1、534 2010 ASHRAEABSTRACT This paper presents the experimental as well as the simu-lated performance studies on the transcritical CO2heat pumpsfor simultaneous water cooling and heating; effects of watermass flow rates and water inlet temperatures of both evapo-rator and gas cooler on the cooling and
2、heating capacities,system COP and water outlets temperatures are investigated.Study shows that both the water mass flow rate and inlettemperature have significant effect on system performances.Test results show that the effect of evaporator water mass flowrate on the system performances and water ou
3、tlet temperaturesis more pronounced (COP increases 0.6 for 1 kg/min)compared to the gas cooler water mass flow rate (COPincreases 0.4 for 1 kg/min) and the effect of gas cooler waterinlet temperature is more significant (COP decreases 0.48 forgiven ranges) compared to the evaporator water inlet temp
4、er-ature (COP increases 0.43 for given ranges). Comparisons ofexperimental values with simulated results show the maximumdeviation of 5% for cooling capacity, 10% for heating capacity,16% for system COP. This study offers useful guidelines forselecting appropriate water mass flow rate to obtain requ
5、iredsystem performance.INTRODUCTIONRecently, natural fluid CO2has been shown to be a prom-ising alternative particularly in heat pump applications due togliding temperature heat rejection in the gas cooler and perfor-mance related various advantages. Some of the current theo-retical and experimental
6、 investigations on several heat pumpapplications of transcritical CO2cycle have been presented byNeksa (2002) and Kim et al. (2004). Neksa et al. (1998) and White et al. (2002) experimentallyinvestigated the effects of discharge pressure, water inlet andoutlet temperatures on the CO2heat pump water
7、heater perfor-mances. Hwang et al. (1999) experimentally showed that theactual CO2cycle performed similar to the R22 cycle when thesame outside volume of the heat exchanger was employed forboth refrigerants. Yarral et al. (1999) experimentally investi-gated to effect of discharge pressure on CO2heat
8、 pump perfor-mance for simultaneous production of refrigeration and waterheating to 90oC for the food processing industry. Adriansyah(2004) experimentally studied the effect of compressordischarge pressure for simultaneous air-conditioning andwater heating. Stene (2005) presented the effect of water
9、 inlettemperature on CO2heat pump performance for combinationof space and water heating. Cho et al. (2005) studied theperformance of the CO2heat pump by varying the refrigerantcharge amount at standard cooling condition to show theimportance of refrigerant charge to achieve better perfor-mance. Kim
10、et al. (2005) have done experimental study onCO2heat pump to study the effect of internal heat exchangerusing water as secondary fluid for both sides with emphasizesonly on heating. Sarkar et al. (2006) numerically studied theeffects of water inlet temperature, compressor speed and heatexchanger inv
11、entory on optimal performance for simultane-ous water cooling and heating applications. Yokoyama (2007)experimentally studied the influence of ambient temperatureon performance of a CO2heat pump water heating system.Cabello et al. (2008) experimentally evaluated the effect ofoperating temperatures o
12、n optimal gas-cooler pressures of aCO2refrigerating plant and showed that Sarkar et al. (2004)correlation matches best with the test data of optimal gas-cooler pressures. However, effects of water mass flow rate onPerformance of a Transcritical CO2Heat Pump for Simultaneous Water Cooling and Heating
13、J. Sarkar, PhD Souvik Bhattacharyya, PhD M. Ram Gopal, PhDJ. Sarkar is a lecturer in the Department of Mechanical Engineering, Institute of Technology-BHU, Varanasi, India. Souvik Bhattacharyyais a professor and M. Ram Gopal is an associate professor in the Department of Mechanical Engineering, Indi
14、an Institute of Technology,Kharagpur India. OR-10-056 2010, American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. (www.ashrae.org). Published in ASHRAE Transactions 2010, Vol. 116, Part 1. For personal use only. Additional reproduction, distribution, or transmission in eith
15、er print or digital form is not permitted without ASHRAEs prior written permission. ASHRAE Transactions 535heat pump behavior have not been reported in the open liter-ature.In the present investigation, both simulation and experi-mental results on the working prototype of a transcritical CO2heat pum
16、p system for simultaneous water cooling and heatingare presented. The cooling and heating capacities, systemCOP and water outlets temperatures have been studied forvarious water mass flow rates and water inlet temperatures ofboth evaporator and gas cooler. Comparisons of simulated andexperimental re
