Performance Evaluation of an Integrated Automotive

www.elsevier.com/locate/enconman

Energy Conversion and Management 47 (2006) 545–559

Performance evaluation of an integrated automotiveair conditioning and heat pump system

M. Hosoz *, M. Direk

Department of Mechanical Education, Kocaeli University, Umuttepe, 41100 Kocaeli, Turkey

Received 5 November 2004; accepted 18 May 2005Available online 14 July 2005

Abstract

This study deals with the performance characteristics of an R134a automotive air conditioning systemcapable of operating as an air-to-air heat pump using ambient air as a heat source. For this aim, an exper-imental analysis has been performed on a plant made up of original components from an automobile airconditioning system and some extra equipment employed to operate the system in the reverse direction.The system has been tested in the air conditioning and heat pump modes by varying the compressor speedand air temperatures at the inlets of the indoor and outdoor coils. Evaluation of the data gathered in steadystate test runs has shown the effects of the operating conditions on the capacity, coefficient of performance,compressor discharge temperature and the rate of exergy destroyed by each component of the system forboth operation modes. It has been observed that the heat pump operation provides adequate heating onlyin mild weather conditions, and the heating capacity drops sharply with decreasing outdoor temperature.However, compared with the air conditioning operation, the heat pump operation usually yields a highercoefficient of performance and a lower rate of exergy destruction per unit capacity. It is also possible toimprove the heating mode performance of the system by redesigning the indoor coil, using another refrig-erant with a higher heat rejection rate in the condenser and employing a better heat source such as theengine coolant or exhaust gases.� 2005 Elsevier Ltd. All rights reserved.

Keywords: Automotive air conditioning; Heat pump; R134a; COP; Exergy

0196-8904/$ - see front matter � 2005 Elsevier Ltd. All rights reserved.doi:10.1016/j.enconman.2005.05.004

* Corresponding author. Tel.: +90 262 3032279; fax: +90 262 3032203.E-mail address: [email protected] (M. Hosoz).

mailto:[email protected]

Nomenclature

COP coefficient of performance_Ed rate of exergy destruction (W)h specific enthalpy (kJ kg�1)_m mass flow rate (g s�1)n compressor speed (rpm)_Q cooling or heating capacity (W)s specific entropy (kJ kg�1 K�1)T temperature (K)T0 environmental temperature (K)_W compressor power (W)

Subscriptsa airc condensercomp compressore evaporatorIC indoor coilIN inletOC outdoor coilr refrigerantrv reversing valvet totalv expansion valve

546 M. Hosoz, M. Direk / Energy Conversion and Management 47 (2006) 545–559

1. Introduction

Automotive air conditioning (AC) systems usually employ a vapour compression refrigerationcircuit, currently using R134a as the working fluid, to achieve summer thermal comfort in the pas-senger compartment. In winter, on the other hand, after an outdoor air stream has absorbed wasteheat from the engine coolant, it is supplied to the passenger compartment to keep it comfortablywarm. However, it is known that some modern high efficiency, direct injection Diesel enginescannot produce sufficient waste heat in this manner to achieve thermal comfort in an acceptabletime-to-comfort period [1,2]. The vehicles with this type of engine currently employ electric, fuelburning or visco-heaters to supplement the main heating system. These devices, however, are usu-ally inefficient, heavy, expensive and not environmentally friendly [3].An attractive method of providing supplemental heat to the passenger compartment is to re-

verse the direction of the refrigerant flow in an automotive AC system, i.e. to operate it as a heatpump (HP). In this case, after an air stream has absorbed heat from the indoor coil serving as acondenser, it is blown into the passenger compartment to warm it. Besides assisting the main heat-ing system of the vehicles with high efficiency internal combustion engines, particularly during the

M. Hosoz, M. Direk / Energy Conversion and Management 47 (2006) 545–559 547

engine warm up period, the automotive HP can be used in electric cars where no waste energy isavailable for comfort heating.Because automotive AC is a competitive and technology oriented industry, the literature pro-

