Proceedings of the Sixth International Exergy, Energy ...

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laminated type indoor coil, two thermostatic expansion valves, a reversing valve to operate the system in reverse direction in the heat pump operations, a brazed ...
Proceedings of the Sixth International Exergy, Energy and Environment Symposium (IEEES-6)

Edited by A. MİDİLLİ and H. KÜÇÜK 1-4 July 2013 ISBN: 978-605-85878-0-9

Recep Tayyip Erdoğan University Rize, Turkey

FLOW AND HEAT TRANSFER CHARACTERISTICS OF AN EMPTY REFRIGERATED CONTAINER

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CERTIFICATION AND MAINTENANCE OF INDUSTRIAL INSTALLATIONS IN THE COMPANY TREFILEST ACCORDING TO STANDARD ISO 9001 /2000

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PLASMA-FUEL SYSTEMS UTILIZATION FOR ECOLOGICAL AND ENERGY EFFICIENCY OF THERMAL POWER PLANTS

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PLASMA TECHNOLOGIES OF SOLID AND GASEOUS FUELS PROCESSING

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EXERGY ANALYSIS OF AN ETHYLENE GYLCOL RECOVERY SYTEM USING BATCH DISTILLATION

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INVESTIGATION OF EXERGY RATIOS OF A SOLAR POND AT VARIOUS REFERENCE TEMPERATURES

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ENERGY AND EXERGY ANALYSIS OF AN R134A AUTOMOTIVE HEAT PUMP SYSTEM FOR VARIOUS HEAT SOURCES IN COMPARISON WITH BASELINE HEATING SYSTEM

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EXERGY AND ENERGY EVALUATION OF DISTRICT HEATING SYSTEM

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EXERGETIC PERFORMANCE ASSESSMENT OF A BINARY GEOTHERMAL POWER PLANT EXERGETIC EVALUATION OF A HIGH-PRESSURE HYDROGEN PRODUCTION SYSTEM POLYPYRROLE/CARBON SUPPORTED Platınum CATALYSTS FOR PEM FUEL CELLS COMPARISON of NATURAL GAS FIRED and INDUCTION HEATING FURNACES THE USE OF HAZELNUT OIL ETHYL ESTER AS A FUEL IN PRE-CHAMBER DIESEL ENGINE

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DEVELOPMENT OF A REDUCED MECHANISM FOR N-HAPTANE FUEL IN HCCI ENGINES

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THREE DIMENSIONAL NUMERICAL MODELLING OF HYDROGEN COMBUSTION IN A SPHERICAL MODEL COMBUSTOR

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PRE-COMBUSTION CO2 CAPTURE BY PD BASED COMPOSITE MEMBRANE DEPOSITED ON POROUS NICKEL SUPPORT

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MATHEMATICAL MODEL OF PETROLEUM DYNAMICS IN CLOSED CONDUIT

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A STATISTICAL ANALYSIS OF LEAN MISFIRES IN A GASOLINE ENGINE AND THE EFFECT OF HYDROGEN ADDITION

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A NEW APPROACH FOR TURBOMACHINERY MODELLING BY USING ANFIS STRUCTURE

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OVERVIEW OF IONIC LIQUIDS USED AS WORKING FLUIDS IN ABSORPTION CYCLES

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FULL SCALE EXPERIMENTAL STUDIES OF A PASSIVE COOLING ROOF IN HOT ARID AREAS

V

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ENERGY AND EXERGY ANALYSIS OF AN R134A AUTOMOTIVE HEAT PUMP SYSTEM FOR VARIOUS HEAT SOURCES IN COMPARISON WITH BASELINE HEATING SYSTEM 1,*

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Murat Hosoz , Mehmet Direk , Kadri S. Yigit , Mustafa Canakci , Ali Turkcan , 4

Ertan Alptekin , 1

Department of Automotive Engineering, Kocaeli University, Kocaeli, 41380, Turkey 2 Yalova Community School, Yalova University, Yalova, 77100, Turkey 3 Department of Mechanical Engineering, Kocaeli University, Kocaeli, 41380, Turkey 4 Department of Mechanical Education, Kocaeli University, Kocaeli, 41380, Turkey * Corresponding author. Tel: +90 262 3032279, Fax: +90 262 3032203, E-mail:[email protected]