17、sults are presented as well.EXPERIMENTS OF A TRANSCRITICAL CO2HEAT PUMPDescription of Test SetupTest facility layout of transcritical CO2heat pump forsimultaneous water cooling and heating with instrumentalpositions is shown in Figure 1. Stainless steel was chosen asthe material for all system compo
18、nents. A Dorin CO2compressor (model TCS113: displacement = 2.2 m3/hr andcapacity = 2.5 kW 2900 rpm) was chosen for the experi-mental investigation. On the basis of minimum and maximumpressure ratios of 80/50 and 120/26 (bar/bar), respectively, aSwagelok integral bonnet needle valve (model SS-1RS4) w
19、asused as the expansion device, which can be used regulate flowrate and degree of superheat. The separator and receiver weredesigned for a total volumetric capacity of 8 L and 2 L, respec-tively. A condensing unit including a fan and a storage tankwas employed for a heat transfer rate of 6 kW to coo
20、l the warmwater to its initial temperature at the inlet to the gas cooler. Awater bath with heater and pump was incorporated in the evap-orator to supply water at constant temperature and flow rate,so that a cooling capacity of 3.5 kW can be obtained. The evap-orator and the gas cooler are counter-f
21、low tube-in-tube heatexchangers, where CO2flows in the inner tube and water in theouter annulus (Table 1). Measuring ranges of instruments withuncertainties are listed in Table 2.Test Procedure and Test ConditionsFor the experimental study, condensing unit for the gascooler and heating unit for the
22、evaporator were started to stabi-lize the temperature in the system and total refrigerant chargepresent in the system was estimated. After recording initialreading the compressor was switched on. Control of dischargepressure was achieved by simultaneous control of the totalmass of CO2in the system a
23、nd degree of opening of the expan-sion device. If the total refrigerant mass in the system needs tobe changed (required to change the discharge pressure), it isimplemented through system charge moderation by addingCO2from a high pressure cylinder or by venting it through thesafety valve. Constant in
24、let water temperatures to gas coolerFigure 1 Test facility layout of the transcritical CO2heatpump.Table 1. Dimensions of Gas Cooler and Evaporator Heat ExchangersGas cooler EvaporatorConfigurationCoaxial, single pass, 14 rows Coaxial, Single pass, 9 rowsInner Tube OD/Outer Ttube OD6.35 mm/12 mm 9.5
25、mm/16mmTotal Length of Tubing14 m 7.2 mTable 2. Ranges and Uncertainties of Measuring InstrumentsParameters Measuring Instruments Ranges AccuracyPressure Dial pressure gauge 0-160 bar 1.5% of full range Pressure loss Differential pressure gauge 0-4 bar 1.5% of full rangeCO2mass flow rate Mass flow m
26、eter 0.210 kg/min 0.1% of full rangeWater mass flow rate Mass flow meter 0.520 kg/min 0.5% of full rangeTemperature Thermocouples (T-type, K-type) Calibrated range: 0-150 oC 0.5 2010, American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. (www.ashrae.org). Published in ASHRA
27、E Transactions 2010, Vol. 116, Part 1. For personal use only. Additional reproduction, distribution, or transmission in either print or digital form is not permitted without ASHRAEs prior written permission. 536 ASHRAE Transactionsfrom condensing unit and to evaporator from heating unit aremaintaine
28、d by thermostatic control. The main operatingparameters are discharge pressure, water inlet temperature andmass flow rate in gas cooler, and water inlet temperature andmass flow rate in evaporator. The compressor power input wasmeasured by using a power meter, the refrigerant mass flowrate was measu
29、red by a Coriolis effect flow meter, the pressureof the refrigerant was measured by using a dial gauge, pressuredrop in the heat exchangers was measured by differential pres-sure transducer and refrigerant and water temperatures at allrequired locations were measured by using T-type Sarkar et al., 2
30、006). Employing LMTD expression,heat transfer in ith segment of the gas cooler is given by,(4)Additionally, energy balance in segment of gas cooler for boththe fluids yield:(5)The overall heat transfer coefficient for the segment of gascooler has been calculated using the fundamental equation forove
31、rall heat transfer coefficient.To estimate the heat transfer coefficient of supercriticalcarbon dioxide for in-tube cooling in gas cooler, Pitla et al.(2002) correlation, incorporating both bulk and wall proper-ties due to large variation of fluid properties in the radial direc-tion, has been used.