vides only a limited number of studies concerning the experimental performance of these systems.Jung et al. [4] studied the thermodynamic performance of supplementary/retrofit refrigerant mix-tures for R12 automotive AC systems produced before 1995. Lee and Yoo [5] conducted perfor-mance analyses of the components of an automotive AC system and developed an integratedmodel to simulate the whole system. Ratts and Brown [6] experimentally analysed the effect ofrefrigerant charge level on the performance of an automotive AC system. Al-Rabghi and Niyaz[7] retrofitted an R12 automotive AC system to use R134a and compared the coefficients of per-formance (COPs) for the two refrigerants. Jabardo et al. [8] developed a steady state computersimulation model for an automotive AC system with a variable capacity compressor and investi-gated its validity on an experimental unit. Joudi et al. [9] presented a computer model simulatingthe performance of an ideal automotive AC system working with several refrigerants. Kaynakliand Horuz [10] analysed the experimental performance of an automotive AC system usingR134a in order to find optimum operating conditions. Bhatti [11] investigated potential augmen-tation of the currently used R134a automotive AC system with the aim of lowering its total equiv-alent global warming impact.Since R134a has a global warming potential (GWP) of 1300 times that of CO2, the industry has

been searching for refrigerants with a low GWP as a candidate for automotive AC systems.Hydrocarbon refrigerants cannot be used in these systems due to their potential flammability,and CO2 seems to be a promising refrigerant for this area. Brown et al. [12] compared the perfor-mance characteristics of CO2 and R134a automotive AC systems using simulation models for thevapour compression refrigeration cycle. In spite of possible safety issues, Ghodbane [13] assessedthe use of hydrocarbons as an alternative to R134a and simulated the performance of an automo-tive AC system with R152a and several other hydrocarbons.Although the automotive HP had been used in concept cars before, it was first utilized in

commercially produced electric vehicles in the 1990s [14]. Domitrovic et al. [15] simulated thesteady state cooling and heating operation of an automotive AC/HP system using R12 andR134a and determined the change of the cooling and heating capacities, COP and powerconsumption with ambient temperature at a fixed compressor speed. They found that R134aand R12 yield comparable results while the heating capacity of the system is insufficient. Schereret al. [16] reported an on-vehicle performance comparison of an R152a and R134a HP using en-gine coolant as a heat source. They presented the air temperatures at several locations inside thepassenger compartment as a function of time and found that both refrigerants yield almost iden-tical performances and heating capacities. Meyer et al. [3] retrofitted a production vehicle to incor-porate an R134a HP and determined the change of air temperatures inside the passengercompartment with time. They compared warm up test results with baseline data and found a sig-nificant improvement when the HP provided supplemental heat. Since the CO2 systems use atranscritical refrigeration cycle, they may offer a high heating capacity and COP when the AC sys-tem is operated as a HP. Bullard et al. [17] investigated the experimental performance of a CO2

automotive AC/HP system and evaluated the cooling and heating capacities, indoor coil air dis-charge temperature and COP as a function of ambient temperature for the cooling and heatingmodes.


It is seen from the literature survey outlined above that a comparative and detailed steady stateexperimental analysis of an integrated automotive AC and air-to-air HP system using R134a hasnot been performed yet. This paper aims to present a thorough performance evaluation of such asystem for both operation modes and determine the effects of the compressor speed, air temper-atures at the inlets of the indoor and outdoor coils and the condensing/evaporating temperatureson the performance of the system. This evaluation is based on not only energy but also exergyanalysis, which has been performed to obtain quantitative information on the losses and pinpointthe components causing inefficiency. For this purpose, an experimental automotive AC systemcapable of operating as a HP has been developed. Using data gathered in more than 100 test runs,some performance parameters, namely the cooling and heating capacities, COP, compressor dis-charge temperature and the rate of exergy destruction in each component of the refrigeration cir-cuit of the system, have been evaluated and presented.

2. Description of the experimental setup

The experimental automotive AC/HP system, as shown in Fig. 1, was mainly made from ori-ginal components from the AC system of a compact automobile. The plant employs a vapourcompression refrigeration circuit consisting of a five cylinder swash plate compressor, a parallelflow microchannel outdoor coil, two internally equalized thermostatic expansion valves (TXV),a laminated type indoor coil, two filter/driers, a reversing valve to operate the system as a HPwhen required and some check valves. In order to operate the system in the comfort coolingmode, i.e. as an air conditioner, the reversing valve is de-energized. Then, the refrigerant is drawnfrom the indoor coil and sent to the outdoor coil as shown in Fig. 1 by solid arrows. When therefrigerant passes through the indoor coil serving as an evaporator, it absorbs heat from a blowerdriven air stream, thereby providing a cool indoor air stream. After the compression, the refrig-erant enters the outdoor coil and rejects heat into another air stream driven by a twin fanarrangement.Alternatively, the system can be operated as a HP when the reversing valve is energized. This

