Performance evaluation of an R134a automotive heat pump system for ...

Applied Thermal Engineering 78 (2015) 419e427

Contents lists available at ScienceDirect

Applied Thermal Engineering journal homepage: www.elsevier.com/locate/apthermeng

Research paper

Performance evaluation of an R134a automotive heat pump system for various heat sources in comparison with baseline heating system Murat Hosoz a, *, Mehmet Direk b, Kadri S. Yigit c, Mustafa Canakci a, Ali Turkcan d, Ertan Alptekin a, Ali Sanli d a

Department of Automotive Engineering, Kocaeli University, Kocaeli 41380, Turkey Department of Energy Systems Engineering, Yalova University, Yalova 77100, Turkey Department of Mechanical Engineering, Kocaeli University, Kocaeli 41380, Turkey d Department of Mechanical Education, Kocaeli University, Kocaeli 41380, Turkey b c

h i g h l i g h t s An R134a automotive heat pump system driven by a diesel engine was developed. The considered heat sources were ambient air, exhaust gas and engine coolant. It was tested by varying engine speed and load at various air temperatures. The system with coolant yielded the highest air temperature at low loads and speeds. The baseline heating system yielded the highest air temperature at other conditions.

a r t i c l e i n f o

a b s t r a c t

Article history: Received 25 September 2014 Accepted 31 December 2014 Available online 8 January 2015

Engine waste heat is commonly used for comfort heating and defogging windows of the passenger compartment of motor vehicles. However, the amount of waste heat to be used for these purposes decreases continuously as a result of increasing engine efficiencies. These requirements can also be met by employing an automotive heat pump (AHP) system. This study presents performance characteristics of an experimental AHP system employing an R134a vapour compression refrigeration circuit and driven by a diesel engine. The AHP system is capable of providing a conditioned air stream by utilizing the heat absorbed from any of the three sources, namely ambient air, engine coolant and exhaust gas. The steadystate and transient performance of the AHP system for each heat source was evaluated by applying energy and exergy analysis to the system based on experimental data. Then, a performance comparison of the AHP system with the three heat sources and baseline heating system was made. It was determined that the AHP system using engine coolant provided the highest indoor coil outlet air temperature during the tests until the steady-state was achieved compared with the AHP system using other heat sources as well as the baseline heating system. © 2015 Elsevier Ltd. All rights reserved.

Keywords: Heat pump Air conditioning Automotive Refrigeration R134a

1. Introduction Comfort heating and defogging windows of the passenger compartment of motor vehicles are commonly achieved utilising waste heat rejected by the engine coolant to an air stream in a heater core. However, sometimes the available waste heat is not enough to achieve a comfortable compartment temperature. The

* Corresponding author. Tel.: þ90 262 3032279; fax: þ90 262 3032203. E-mail address: [email protected] (M. Hosoz). http://dx.doi.org/10.1016/j.applthermaleng.2014.12.072 1359-4311/© 2015 Elsevier Ltd. All rights reserved.

lack of sufficient waste heat is more critical for vehicles employing diesel engines. It has been observed that certain modern high efficiency-direct injection diesel engines cannot produce sufficient waste heat to achieve thermal comfort in an acceptable period of time [1]. In order to obtain thermal comfort in the compartment in a short time after starting up the engine, certain vehicles employ heaters using fuel or electricity. However, these systems have disadvantages in terms of high initial and operating costs, low efficiency, air pollution and global warming. As a remedy to these problems, certain low cost components can be added to the present air conditioning system of the vehicle to operate it as a heat pump