ABSTRACT Performance of an automotive heat pump (AHP) system using R134a and driven by a diesel engine has been evaluated in this study. For this purpose, an experimental AHP system, capable of providing a conditioned air stream by utilizing the heat absorbed from the ambient air, engine coolant and exhaust gas, was developed. The experimental system was equipped with instruments for measuring engine torque and speed, refrigerant and coolant mass flow rates, refrigerant and air temperatures as well as refrigerant pressures. The system was tested by varying the engine speed, engine load and air temperatures at the inlets of the indoor and outdoor coils. Using experimental data, an energy analysis of the system was performed, and its performance parameters for each heat source were evaluated for transient and steady-state regimes. Then, the performance of the AHP system for each source was compared with that of the system using other heat sources and with that of the baseline heating system. The investigated performance parameters include air temperature at the outlet of the indoor coil, heating capacity, coefficient of performance, and exergy destruction in the components of the AHP system. The exergy destruction rate in the AHP with engine coolant is higher than those in the AHP with ambient air and with exhaust gas mainly because of the greater refrigerant mass flow rate and heating capacity. Keywords: automotive heat pump, R134a, exergy

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INTRODUCTION Under cold weather conditions, comfort heating of the passenger compartments of vehicles with an internal combustion engine is usually performed by utilizing the engines waste heat. However, this coolant-based heating system cannot provide an appropriate thermal comfort in the compartment until the coolant temperature rises to a certain value. This problem is more critical for the vehicles employing high-efficiency diesel engines due to the lack of sufficient waste heat within an acceptable duration of operation after the engine is started (Wienbolt and Augenstein, 2003). With the intention of obtaining thermal comfort rapidly, some vehicles utilize heaters using fuel or electricity. However, these systems contain several disadvantages such as high initial and operating costs, low efficiency and leading to air pollution as well as global warming. On the other hand, the problem of insufficient heating can be solved by adding some low cost components to the present air conditioning system of the vehicle to operate it as a heat pump. The automotive heat pump (AHP) system can heat the passenger compartment individually, or it can support the present heating system of the vehicle (Meyer et al., 2004). There have been very few research studies on AHP systems due to its competitive nature. Domitrovic et al. (1997) simulated the steady state cooling and heating operation of an automotive air conditioning (AAC) and AHP system using R12 and R134a, and determined the change of the cooling and heating capacities, COP and power consumption with ambient temperature at a fixed compressor speed. They found that R134a and R12 yield comparable results, while the heating capacity of the AHP system was insufficient in both refrigerant cases. Antonijevic and Heckt (2004) developed and evaluated the performance an R134a AHP system, which was employed as a supplementary heating system. They carried out the tests at very low ambient temperatures and compared the performance of the AHP system with that of other supplemental heating systems. Rongstam and Mingrino (2003) evaluated the performance of an R134a AHP system using engine coolant as a heat source, and compared it with the performance of a coolant-based heating system at –10ºC ambient temperature. They found that, during a drive cycle operation, the average cabin temperature provided by the heat pump system at the 10th minute of the tests was 8ºC higher than those provided by the baseline heating system. Scherer et al. (2003) reported an on-vehicle performance comparison of an R152a and R134a AHP system using engine coolant as a heat source. They presented the air temperatures at several locations inside the passenger compartment as a function of time, and found that both refrigerants yield almost identical performances and heating capacities. Hosoz and Direk (2006) evaluated the performance of an air-to-air R134a AHP system, and compared its performance with the performance of an air conditioning system. They observed that the AHP system using ambient air as a heat source could not meet the heating requirement of the compartment when ambient temperature was extremely low. Direk and Hosoz (2008) carried out an energy and exergy analysis of an R134a AHP system using ambient air as a heat source. They observed that the heat exchangers of the system are responsible for most of the exergy destruction. Tomoichiro et al. (2005) studied on the