results in the outdoor and indoor coils serving as an evaporator and a condenser, respectively. Inthis comfort heating mode, the refrigerant circulates in the direction of the dashed arrows shownin Fig. 1. As the refrigerant passes through the outdoor coil, it absorbs heat from the outdoor airstream, and this heat, along with the heat equivalent of the work of compression, is rejected intothe indoor air stream at the indoor coil. In each operation mode, only one TXV, located upstreamof the active evaporator will function. Each expansion valve is connected to a check valve in par-allel to allow refrigerant flow in the opposite direction when the expansion valve is not in action.The compressor is belt driven by a three phase 4 kW electric motor energized through a fre-

quency inverter. The capacity of an automotive AC system is usually controlled by a thermostatwhich de-energizes an electromagnetic clutch to disengage the compressor shaft from the rotatingpulley when the required compartment air temperature is achieved. However, in order to test thesystem in steady state operation without interruption, the experimental system does not employ athermostat. The indoor blower and outdoor fan motors were energized by separate direct currentpower sources with adjustable output voltages. Because the air velocity at the outdoor coil de-pends on the voltage across the fan motors, varying this voltage allows obtaining a broad range

Fig. 1. Schematic diagram of the experimental automotive air conditioning/heat pump system.


of condensing and evaporating temperatures in the cooling and heating modes, respectively,regardless of the air temperature at the inlet of the outdoor coil.The outdoor coil has a frontal area of 0.227 m2, and it was inserted into an air duct of 1.0 m

length. In order to provide a uniform air flow through the evaporator, a flow straightenerwas used upstream of the coil. The indoor coil has a frontal area of 0.044 m2, and it wasinserted into another duct of 1.0 m length. This duct also contains the blower, electric heaterand another flow straightener located upstream of the indoor coil. The electric heater, whichcan be controlled between 0 and 2000 W, is used to achieve the required air temperature at theindoor coil inlet.The refrigerant lines of the system were made from copper tubing and insulated by elastomeric

material. The duct containing the indoor coil was also insulated by a 5 cm slab of rock wool. Therefrigeration circuit was charged with 700 g of R134a.The locations where the measurements were performed are also depicted in Fig. 1. The refrig-

erant and air temperatures at various points of the system were detected by type K thermocouples.Thermocouples for the refrigerant temperatures were soldered to the copper tube. The air sidetemperature measurements consist of dry and wet bulb temperatures of the air entering and

Table 1Characteristics of the instrumentation

Measured variable Instrument Range Uncertainity

Temperature Type K thermocouple �50/100 �C 0.3 �CPressure Bourdon gauge �100/1000, 0/3000 kPa 10/50 kPaAir flow rate Anemometer 0.1/15 m s�1 ±3%Compressor speed Digital tachometer 10/100000 rpm ±2%


leaving the indoor and outdoor coils. The dry and wet bulb temperatures at the outlet of the in-door coil were determined at three different locations, and they were averaged to find the meanvalues. The suction and discharge pressures were measured by Bourdon tube gauges. It was as-sumed that the evaporating and condensing pressures were equal to the suction and dischargepressures, respectively.In order to determine the air mass flow rates through the indoor and outdoor coils, the air

velocities in the related ducts were measured by an anemometer and were averaged. Then, themean velocities, along with the air densities and duct flow areas were evaluated with the continuityequation. The compressor speed was detected by a digital photoelectric tachometer, while the elec-tric power consumption of the compressor motor was measured by an analogue Wattmeter. Somefeatures of the instrumentation are summarized in Table 1, and further details of the experimentalsetup can be found in Direk [18].

3. Description of the testing procedure

The experimental performance of the automotive AC/HP system was evaluated by conductingtwo groups of tests for each operation mode, namely the maximum outdoor fan speed andthe constant condensing temperature tests. In the first group of tests, a voltage of 12 V, yieldinga volumetric air flow rate of 0.174 m3 s�1, was applied to the outdoor fan motors. After thecompressor speed was adjusted by means of a potentiometer connected to the inverter, the re-quired air temperature at the inlet of the indoor coil was achieved by varying the energy inputto the electric heater. In the second group of tests, performed at constant condensing temperaturesof 45, 50 and 55 �C, the voltage applied to the outdoor fan motors was changed between 0 and12 V to obtain a varying air flow rate. In the cooling mode, a change in the air flow rate throughthe outdoor coil influences the condensing temperature directly, while in the heating mode, it ini-tially affects the evaporating temperature, and then this causes the condensing temperature tovary.In the cooling mode tests, the air dry bulb temperature at the indoor coil inlet was kept at 26, 31