420

M. Hosoz et al. / Applied Thermal Engineering 78 (2015) 419e427

[2]. The automotive heat pump (AHP) system developed in this way can heat the passenger compartment individually, or it can support the present heating system of the vehicle. The literature contains only a limited number of research studies on the AHP systems due to its competitive nature. Vargas and Parise [3] developed a mathematical model to predict the performance of a heat pump system with a variable capacity compressor. Antonijevic and Heckt [4] developed and evaluated the performance of an R134a AHP system, which was employed as a supplementary heating system. They found that the ratio of the heating capacity to the fuel consumption of the developed system was significantly better compared to other supplemental heating systems such as the glow plug heater and PTC heating system. Domitrovic et al. [5] simulated the steady-state cooling and heating operation of an automotive air conditioning (AAC) and AHP system. They found that R134a and R12 yielded comparable results, while the heating capacity of the AHP system was insufficient in both refrigerant cases. Hosoz and Direk [6] evaluated the performance of an air-to-air R134a AHP system, and compared its performance with the performance of the system operated in summer air conditioning mode. Rongstam and Mingrino [7] 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 10 C ambient temperature. Yigit [8] determined experimental performance of an automotive comfort heating system based on recovery of heat from the exhaust gas of vehicles with an air-cooled engine.

Lee et al. [9] conducted a study on the performance characteristics of a mobile heat pump for an electric bus using the wasted heat of electric devices as a heat source. Scherer et al. [10] reported an onvehicle performance comparison of an R152a and R134a AHP system using engine coolant as a heat source. Direk and Hosoz [11] carried out an energy and exergy analysis of an R134a AHP system using ambient air as a heat source. They determined that the heat exchangers of the system were responsible for most of the exergy destruction. Tamura et al. [12] studied the experimental performance of an AHP system using CO2 as a refrigerant. They found that the performance of the system with CO2 was equal to or exceeding that of the system with R134a. Kim et al. [13] 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. Hosoz and Ertunc [14] and Kamar et al. [15] modelled AAC systems using artificial neural networks (ANN) for various compressor speeds, air temperatures and velocities at evaporator inlet and air temperatures at condenser inlet. They determined that the developed ANN models predicted the performance of AAC systems with a high accuracy. As seen in the literature survey outlined above, an R134a AHP system driven by an internal combustion engine and capable of using more than one kind of heating source has not been studied yet. In this study, the performance parameters of an experimental AHP system for the cases of using the heat from ambient air, engine

Fig. 1. Schematic illustration of the experimental AHP system for the cases of using ambient air and exhaust gas as a heat source.


coolant and exhaust gas were evaluated and compared with each other as well as with the performance of the baseline heating system.

Table 1 Specifications of the components of the automotive heat pump system. Component

Specification

Compressor

Type: fixed-capacity swash-plate Stroke volume: 155 cc No. of cylinders: 7 Max. Speed: 6000 rpm Type: parallel-flow micro-channel coil Capacity: 6.44 kW Dimensions: 610 340 17 mm3 No. of channels: 31 Type: laminated coil Capacity: 4.39 kW Dimensions: 230 215 80 mm3 No. of channels: 19 Type: internally equilized with bulb Capacity: 5.27 kW Type: Brazed plate Dimensions: 282 127 37.7 mm3 No. of plates: 14 Heat transfer area: 0.56 m2 Internal volume: 0.33 þ 0.39 dm3 Type: parallel-flow micro-channel coil Dimensions: 235 186 41 mm3 No. of channels: 20 Internal volume: 0.40 dm3

2. Description of the experimental setup The experimental AHP system was generally made from original components of a compact size AAC system, as schematically shown in Fig. 1. It employed a seven-cylinder fixed-capacity swash-plate compressor, a parallel-flow micro-channel outdoor coil, a laminated type indoor coil, two thermostatic expansion valves (TXVs), 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 with dimensions of 400 600 400 mm3 to extract heat from the exhaust gas. Fig. 2 shows a photograph of the experimental AHP system taken before the refrigerant lines and air ducts were insulated. Table 1 indicates the specifications of the components used in the AHP system. 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 contained electric heaters located upstream of the indoor and outdoor coils. The indoor and outdoor coil electric heaters could be controlled between 0e2 kW and 0e6 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. In order to gather data for the performance evaluation of the experimental AHP system, some mechanical measurements were conducted on it. The instruments and their locations are also depicted in Fig. 1. The refrigerant and coolant mass flow rates were measured by Coriolis and electromagnetic mass flow meters, respectively. The temperatures measurement was performed by type K thermocouples and refrigerant pressures were measured by pressure transmitters. Most of the measured variables were acquired through a data acquisition system and recorded on a computer. The characteristics of the instrumentation can be seen in Table 2. The AHP system was driven by a Fiat Doblo JTD diesel engine with a cylinder volume of 1.9 L and a maximum power of 77 kW at