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experimental performance of an AHP system using CO2 as a refrigerant. They found that the performance of the heat pump system using CO2 is equal to or exceeding that of the system using R134a. Kim et al. (2007) investigated the heating performance of an AHP system with CO2. They performed transient and steady state tests under various operating conditions, finding that the use of their system improved heating capacity compared to the baseline heating system. Cho et al. (2009) reported on the heating performance characteristics of a coolant source heat pump using the waste heat of electric devices for an electric bus and suggested a heat pump for an electric bus to be a heating system of over 23.0 kW for cold weather conditions. In this study, the performance parameters of an experimental AHP system for the cases of using the heat from ambient air, engine coolant and exhaust gas were evaluated, and compared with each other along with the performance of the baseline heating system. The investigated performance parameters are heating performance, compressor power, coefficient of performance, exergy destruction in the components and total exergy destruction in the AHP system. DESCRIPTION OF THE EXPERIMENTAL SETUP As schematically shown in Fig. 1, the experimental AHP system is usually made from original components of a compact size AAC system. It employs a seven-cylinder fixedcapacity swash-plate compressor, a parallel-flow micro-channel outdoor coil, a laminated type indoor coil, two thermostatic expansion valves, a reversing valve to operate the system in reverse direction in the heat pump operations, a brazed plate heat exchanger between the engine coolant and the refrigerant to serve as an evaporator and another plate heat exchanger to extract heat from the exhaust gas. All lines in the refrigeration circuit of the system were made from copper tubing, and insulated by elastomeric material. The indoor and outdoor coils were inserted into separate air ducts of 1.0 m length. In order to provide the required air streams in the air ducts, a centrifugal fan and an axial fan were placed at the entrances of the indoor and outdoor air ducts, respectively. These ducts also contain electric heaters located upstream of the indoor and outdoor coils. The indoor and outdoor coil electric heaters can be controlled between 0–2 kW and 0–6 kW, respectively, to provide the required air temperatures at the inlets of the related coils. The refrigeration circuit was charged with 1600 g of R134a. Table 1 Characteristics of the instrumentation Measured variable Instrument Temperature Type K thermocouple Pressure Pressure transmitter Air flow rate Anemometer Refrigerant mass flow rate Coriolis flow meter Compressor speed Digital tachometer Torque Hydraulic dynamometer

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Range –50–500 ºC 0–25 bar -1 0.1–15 m s -1 0–350 kg h 10–100000 rpm 5–750 Nm

Uncertainty ± % 0.3 ± % 0.2 ±%3 ± % 0.1 ±%2 ±%2

Fig. 1. Schematic illustration of the experimental AHP system using various heat sources

In order to gather data for the performance evaluation of the experimental AHP system, some mechanical measurements were conducted on the system. The employed instruments and their locations are depicted in Fig. 1. The characteristics of the instrumentation can be seen in Table 1. The refrigerant mass flow rate was measured by a Coriolis mass flow meter. In order to ensure that the refrigerant is in a liquid phase before entering the mass flow meter, a liquid receiver with a volume of 1.1 litres was placed between the condenser and filter/drier. Additionally, in order to observe if the refrigerant stream contains moisture, a sight glass was located just upstream of the mass flow meter. The temperatures of the refrigerant and air at the inlet and outlet of each component were measured by K-type thermocouples. The refrigerant pressures at the inlet and outlet of the compressor were monitored by both Bourdon gauges and pressure transmitters. The AHP system was driven by a Fiat Doblo JTD diesel engine with a cylinder volume of 1.9 litres and a maximum power of 77 kW at 4000 rpm. The engine torque and speed were measured by means of a hydraulic dynamometer (Baturalp-Taylan BT-190 FR) with a maximum measuring power of 100 kW, a maximum torque of 750 Nm and a maximum speed of 6000 rpm. The fuel consumption was measured using an electronic scale and a chronometer. The values of the measured variables were usually acquired through a data acquisition system and recorded on a computer. The data acquisition system has 16 bit - 200 KHz frequency with a 56 channel thermocouple input module and an 8 channel transducer interface module.