and 37 �C, while the compressor speed was varied between 750 and 2000 rpm. In the heating modetests, on the other hand, the dry bulb temperature of the air entering the indoor coil was kept at13, 18, 24 and 30 �C, while the compressor speed was maintained at 750, 1000 and 1250 rpm. Inboth modes, the minimum compressor speed was chosen as 750 rpm to avoid insufficient lubrica-tion that might arise at lower speeds. In the heating mode tests, low air temperatures at the out-door coil inlet caused quite low evaporating temperatures when the compressor speed exceeded


1250 rpm. This resulted in frosting and activation of the compressor capacity control system, thusimpairing the steady state and yielding inefficient operation. The capacity control system controlsthe refrigerant flow rate by adjusting the stroke of the compressor pistons when the suction pres-sure falls below a predetermined value. In order to avoid this interference, the upper limit of thecompressor speed was chosen as 1250 rpm in the heating mode tests. In the cooling mode tests,however, the air temperatures at the indoor coil inlet were significantly higher, and this allowedselecting a speed limit up to 2000 rpm.In both groups of tests, the volumetric flow rate of the air stream passing over the indoor coil

was fixed at 0.114 m3 s�1. The air temperature at the outdoor coil inlet was maintained at 26 and31 �C in the cooling mode tests and kept at 13 and 18 �C in the heating mode tests. In order toachieve this, a conditioned air stream was supplied to the laboratory when needed. It was acceptedthat when temperature deviations at the key points considered were lower than 0.5 �C for 10 min,the steady state was achieved. The experimental plant was usually brought to steady state within20–40 min after the input conditions were changed. Data were collected to evaluate the perfor-mance of the system as soon as stabilized conditions occurred.

4. Thermodynamic analysis

Referring to Fig. 1, the cooling and heating capacities of the experimental system in the AC andHP operation modes can be expressed in terms of the mass flow rate and the enthalpies of the airstream at the inlet and outlet of the indoor coil. That is,

_Q ¼ _majhB � hCj ð1Þ

Since the air stream passing over the indoor coil exchanges heat only with the refrigerant, therefrigerant mass flow rate can be calculated as

_mr ¼_Q

jh8 � h7jð2Þ

Assuming that the compression process is adiabatic, the compressor power absorbed by therefrigerant in both operation modes can be evaluated as

_W ¼ _mrðh2 � h1Þ ð3Þ
The ratios of the cooling and heating capacities to the compressor power give the energetic per-
formances of the system in the air conditioning and heating modes, respectively. That is,

COP ¼_Q_W

ð4Þ

In the adiabatic compressor, the rate of exergy destruction, which is due to gas friction,mechanical friction of the moving parts, and internal heat transfer, can be expressed as

_Ed;comp ¼ _mrT 0ðs2 � s1Þ ð5Þ
where T0 is the environmental temperature representing the dead state.


Assuming no heat transfer to/from the environment, the rate of exergy destruction in thereversing valve can be computed from Eqs. (6.a) and (6.b) for the cooling and heating modes,respectively.

_Ed;rv ¼ _mrT 0½ðs3 � s2Þ þ ðs1 � s8Þ� ð6:aÞ_Ed;rv ¼ _mrT 0½ðs8 � s2Þ þ ðs1 � s3Þ� ð6:bÞ

Because it is assumed that there is no pressure drop in any component of the refrigeration circuitexcept the expansion valve, the source of exergy destruction in the reversing valve is the irrever-sibilities associated with stream-to-stream heat transfer.The rate of exergy destruction in the condenser and liquid line is due to the heat transfer result-

ing from the temperature difference between the air and refrigerant streams. This rate can bedetermined from Eqs. (7.a) and (7.b) for the cooling and heating modes, respectively.

_Ed;c ¼ _mrT 0 ðs6 � s3Þ �ðh6 � h3Þ

T E

� �ð7:aÞ

_Ed;c ¼ _mrT 0 ðs5 � s8Þ �ðh5 � h8Þ

T B

� �ð7:bÞ

Neglecting the heat transfer, the rate of exergy destruction in the expansion valve, which is pri-marily due to the refrigerant friction accompanying the expansion across the valve, is given byEqs. (8.a) and (8.b) for the cooling and heating modes, respectively.