32、The pressure drop for supercritical carbondioxide in-tube cooling has been calculated by Petrov andPopov equation (Fang et al., 2001), neglecting inertia effect.The waterside heat transfer coefficient has been evaluated bythe Gnielinski (1976) equation for annular flow. All waterproperties are assum
33、ed to be temperature dependent only, andpolynomial expressions based on text book values have beenused. Evaporator ModelThe evaporator consists of two zones: two-phase (boiling)zone and superheated zone. Similar to the gas cooler, bothzones in the evaporator are divided into a finite number ofequal-
34、length segments along the refrigerant flow direction.Each segment is treated as one counter-flow heat exchangerand the outlet conditions of each segment should become inletconditions for the next segment. For each segment LMTDmethod is used and properties are evaluated based on meantemperature and p
35、ressure. Energy balance in each segment ofthe evaporator for the refrigerant (CO2) and water, respec-tively, yields:(6)The overall heat transfer coefficient for each segment ofthe evaporator has been calculated in the same way as for thegas cooler. In this analysis, the recently developed Yoon et al
36、.(2004) correlation has been employed to estimate the boilingheat transfer coefficient. For superheated zone, Gnielinski(1976) equation has been used to estimate convective heattransfer coefficient of carbon dioxide. Jung and Radermacher(1989) correlation has been used for boiling pressure drop andB
37、lasius correlation has been used for single phase pressuredrop of carbon dioxide. The waterside heat transfer coefficienthas been evaluated by Gnielinski (1976) equation for annularflow for both two-phase and superheated sections.Using discretization, the heat exchanger is made equiva-lent to a numb
38、er of counter flow heat exchangers arranged inseries and the combined heat transfer of all the segments is thetotal heat transfer of the heat exchanger. Therefore, fast chang-ing properties of CO2have been modeled accurately in bothevaporator and gas cooler.Expansion Device ModelThe expansion proces
39、s is considered to be isenthalpicunder the assumption that the heat exchange with its surround-ings is negligible, yielding:(7)Numerical ProcedureA computer code, incorporating the subroutineCO2PROP (Sarkar et al., 2006) for thermophysical andtransport properties, has been developed to simulate the
40、tran-scritical carbon dioxide system for simultaneous water cool-ing and heating at various operating conditions. Water inlettemperatures and water mass flow rates for both heat exchang-ers, compressor data, evaporator and gas cooler dimensions,compressor suction pressure and discharge pressure are
41、theinput data for the simulation. The flow chart of the simulation is shown in Figure 3.Pressure drop and heat loss in connecting lines are not consid-ered; therefore, the outlet state of one component becomes theinlet state of the next component. In the simulation, by assum-ing the suction temperat
42、ure, refrigerant mass flow rate andcompressor outlet conditions are calculated by compressormodel, and refrigerant conditions at evaporator inlet and at gascooler outlet as well as water outlet temperatures of both evap-orator and gas cooler are calculated based on mathematicalmodel of evaporator an
43、d gas cooler. The suction temperatureis adjusted by the iteration in order for the enthalpy of inlet (h3)and outlet (h4) of expansion valves to converge within aprescribed tolerance and performances such as cooling andheating output, compressor work and COP are calculated.Tolerance has been maintain
44、ed in the range of 103for simu-lation. RESULTS AND DISCUSSIONThe performance of the CO2heat pump system in termsof cooling or heating capacity and system COP (cooling +heating outputs divided by compressor power) consideringboth cooling and heating as useful outputs are studied for theis c,0.61 0.03
45、56PdisPsuc-0.0257PdisPsuc-20.0022PdisPsuc-3+=QgciUA()gciTgc r,iTgc w,i()Tgc r,i 1+Tgc w,i 1+()Tgc r,iTgc w,iTgc r,i 1+Tgc w,i 1+-ln-=Qgcimrhgc r,ihgc r,i 1+()mgc w,cpwTgc w,iTgc w,i 1+()=Qevimrhev r,i 1+hev r,i()mev w,cpwTev w,i 1+Tev w,i()=h4h3= 2010, American Society of Heating, Refrigerating and
46、Air-Conditioning Engineers, Inc. (www.ashrae.org). Published in ASHRAE Transactions 2010, Vol. 116, Part 1. For personal use only. Additional reproduction, distribution, or transmission in either print or digital form is not permitted without ASHRAEs prior written permission. 538 ASHRAE Transactions
47、suction and discharge pressure of 40 bar and 90 bar, respec-tively. Both the numerical and experimental results arepresented to study the effect of water inlet temperatures andmass flow rates on the performances and water outlettemperatures. Effects of water mass flow rate to evaporator on theperfor
48、mances and water outlet temperatures for both gascooler and evaporator are shown in Figures 4 and 5, respec-tively, for evaporator water inlet temperature of 29oC, gascooler water inlet temperature of 33oC and water mass flowrate of 1 kg/min. With increase in water mass flow rate to evap-orator, the
49、 cooling capacity increases due to increase in waterside heat transfer coefficient and both the heating capacity andcompressor work increase modestly due to minor increase inthe suction temperature (increase in degree of superheat) andalso discharge temperature. Water outlet temperature of evap-orator increases due to dual effect of increase in cooling capac-ity and water mass flow rate; whereas water ou