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Fig. 1 also illustrates the refrigerant flow paths in the experimental heat pump system for the cases of using ambient air or exhaust gas as a heat source. In order to perform the heat pump operation, the reversing valve is energized. Then, the reversing valve directs the high temperature superheated vapour refrigerant discharged from the compressor to the indoor coil (condenser). The refrigerant passing through the indoor coil rejects heat to the indoor air stream, thus providing a warm air stream for the potential passenger compartment. After rejecting heat to the indoor air stream, the refrigerant condenses and leaves the indoor coil as subcooled liquid. In heat pump operations, thermostatic expansion valve #1 (TXV1) is bypassed. Then, the refrigerant flows through valves V1 and V2 and reaches the receiver tank, which keeps the unrequired refrigerant in it when the thermostatic expansion valve decreases the refrigerant flow rate at low cooling loads. After passing through the filter, sight glass and Coriolis flow meter, the refrigerant reaches TXV2 located at the inlet of the outdoor coil. Since valves V5 and V8 are closed, the refrigerant passes through TXV2, which reduces the pressure, and thereby the temperature, of the liquid refrigerant. TXV2 also controls the refrigerant mass flow rate circulating through the circuit so that a constant superheat at the outlet of the outdoor coil is maintained under all operating conditions. Next, it enters the outdoor coil, where it absorbs heat taken from the outdoor air stream, and leaves the outdoor coil as low pressure superheated vapour. After leaving the outdoor coil, the refrigerant passes through valve V6, and enters the reversing valve. The refrigerant flow paths in the experimental AHP system using exhaust gas as a heat source are the same as those shown in Fig. 1. In this case, the outdoor air stream is heated by the exhaust gas in a heat exchanger located upstream of the outdoor coil, and then the evaporating refrigerant absorbs its heat in the outdoor coil. Note that in this operation, the bulb of TXV2 is located at the outlet of the outdoor coil serving as an evaporator to sense the superheat of the leaving refrigerant and regulate the refrigerant mass flow rate circulating through the circuit properly. Similar to the operations in the previous cases, in the AHP system using engine coolant, the liquid refrigerant enters TXV2 located at the inlet of the outdoor coil. Since valve V4 is closed, it flows through valve V8 and enters the heat exchanger serving as an evaporator. After absorbing heat from the engine coolant, the refrigerant evaporates and leaves the heat exchanger as superheated vapour. Then, through valve V7, the refrigerant enters the reversing valve, and the operation goes on similar to the previous cases. THERMODYNAMIC ANALYSIS The performance parameters of the experimental AHP system can be evaluated by applying the first law of thermodynamics to the system. Using this law for the indoor coil (condenser), the heating capacity of the experimental AHP system can be evaluated from  r (hcond,in  hcond,out ) (1) Q cond m Assuming that the compressor is adiabatic, the power absorbed by the refrigerant during the compression process can be expressed as

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Wcomp m r (hcomp,out  hcomp,in )

(2)

Coefficient of performance of the system can be defined as

COP 

Q cond Wcomp

(3)

The general form of exergy rate balance equation given below can be utilized (Ozgener and Hepbasli, 2007).



T0   Q Wcv   m in in   m out out  E xd  j j 

 1  T 

(4)

The specific flow exergy in this equation can be evaluated from (Ozgener and Hepbasli, 2007)

  (h  h0 )  T0 (s  s0 )

(5)

where subscript “0” stands for reference (dead) state. In the adiabatic compressor, the rate of exergy destruction can be evaluated from

 r  comp,in  comp, out  E xd , comp  m

(6)

The rate of exergy destruction in the condenser can be determined from

E xd ,cond  m r ( cond,in  cond,out )  m a,cond ( B  C )

(7)

The total flow exergy of air to be used in Eq. (7) at the locations B and C can be calculated from (Ozgener and Hepbasli, 2007)

 a  (C p ,a  C p ,v )T0 (T / T0 ) - 1 - ln(T / T0 )  (1  6078 )( RaT0 ln( P / P0 ) (1  1.6078 ) ln(1  1.60780 ) /(1  1.6078 )  RaT0    1.6078 ln( / 0 ) 

(8)

Neglecting the heat transfer, the rate of exergy destruction in the expansion valve can be obtained from

 r ( valve, in  valve, out ) E xd ,valve  m

(9)

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Then, using Eq. (5) and considering that

h5  h4 , we obtain

E xd ,valve  m rT0 svalve, in  svalve, out 

(10)

The rate of exergy destruction in the evaporator can be evaluated from

E xd ,evap  m r ( evap,in  evap,out )  m a,evap ( E  F )

(11)

Finally, the total rate of exergy destruction in the refrigeration circuit of the system can be determined by summing up the individual destructions, i.e.