_Ed;v ¼ _mrT 0ðs7 � s6Þ ð8:aÞ_Ed;v ¼ _mrT 0ðs4 � s5Þ ð8:bÞ

The rate of exergy destruction in the evaporator stems from the temperature difference betweenthe refrigerant and the air stream. This rate can be evaluated from Eqs. (9.a) and (9.b) for thecooling and heating modes, respectively.

_Ed;e ¼ _mrT 0 ðs8 � s7Þ �ðh8 � h7Þ

T B

� �ð9:aÞ

_Ed;e ¼ _mrT 0 ðs3 � s4Þ �ðh3 � h4Þ

T E

� �ð9:bÞ

Finally, the total rate of exergy destruction in the refrigeration circuit of the system in bothoperation modes can be calculated by

_Ed;t ¼ _Ed;comp þ _Ed;rv þ _Ed;c þ _Ed;v þ _Ed;e ð10Þ

5. Results and discussion

The variations in some of the performance parameters of the experimental system with com-pressor speed are shown in Figs. 2–6 for various air temperatures at the inlets of the indoor

500 750 1000 1250 1500 1750 2000 22502400

2800

3200

3600

4000

4400

4800TIC, IN=37 oC, TOC, IN=26 oC

TIC, IN=31 oC, TOC, IN=31 oC


Coo

ling

capa

city

(W)

Compressor speed (rpm)

700 800 900 1000 1100 1200 13002500

3000

3500

4000

4500

5000

5500

a

TOC, IN=18oC, TIC, IN=30oC




Hea

ting

capa

city

(W)

Compressor speed (rpm)b

Fig. 2. Variations in the cooling capacity (a) and heating capacity (b) with compressor speed.

500 750 1000 1250 1500 1750 2000 22502.4

2.8

3.2

3.6

4.0

4.4

4.8

a b




CO

P


700 800 900 1000 1100 1200 13003.0

3.5

4.0

4.5

5.0

5.5

6.0 TOC, IN=18oC, TIC, IN=30oC




CO

P


Fig. 3. Variations in the COP with compressor speed for the cooling mode (a) and heating mode (b) operations.


and outdoor coils. The graphs in these figures have been obtained from the maximum outdoor fanspeed tests.The cooling and heating capacities as a function of compressor speed are plotted in Fig. 2. The

cooling capacity increases with compressor speed while it drops with decreasing air temperature atthe indoor coil inlet. Because of activation of the capacity control system, the cooling capacitiestend to decrease when a certain compressor speed, whose value is dependent upon the tempera-tures of the air streams entering the coils, is exceeded. Similarly, the heating capacity also in-creases with compressor speed and declines on decreasing air temperature at the outdoor coilinlet. It is observed that the experimental system provides a significant amount of heat to the in-door air stream, and the heating capacity is competitive with the cooling capacity for the givenoperating conditions.

500 750 1000 1250 1500 1750 2000 225065

70

75

80

85

90

95

100

a

TIC, IN

=37 oC, TOC, IN

=26 oC




700 800 900 1000 1100 1200 130060

70

80

90

100

110

120

130





Dis

char

ge te

mpe

ratu

re (

”C)

Dis

char

ge te

mpe

ratu

re (

”C)


Fig. 4. Variations in the compressor discharge temperature with compressor speed for the cooling mode (a) and heatingmode (b) operations.

500 750 1000 1250 1500 1750 2000 22500.24

0.28

0.32

0.36

0.40

0.44

a




Ed/

Q


700 800 900 1000 1100 1200 13000.12

0.16

0.20

0.24

0.28

0.32

TOC, IN

=18oC, TIC, IN

=30oC




Ed/

Q


Fig. 5. The rate of total exergy destruction per unit capacity as a function of compressor speed for the cooling (a) andheating (b) modes.


