature on strong and weak solution con-centration difference.

Figure 10 Effect of evaporator temperature on circulation ratio.

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Figure 11 Effect of evaporator temperature on coefficient of performance.

experimental circulation ratio compared to theoretical values. Itis observed from Figure 5 that the coefficient of performance(COP) increases at high generator temperatures. The reason isthat at lower CR, the heat input required at the generator is lessto generate the unit mass of refrigerant vapor. Since COP is theratio of refrigeration effect to generator heat input, the value ishigher at high generator temperatures. The COP is also foundhigher at higher evaporator temperatures. It can be seen fromFigure 5 that values of theoretical COP are higher than the ex-perimental values. The reason is that in theoretical COP, pumpwork has been added as energy input based on enthalpy differ-ence across pump, which is a very small value. In experimentalCOP, pump work has been added as energy input based on pumpmotor energy meter readings, which is a larger value. The pumpused in the experiments was a hermetically sealed multistagecentrifugal pump with a minimum capacity of 1 m3 h−1 and amaximum capacity of 12 m3 h−1. The required capacity for thisoperation is only 0.1 to 0.3 m3 h−1. But such a low-capacity,hermetically sealed high-pressure centrifugal pump could notbe obtained. Hence a higher capacity pump was used and mostof the solution was bypassed to strong solution tank. Due to this,additional electrical power of 1.5 kW was consumed. This makesa large difference between the theoretical COP and experimentalCOP. Figure 6 shows that the concentration difference betweenstrong solution and weak solution decreases as the condensertemperature increases and the evaporator temperature decreases.At higher condenser temperatures, weak solution concentrationincreases, and at lower evaporator temperatures, strong solutionconcentration decreases. Both lead to decrease in concentrationdifference between strong and weak solutions. Figure 7 showsthe effect of condenser temperature on CR at different evapo-rator temperatures. CR increases as the condenser temperatureincreases. Since the concentration difference is less at highercondenser temperatures, the required solution circulation rate ismore to generate unit mass of refrigerant vapor. CR is lower athigher evaporator temperatures due to increase in strong solu-tion concentration. In Figure 8, COP decreases at high condensertemperatures. The reason is that at high CR, heat input required

at the generator is more to generate the unit mass of refrigerantvapor. Since COP is the ratio of refrigeration effect to generatorheat input, the value is less at high condenser temperatures. COPis more at higher evaporator temperatures. Figures 9, 10, and11 show the effect of evaporator temperature on concentrationdifference between strong and weak solution, CR, and COP, re-spectively. At high evaporator temperatures, the strong solutionconcentration increases, resulting in an increase in concentra-tion difference between strong and weak solutions. CR is lowerat high evaporator temperatures due to increased concentrationdifference. Also, COP is more at high evaporator temperaturesdue to lower circulation ratios.

CONCLUSIONS

An experimental setup for 1 TR capacity vapor absorptionrefrigeration system with plate heat exchangers is designed anddeveloped. R134a–DMF is used as the working fluid and hotwater is used as the heat source to supply heat to the generator.The bubble absorption principle is employed in the absorber.Experiments are carried out to investigate the effect of operat-ing parameters such as the generator, condenser, and evaporatortemperatures on system performance, and the following conclu-sions have been drawn from the experimental results.

• The concentration difference between the strong and weak so-lution is higher at high generator and evaporator temperatures,whereas it is lower at higher condenser temperatures.

• The circulation ratio is lower at high generator and evapo-rator temperatures, whereas it is higher at higher condensertemperatures.

• The coefficient of performance is higher at high generatorand evaporator temperatures, whereas it is lower at highercondenser temperatures.

• With addition of a rectifier as well as improvement of vaporseparation in the generator storage tank, the R134a/DMF-based compact vapor absorption refrigeration system withplate heat exchangers could be very competitive for appli-cations ranging from –10◦C to 10◦C, with hot water as heatsource in the temperature range of 80◦C to 90◦C and withcooling water as coolant for the absorber and condenser in thetemperature range of 20◦C to 35◦C.