421

Outdoor unit

Indoor unit

TXVs Refrigerant heat exchanger

Heater core

Table 2 Characteristics of the instrumentation. Measured variable Temperature Pressure Humidity Air flow rate Refrigerant mass flow rate Coolant mass flow rate Engine speed Torque

Instrument

Range

Uncertainty

Type K thermocouple Pressure transmitter Hygrometer Anemometer Coriolis flow meter

50 to 500 C 0e25 bar 10e100% 0.1e15 m s1 0e350 kg h1

±% ±% ±% ±% ±%

Electromagnetic flow meter Digital tachometer Hydraulic dynamometer

0e1 m3 h1

±% 0.3

10e100,000 rpm 5e750 Nm

±% 2 ±% 2

0.3 0.2 3 3 0.1

4000 rpm. The engine torque and speed were measured by means of a hydraulic dynamometer with a maximum measuring power of 100 kW. Fig. 1 also illustrates the refrigerant flow paths in the experimental heat pump system for the cases of using ambient air and

Fig. 2. Photograph of the experimental AHP system.

422


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 to the 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 the valve 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 loads. After passing through the filter/drier, sight glass, Coriolis flow meter and V3, the refrigerant reaches TXV2 located at the inlet of the outdoor coil. Since the 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, in which it absorbs heat from the outdoor air stream, and leaves this coil as low pressure superheated vapour. Then, the refrigerant passes through valve V6, and enters the reversing valve. This valve directs the refrigerant to

the compressor, which receives the low pressure refrigerant vapour and compresses it to a high pressure. 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, before absorbing heat in the outdoor coil, the air stream is heated by the exhaust gas in a heat exchanger located upstream of the outdoor coil. Note that in this operation, the bulb of TXV2 was located at the outlet of the outdoor coil to sense the superheat of the refrigerant leaving this coil, thus regulating the refrigerant mass flow rate circulating through the circuit properly. Fig. 3 illustrates the refrigerant flow paths in experimental AHP system using engine coolant as a heat source. Similar to the operations in the previous cases, the liquid refrigerant enters TXV2 located at the inlet of the outdoor coil. Since the valve V4 is closed, it flows through the 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 the valve V7, the refrigerant enters the reversing valve, and the operation goes on as mentioned in the previous cases. Note that, in this case the bulb of TXV2 was located at the outlet of the heat exchanger to regulate the refrigerant mass flow rate properly. Furthermore, the engine coolant-refrigerant heat exchanger was located upstream of the engine oil coiler, and

Fig. 3. Schematic illustration of the experimental AHP system for the case of using engine coolant as a heat source.


423

the coolant stream was sent to the oil cooler after passing through this heat exchanger.

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

3. Thermodynamic analysis

_ comp _ _ r ðex1 ex2 Þ þ W Ex d;komp ¼ m

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

Q_ cond ¼ m_ r hcond;in hcond;out

(1)

where m_ r is the refrigerant mass flow rate. In the tests with the baseline heating system, the heating capacity provided by the heater core can be determined from

Q_ hc ¼ m_ c hhc;in hhc;out

(2)

where m_ c is the mass flow rate of the engine coolant Assuming that the compressor is adiabatic, the power absorbed by the refrigerant during the compression process can be expressed as

_ comp ¼ m_ hcomp;out h W comp;in r

(3)

The power output from the engine, which is called the engine effective power, can be determined from the product of the engine torque and engine speed, i.e.

Pe ¼ T

pn 30

(4)

The mechanical power input to the AHP, which is the sum of the shaft power given to the compressor and the power consumed by indoor unit blower, outdoor unit fan and compressor electromagnetic clutch, can be evaluated using the difference of the engine fuel consumptions between the operations with and without the AHP system. The coefficient of performance of the AHP system is defined as the ratio between heating capacity and power input to the system, i.e.