E xd ,total  E xd , cond  E xd , evap  E xd , comp  E xd , valve

(12)

RESULTS AND DISCUSSIONS Fig. 2 shows the variations in the heating capacity versus compressor speed at the fifth minute of the tests when the temperatures of the air streams entering the evaporator and condenser are maintained at 5 ºC. The coolant based AHP system has the highest heating capacity at all compressor speeds. The heating capacity at the end of the fifth minute increases on rising the compressor speed for a fixed dynamometer load of 60 Nm. In all tests performed at 60 Nm, the coolant based AHP system has the highest fifth minute heating capacity, which is followed by the baseline heating system, AHP with exhaust gas and AHP with ambient air in descending order.

Fig. 2. The change of the heating capacity as a function of the compressor speed at the end of the five minute operation period

The effects of the compressor speed on some of the performance parameters of the AHP system at steady state conditions are shown in Figs. 3–6. Fig. 3 shows the variations of the heating capacity versus compressor speed when the temperatures of

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the air streams entering the evaporator and condenser are maintained at 5 ºC and at the fixed dynamometer load of 60 Nm. It is seen that the heating capacity usually gets higher on increasing the compressor speed. In the AHP system with the engine coolant, the refrigerant absorbs a greater amount of heat from the coolant heat exchanger, thereby providing considerably higher heating capacity compared with the AHP system with ambient air or exhaust gas. On the other hand, when the engine load is kept at 60 Nm, the baseline heating system yields higher steady-state heating capacities than all AHP systems at all speeds. In all tests, The AHP system using exhaust gas usually results in higher heating capacities than the AHP system using ambient air as a heat source.

Fig. 3. The change of the heating capacity as a function of the compressor speed

Fig. 4 indicates that the COP for heating decreases by increasing the compressor speed. Because the compressor power increases faster than the heating capacity does, the COP for heating decreases with increasing compressor speed. The AHP system using engine coolant provides the highest COP compared with the AHP system using ambient air and exhaust gas.

Fig. 4. The change of the coefficient of performance for heating as a function of the compressor speed

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The effect of the compressor speed on the distribution of the exergy destructions in the components of the AHP system can be shown in Fig. 5 for the engine speed of 850 rpm at 5 ºC of air inlet temperatures. It is seen that the indoor coil (condenser) usually causes the highest exergy destruction followed by the outdoor coil (evaporator), compressor and thermostatic expansion valve in descending order.

Fig. 5. The comparison of the exergy destructions at n=850 rpm and T=5 Nm

Fig. 6 indicates that the AHP with engine coolant has the highest exergy destruction at 1550 rpm engine speed and 60 Nm dynamometer load settings. As the heat transfer in the outdoor and indoor coils takes place across a higher temperature difference, the exergy destructions in these components are higher than other components.

Fig. 6. The comparison of the exergy destructions at n=1550 rpm and T=60 Nm

Fig. 7 shows the effect of the compressor speed on the rate of total exergy destruction in the AHP system. It can be seen that the rate of total exergy destruction in the AHP system increases on rising the compressor speed. This is mainly due to the increased refrigerant mass flow rate and condensing pressure together with decreased

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evaporating pressure with increasing compressor speed. Furthermore, as the compressor speed increases, the heat transfer in the condenser (indoor coil) and evaporator (outdoor coil) takes place across a higher temperature difference, and the exergy destructions in these components get higher. The exergy destruction rate in the AHP with coolant is higher than those in the AHP with ambient air and with exhaust gas mainly because of the greater refrigerant mass flow rate and heating capacity.