NOMENCLATURE

Cp specific heat at constant pressure, kJ kg−1 K−1

E electrical energy, kWh specific enthalpy, kJ kg−1

m mass flow rate, kg s−1

P pressure, kPaQ heat energy, kWT temperature, ◦CX concentration by mass, kg kg−1

v specific volume, m3 kg−1

W power, kW

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Subscripts

a, A absorberc, C condenserchw chilled watercw cooling watere, E evaporatoreq equilibriumg generator, gasG generatorhw hot waterHX heat exchangerP pumpr refrigerantss strong solutionth thermodynamic, theoreticalws weak solution

REFERENCES

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[2] Fatouh, M., and Srinivasa Murthy, S., HCFC22-BasedVapour Absorption Refrigeration System, Part I. Paramet-ric Studies, International Journal of Energy Research, vol.20, 297–312, 1996.

[3] Fatouh, M., and Srinivasa Murthy, S., HCFC22-BasedVapour Absorption Refrigeration System, Part II. Influ-ence of Component Effectiveness, International Journalof Energy Research, vol. 20, 371–384, 1996.

[4] Fatouh, M. and Srinivasa Murthy, S., HCFC22-BasedVapour Absorption Refrigeration System, Part III. Effectsof Different Absorber and Condenser Temperatures, Inter-national Journal of Energy Research, vol. 20, 483–494,1996.

[5] Karthikeyan, G., Mani, A., and Srinivasa Murthy, S., Anal-ysis of an Absorption Refrigeration System With TransferTank, Renewable Energy, vol. 4, pp. 129–132, 1994.

[6] Karthikeyan, G., Mani, A., and Srinivasa Murthy, S., Per-formance of Different Working Fluids in Transfer-TankOperated Vapour Absorption Refrigeration Systems, Re-newable Energy, vol. 6, pp. 835–842, 1995.

[7] Sujatha, K. S., Mani, A., and Srinivasa Murthy, S., Fi-nite Element Analysis of a Bubble Absorber, InternationalJournal for Numerical Methods for Heat & Fluid Flow,vol. 7, pp. 737–750, 1997.

[8] Sujatha, K. S., Mani, A., and Srinivasa Murthy, S., Analysisof a Bubble Absorber Working With R22 and Five OrganicAbsorbents, Heat and Mass Transfer, vol. 32, pp. 255–259,1997.

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[10] Borde, I., Jelinek, M. and Daltrophe, N. C., Refrigerant–Absorbent Mixtures Based on the Refrigerant R134a, Proc.XVIIIth Int. Conf. of Refrigeration, Montreal, Canada, pp.653–658, 1991.

[11] Borde, I., Jelinek, M., and Daltrophe, N. C., AbsorptionSystem Based on the Refrigerant R134a, InternationalJournal of Refrigeration, vol. 18, pp. 387–394, 1995.

[12] Songara, A. K., Fatouh, M., and Srinivasa Murthy, S,Comparative Performance of HFC134a- and HCFC22-Based Vapour Absorption Refrigeration Systems, Inter-national Journal of Energy Research, vol. 22, pp. 363–372, 1998.

[13] Songara, A. K., Fatouh, M., and Srinivasa Murthy, S, Ther-modynamic Studies on HFC134a-DMA Double Effectand Cascaded Absorption Refrigeration Systems, Interna-tional Journal of Energy Research, vol. 22, pp. 603–614,1998.

[14] Nezu, Y., Hisada, N., Ishiyama, T., and Watanabe, K.,Thermodynamic Properties of Working-Fluid Pairs WithR-134a for Absorption Refrigeration System, in NaturalWorking-Fluids, IIR Gustav Lorentzen Conf. 5th, China,September 17–20, pp. 446–453, 2002.

[15] Yokozeki, A., Theoretical Performances of VariousRefrigerant–Absorbent Pairs in a Vapour Absorption Re-frigeration Cycle by the Use of Equation of State, AppliedEnergy, vol. 80, pp. 383–399, 2005.

[16] Arivazhagan, S., Murugesan, S. N., Saravanan, R., andRenganarayanan, S., Simulation Studies on R134a-DMACBased Half Effect Absorption Cold Storage Systems, En-ergy Conversion and Management, vol. 46, 1703–1713,2005.

[17] Muthu, V., Saravanan, R., and Renganarayanan, S., Experi-mental Studies on R134a-DMAC Hot Water Based VapourAbsorption Refrigeration Systems, International Journalof Thermal Sciences, vol. 47, 175–181, 2008.