The changes in COP as a function of compressor speed are exhibited in Fig. 3. As can be seen fromthis figure, the COP for the cooling mode declines with increasing compressor speed and decreasingair temperature at the indoor coil inlet. Similarly, the COP for the heating mode also declines withincreasing compressor speed, and it increases with decreasing outdoor temperature. Operation witha higher COP is accomplished at the expense of a lower capacity in both modes, as can be seen byexamining Fig. 2. Because the COP for heating takes into account the heat added to the refrigerantby the compressor, it surpasses the COP for cooling for the given operating conditions.The compressor discharge temperatures are indicated in Fig. 4 as a function of compressor

speed. The discharge temperatures in both operation modes rise on increasing compressor speedand air temperatures at the inlets of the indoor and outdoor coils. This is a result of increasingcondensing temperatures with compressor speed and with air inlet temperatures. It is known thatthe higher is the discharge temperature, the higher is the possibility of thermal destruction of the

600 800 1000 1200 1400 16000

250

500

750

1000

1250

1500

1750

a

TIC, IN =31 oC, TOC, IN=31 oC

reversing valve evaporatorexpansion valve condensercompressor total

Rat

e of

exe

rgy

dest

ruct

ion

(W)


700 800 900 1000 1100 1200 13000

250

500

750

1000

1250

1500

TOC, IN=13 oC, TIC, IN =13 oC

reversing valve evaporatorexpansion valve condensercompressor total

Rat

eof

exer

gyde

stru

ctio

n(W

)


Fig. 6. The rates of exergy destroyed by the refrigeration circuit components as a function of compressor speed for thecooling (a) and heating (b) modes.


lubricating oil, which consequently causes excessive wear and decreases the durability of the com-pressor. It is seen that the discharge temperatures observed in the heating mode for the given mildweather conditions are slightly higher than those experienced in the cooling mode. However, thedischarge temperatures in the heating mode would be lower if more severe weather conditionsprevailed.The ratio of the rate of total exergy destruction to the capacity versus compressor speed is pre-

sented in Fig. 5. For both operation modes, this ratio rises on increasing the compressor speed,while it drops on decreasing the air temperatures at the inlets of both coils. Furthermore, the heat-ing operation yields lower ratios, thus giving higher COPs than those observed in the coolingoperation, as depicted before in Fig. 3.The rates of exergy destroyed by the components of the refrigeration circuit versus compressor

speed are shown in Fig. 6. For both operation modes, the exergy destruction in each componentincreases with compressor speed. This can be explained by the fact that an increase in the com-pressor speed raises the refrigerant flow rate and condensing pressure while dropping the evapo-rating pressure. As the pressure difference across the compressor and expansion valve increases,these components destroy more exergy. On the other hand, the increasing mean temperature dif-ference between the refrigerant and the air due to rising condensing pressure causes higher exergydestruction in the condenser. Similarly, when the evaporating pressure decreases, the mean tem-perature difference between the air and the refrigerant increases, thus raising the exergy destruc-tion in the evaporator. It is seen that for the cooling mode, the contributions of the indoor andoutdoor coils to the total exergy destruction are almost equal, whereas for the heating mode,the exergy destruction in the indoor coil is twice as that in the outdoor coil. This means thatthe indoor coil cannot perform adequately, which is possibly due to its relatively small heat trans-fer area, as a condenser. For the heating mode, the exergy destruction in the outdoor coil is evenlower than that in the compressor, indicating that this coil, originally designed as a condenser,also performs well as an evaporator.The effects of the evaporating and condensing temperatures on some of the performance

parameters are presented in Figs. 7–9. The variations in the cooling and heating capacities with

-2.5 0.0 2.5 5.0 7.5 10.0 12.52000

2400

2800

3200

3600

4000

4400

a

TC=50oC, n=1000 rpm TC=50oC, n=750 rpm



Coo

ling

capa

city

(W)

Evaporating temperature ( oC)

-4 -2 0 2 4 6 82500

2600

2700

2800

2900

3000

3100

3200

3300

TC=45oC, n=750 rpm

TC=50oC, n=750 rpm

TC=55oC, n=750 rpm

Hea

ting

capa

city

(W)

Evaporating temperature ( oC)b

Fig. 7. Variations in the cooling capacity (a) and heating capacity (b) with evaporating temperature.