Fig. 7. The change of the total exergy destruction as a function of the compressor speed

CONCLUSIONS The performance of an experimental AHP system for the cases of using ambient air, exhaust gas and engine coolant as a heat source has been evaluated and compared with each other as well as with the performance of the baseline heating system. AHP systems using different heat sources provided higher heating capacities than the baseline heating system at the end of five minute operation period when the engine was operated at the idling conditions (n=850 rpm and T= 5 Nm). However, only the AHP system using engine coolant as a heat source yielded a better fifth-minute performance than the baseline heating system at higher speeds when the engine load was kept at 60 Nm. On the other hand, when the engine load was 60 Nm, the baseline heating system provided higher steady-state heating capacities than all AHP systems at all speeds. In all tests, the AHP system using exhaust gas usually resulted in higher heating capacities than the AHP system using ambient air as a heat source. It was determined that the AHP system with engine coolant yielded the highest COP values, while the AHP system with ambient air yielded the lowest ones. The exergy destructions in the components of the AHP system and the total exergy destruction in the AHP system increased on rising compressor speed and temperatures of the air streams entering the evaporator and condenser. Among the AHP system components, the condenser usually caused the highest exergy destruction, followed by

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evaporator, compressor and expansion valve. These results reveal that the condenser has a high potential for further improvement. Acknowledgement The authors would like to thank The Scientific and Technological Research Council of Turkey (TUBITAK) for supporting this study through a Research Project (Grant No. 108M132). NOMENCLATURE

Tj

AAC automotive air conditioning AHP automotive heat pump COP coefficient of performance c p ,a specific heat of air, kJ/kg K

W

c p ,v

specific heat of water vapour,

E x d

kJ/kg K the rate of exergy destruction, W

R s T

T0

j

power, W power produced in the control

volume, W Greek symbols  humidity ratio  specific flow exergy Subscripts 0 reference (dead) state a air comp compressor cond condenser cv control volume evap evaporator in inlet ind indoor out outlet outd outdoor r refrigerant uni unit

enthalpy, kJ/kg mass flow rate, g/s engine speed, rpm pressure, Pa heating capacity, W

h

m n p Q Q

Wcv

instantaneous temperature, K

time rate of heat transfer, W ideal gas constant, kJ/kg K entropy, kJ/kg K temperature, °C environmental temperature representing the dead state, K

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Direk, M. and M. Hosoz. 2008. Energy and exergy analysis of an automobile heat pump system. International Journal of Exergy 5:556-566. Domitrovic, R.E., V.C. Mei and F.C. Chen. 1997. Simulation of an Automotive Heat Pump. ASHRAE Transactions 103:291-296. Hosoz, M. and M. Direk. 2006 Performance evaluation of an integrated AAC and HP system. Energy Conversion and Management 47:545-559. Kim, S.C., M.S. Kim, I.C.Hwang and T. Lim. 2007. Heating performance enhancement of a CO2 heat pump system recovering stack exhaust thermal energy in fuel cell vehicles. International Journal of Refrigeration 30:1215-1226. Meyer, J., G. Yangand E. Papoulis. 2004. R134a heat pump for improved passenger comfort. SAE paper 2004-01-1379. Ozgener, O. and A. Hepbasli. 2007. Modeling and performance evaluation of ground source (geothermal) heat pump systems. Energy and Buildings 39:66-75. Rongstam, J and F.A. Mingrino. 2003. Coolant – based system. C599/067/2003 VTMS 6 Conference.

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heat

pump

Scherer, L.P., M. Ghodbane, J.A. Baker and P.S. Kadle. 2003. On-Vehicle performance comparison of an R-152a and R-134a HP system. SAE Word Congress, Detroit, Michigan, USA. Tamura, T., Y. Yakumaru and F. Nishiwaki. 2005. Experimental study on automotive cooling and heating air conditioning system using CO2 as a refrigerant. International Journal of Refrigeration 28:1302-1307. Wienbolt, H.W. and C.D. Augenstein. 2003. Visco heater for low consumption vehicles SAE Word Congress, Detroit, Michigan, USA. 2003-01-0738.

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