Q_ COP ¼ cond _ W

(5)

in

_ is the sum of compressor power and power inputs to the where W in indoor unit blower, outdoor unit fan and compressor electromagnetic clutch. The coefficient of performance of the system based on mechanical power input to the AHP can be obtained from

COPm ¼

Q_ cond _m W

1

! X X T0 _ _ cv þ _ m_ in exin m_ out exout ¼ Ex Qj W d Tj

(7)

The specific flow exergy can be evaluated from Ref. [16].

ex ¼ ðh h0 Þ T0 ðs s0 Þ where subscript “0” stands for reference (dead) state.

The rate of exergy destruction in the condenser and liquid line, which stems from the temperature difference between the refrigerant and the air streams, can be evaluated from

_ _ r excond;in excond;out þ m_ a;cond exC exD Ex d;cond ¼ m

(10)

where m_ a;cond is the mass flow rate of the condenser air stream. The total flow exergy of air to be used in Eq. (10) at the locations C and D can be calculated from Ref. [16].

exa ¼ cp;a þ ucp;v T0 ½ðT=T0 Þ 1 lnðT=T0 Þ þ ð1 þ 1:6078uÞðRa T0 lnðp=p0 Þ ð1 þ 1:6078uÞln½ð1 þ 1:6078u0 Þ=ð1 þ 1:6078uÞ þ Ra T0 þ1:6078u lnðu=u0 Þ (11) Exergy destruction in the expansion valve is mainly due to friction. Neglecting heat transfer, the rate of exergy destruction in the expansion valve can be obtained from

_ _ r exvalve;in exvalve;out Ex d;valve ¼ m

(12)

Using Eq. (8) and considering that the enthalpies of the refrigerant at the inlet and outlet of the expansion valve will be the same, we obtain

_ _ r T0 svalve;out svalve;in Ex d;valve ¼ m

(13)

The rate of exergy destruction in the evaporator, which is due to the heat transfer resulting from the temperature difference between the air and refrigerant streams, can be evaluated from

_ _ r exevap;in exevap;out þ m_ a;evap exF exG Ex d;evap ¼ m

(14)

The total flow exergy of air to be used in Eq. (14) at the locations F and G can be calculated from Eq. (11). Finally, the total rate of exergy destruction in the refrigeration cycle of the system is evaluated by summing up the individual destructions, i.e.

_ _ _ _ _ Ex d;total ¼ Exd;comp þ Exd;cond þ Exd;valve þ Exd;evap

(15)

4. Testing procedure

(6)

In order to determine the locations and magnitudes of the thermodynamic inefficiencies, an exergy analysis of the system can be performed. For this aim, the general form of exergy rate balance equation given below can be utilized [16].

X

(9)

(8)

Performance tests were carried out at five different compressor speeds, namely 850, 1200, 1550, 1900 and 2250 rpm. The tests at 850 rpm were performed at the engine (dynamometer) loads of both 5 Nm (idling operation) and 60 Nm, while tests at other speeds were performed only at the engine load of 60 Nm. In all tests the indoor coil air flow rate was adjusted to its maximum (0.147 m3/s), while in the tests using ambient air and exhaust gas as heat sources the outdoor coil air flow rate was also adjusted to its maximum (0.470 m3/s). Before performing the tests, the air temperatures at the inlets of the evaporator (Tevap, ain) and condenser (Tcond, aout) were both fixed to either 0 or 5 C. In the tests using engine coolant as a heat source, the coolant mass flow rate passing through the refrigerant heat exchanger was 63 cm3/s for 850 rpm engine speed and 5 Nm load. This rate increased on rising engine speed and