-2.5 0.0 2.5 5.0 7.5 10.0 12.52.0

2.5

3.0

3.5

4.0

4.5

a




CO

P

Evaporating temperature ( oC)

-4 -2 0 2 4 6 83.6

4.0

4.4

4.8

5.2

5.6

TC=45oC, n=750 rpm

TC=50oC, n=750 rpm

TC=55oC, n=750 rpm

CO

P


Fig. 8. Variations in the COP with evaporating temperature for the cooling mode (a) and heating mode (b) operations.


evaporating temperature are reported in Fig. 7. It is seen that an increase in evaporating tem-perature or a decrease in condensing temperature results in increased capacity for both opera-tion modes. Furthermore, the cooling and heating capacities rise with increasing compressorspeed. At 50 �C condensing temperature and 750 rpm, the ratio of the heating to coolingcapacities ranges from 1.10 ( at 1 �C evaporating temperature) to 1.11 (at 6.5 �C evaporatingtemperature).The COPs for both operation modes as a function of evaporating temperature are indicated in

Fig. 8. In both cases, the COP rises with increasing evaporating temperature and decreasing con-densing temperature. This means that the required compressor power to provide a certain coolingcapacity drops on increasing the air temperature at the indoor coil inlet and on decreasing the airtemperature at the outdoor coil inlet. Similarly, the compressor power to provide a certain heatingcapacity drops with increasing air temperature at the outdoor coil inlet and decreasing air

-2.5 0.0 2.5 5.0 7.5 10.0 12.575

80

85

90

95

100

105

110

115

a




Dis

char

gete

mpe

ratu

re(o C

)

Evaporating temperature ( oC)-4 -2 0 2 4 6 8

70

75

80

85

90

95

100

105

TC=45oC, n=750 rpm

TC=50oC, n=750 rpm

TC=55oC, n=750 rpm

Dis

char

gete

mpe

ratu

re(o C

)


Fig. 9. Variations in the compressor discharge temperature with evaporating temperature for the cooling mode (a) andheating mode (b) operations.


temperature at the indoor coil inlet. At 50 �C condensing temperature and 750 rpm, the COP forheating surpasses the COP for cooling by 37–69% in the evaporating temperature range of 1–6.5 �C.The changes in the compressor discharge temperature with evaporating temperature are shown

in Fig. 9. The discharge temperature reduces with increasing evaporating temperature anddecreasing condensing temperature. At 50 �C condensing temperature and 750 rpm, the coolingmode discharge temperatures are about 5 �C higher than the heating mode ones in the evaporat-ing temperature range of 1–6.5 �C.The AC mode results presented above are usually in good agreement with those given by Joudi

et al. [9], Kaynakli and Horuz [10] and Domitrovic et al. [15], while the HP mode results agree wellwith those given by Domitrovic et al. [15].

6. Conclusions

The performance characteristics of an integrated automotive AC and air-to-air HP systemusing R134a as the working fluid have been experimentally evaluated. Based on the experimentalevidence, the final conclusions reached in this study can be summarized as follows.

• Although the HP operation provides sufficient amounts of heat to the indoor air stream at mildweather conditions, the heating capacity would drop at more severe conditions due to bothdecreasing evaporating temperatures and activation of the capacity control system. Therefore,an air-to-air automotive HP must be considered only as a supplementary heating method to beused in energy efficient automobiles lacking waste heat.

• Both the heating and cooling capacities of the system increase with compressor speed, while theCOPs for both cases decrease with it. Furthermore, the COPs for heating outperform the COPsfor cooling due to the fact that the former takes into account the heat equivalent of the work ofcompression.


• For the same compressor speed and condensing/evaporating temperatures, the HP operationyields lower compressor discharge temperatures.

• In both operation modes, the ratio of the rate of total exergy destruction in the refrigerationcircuit to the capacity increases with compressor speed, while the heating mode operationresults in lower ratios.

• In the cooling mode, both coils destroy almost equal amounts of exergy, whereas in the heatingmode, the exergy destroyed by the indoor coil is double that by the outdoor coil, meaning thatthe indoor coil cannot perform as a condenser adequately. The poor heat rejection in the indoorcoil also limits the amount of heat absorbed by the outdoor coil. Therefore, an automotive AC/HP system must employ an indoor coil with a larger heat transfer area and a higher air flowrate.

Although the experimental system consists of components originally designed for the AC oper-ation, it offers a comparable performance in the HP operation as well. The heating mode perfor-mance will further increase if a better heat source such as engine coolant or exhaust gases isutilized. Since the provision of supplemental heating by operating an existing AC system in thereverse direction requires only a few low cost components with negligible packaging and a reason-able amount of energy compared to its alternatives, this method may be an optimum solution tothe problem of insufficient comfort heating in energy efficient automobiles.

Acknowledgement

The authors would like to acknowledge the support provided by Kocaeli University under theproject number 2002/37.

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Performance Evaluation of an Integrated Automotive

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