[18] Flamensbeck, M., Summerer, F., Riesch, P., Ziegler, F., andAlefeld, G., A Cost Effective Absorption Chiller With PlateHeat Exchangers Using Water and Hydroxides, AppliedThermal Engineering, vol. 18, 413–425, 1998.

[19] Reid, R. C., Prausnitz, J. M., and Poling, B. E., The Prop-erties of Gases and Liquids, 4th ed., McGraw-Hill, NewYork, NY, 1987.

APPENDIX

Thermodynamic Heat Quantities

Heat absorbed at evaporator:

QE = mr (h6 − h5) (1)

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984 S. MARIAPPAN AND M. ANNAMALAI

Heat rejected at condenser:

QC = mr (h2 − h3) (2)

Heat rejected at absorber:

Q A = mr h6 + mwsh14 − mssh7 (3)

where

mss = (mws + mr ) (4)

Heat supplied at generator:

QG = mr h11r + mwsh11s − mssh10 (5)

where h11r and h11s are refrigerant vapor enthalpy and weaksolution enthalpy at state 11, respectively.

Heat recovered by strong solution at solution heat exchanger:

Q H X,SS = mss(h10 − h9) (6)

Pumping power required to pump strong solution:

WP = mssvss(P9 − P8) (7)

where vss is specific volume of strong solution at solution pumpinlet, that is, state 8.

Experimental Heat Quantities

Experimental heat quantities determined from measured wa-ter flow rate and temperature difference across each componentare as follows.

Heat absorbed at evaporator:

Qe = mchwCpchw(T21 − T22) (8)

Heat rejected at condenser:

Qc = mcwCpcw(T20 − T19) (9)

Heat rejected at absorber:

Qa = mcwCpcw(T18 − T17) (10)

Heat supplied at generator:

Qg = mhwCphw(T15 − T16) (11)

Specific heat capacities are calculated at average temperatureof water through each component.

Performance Characteristics

System performance is evaluated by means of two impor-tant parameters: circulation ratio (CR) and coefficient of perfor-mance (COP). CR is defined as the ratio of mass flow rate ofstrong solution to that of refrigerant.

Thermodynamic CR is calculated by assuming equilibriumconditions at absorber and generator outlets:

CRth = Xg − Xws,eq

Xss,eq − Xws,eq(12)

whereXg is considered the concentration of R134a gas, givenas 99.9% as per manufacturer’s specifications; Xss,eq is equi-librium concentration calculated from absorber outlet pressureand temperature; and Xws,eq is equilibrium concentration calcu-lated from generator outlet pressure and temperature, using ex-perimental bubble point pressure correlations for R134a–DMFsolution developed by Nezu et al. [14].

Actual CR is determined from measured values of refrigerantand weak solution flow rates:

CR = mr + mws

mr(13)

COP is defined as the ratio of refrigeration effect producedby the system to energy supplied to the system. Cooling load atthe evaporator is termed the refrigeration effect from the system.Heat input supplied to the generator and power supplied to thesolution pump are termed energy inputs supplied to the system.Theoretical COP is calculated as

COPth = QE

QG + WP(14)

Actual COP is determined as

COP = Qe

Qg + E p(15)

where E p is electrical energy required for solution pump topump strong solution and is determined from energy meterreadings. All numerical subscripts in this appendix are withreference to the measurement locations indicated in Figure 1.

Suresh Mariappan is currently an associate profes-sor at SSN College of Engineering, India. He receivedhis M.E. degree from Anna University, India, and hisPh.D. degree from the Indian Institute of TechnologyMadras, India, under the supervision of Prof. ManiAnnamalai. He is currently working on vapor absorp-tion refrigeration.

Mani Annamalai is a professor of mechanical engi-neering at the Indian Institute of Technology Madras(IITM), Chennai, India. He received his M.Tech. andPh.D. degrees from IITM. He has been the head ofthe Refrigeration and Air-Conditioning Laboratoryof the Department of Mechanical Engineering for sixyears and also held several administrative positionsin IITM. His research contributions are in the fieldof vapor compression refrigeration, vapor absorptionrefrigeration, vapor jet refrigeration, liquid effluent

concentration systems, desalination, cryogenic heat transfer, and solar thermalapplications. He has contributed more than 100 papers in the area of thermalengineering.

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