424


torque, reaching to 175 cm3/s for 2250 rpm and 60 Nm. Data was collected until the steady-state was achieved, which took a time period of 10e30 min. However, when the discharge pressure exceeded 16.3 bar during the test, which was the maximum test pressure of the indoor coil, normally designed as an evaporator but employed as a condenser, the high pressure switch cut out the compressor by deenergising the coil of the compressor electromagnetic clutch due to the burst risk of the indoor coil. Using data acquired in the test operations, the air temperature at the outlet of the indoor coil, heating capacity, COP, COPm and total exergy destruction in the AHP system were evaluated. After the heat pump tests were completed, a heater core was placed at the outlet of the indoor coil and connected to the engine cooling system using rubber hoses in order to provide data for comparing the results of the heat pump operations with those of the baseline heating system. The specifications of the heater core were also reported in Table 1. In the tests of this system, the air temperature at the inlet of the heater core was maintained at 5 C, and the engine was operated at the speeds and loads equal to those in the AHP tests. The coolant mass flow rate passing through the heater was 119 cm3/s for 850 rpm engine speed and 5 Nm load. This rate increased on rising engine speed and torque, reaching to 330 cm3/s for 2250 rpm and 60 Nm. Data was collected until the steady-state was achieved. Finally, the change of the air temperature at the outlet of the heater core as a function of time was determined, and the heating capacity of the system was evaluated. 5. Results and discussion Fig. 4 presents the variations in the air temperatures at the outlet of the indoor coil and heater core versus time for the compressor speed of 850 rpm and engine load of 5 Nm when the temperatures of the air streams entering the evaporator (Tevap, ain) and condenser (Tcond, ain) were maintained at 5 C. In the tests of the AHP, the air temperature at the outlet of the indoor coil increased rapidly within 2 or 3 min of the tests for all heat sources, whereas the air temperature at the outlet of the heater core of the baseline heating system increased moderately during this period of time. The AHP system using engine coolant as a heat source provided the highest air temperature at the outlet of the indoor coil during the entire test duration compared with the

Fig. 4. The change of the conditioned air temperature as a function of time at n ¼ 850 rpm and T ¼ 5 Nm.

AHP systems using ambient air or exhaust gas and the baseline heating system. This is due to the fact that the AHP system with engine coolant experienced the highest evaporating temperatures, thus yielding highest condensing temperatures. Consequently, the air temperature at the outlet of the indoor coil increased rapidly after starting up the compressor. The first five minutes of the test is very important in determining how an AHP system utilizing different heat sources will perform. At the fifth minutes of the tests, the AHP system with engine coolant yielded a conditioned air stream temperature of 36 C, while the AHP system with air and exhaust gas provided about the same temperatures of 28 C, and the baseline heating system resulted in a temperature of only 19 C, which is not acceptable. On the other hand, the air temperature at the outlet of the heater core in the baseline heating system reached 28 C at the end of 22nd minute. Since the AHP system utilizing ambient air or exhaust gas as a heat source did not absorb as much heat as the AHP system with engine coolant did, the baseline heating system surpassed the AHP system with ambient air and exhaust gas towards the end of the test period. Fig. 5 indicates the changes in the air temperatures at the outlet of the indoor coil and heater core versus time for the compressor speed of 1550 rpm and engine load of 60 Nm when the temperatures of the air streams entering the evaporator and condenser were maintained at 5 C. Compared with the previous case, the air temperature at the outlet of the indoor coil increased for all AHP systems when the engine speed and load were increased. In AHP operations with engine coolant and exhaust gas, this increase was mainly due to the elevated coolant and exhaust gas temperatures. On the other hand, increased engine speed caused a greater refrigerant mass flow rate, thus yielding higher heat rejection in the condenser, and consequently higher conditioned air temperatures for all AHP systems. In the tenth minute of the test, the AHP system with engine coolant yielded an air temperature at the outlet of the indoor coil of about 52 C, while the AHP systems with exhaust gas and ambient air resulted in air temperatures of about 37 C and 30 C, respectively. Because increased engine load and speed caused an increase in the engine waste heat, the baseline heating system yielded higher conditioned air temperatures than the AHP system with engine coolant starting from the 12th minute of the tests.

Fig. 5. The change of the conditioned air temperature as a function of time at n ¼ 1550 rpm and T ¼ 60 Nm.


425

system using engine coolant still yielded the highest heating capacity at 850 rpm and 5 Nm conditions. However, due to the increased waste heat and elevated coolant temperature, the baseline heating system exceeded all AHP systems in a time period of maximum 30 min at all speeds when the engine torque was maintained at 60 Nm. The baseline heating system was followed by the AHP systems with engine coolant, exhaust gas and ambient air in decreasing order. Figs. 6 and 7 reveal that AHP systems could provide heating capacities over the baseline heating system when the temperature of the engine coolant was not sufficiently high, i.e. in the first five minutes after the engine was started up or at the operations with low engine torque. In the tests of the AHP system with engine coolant, the heat exchanger inlet and outlet temperatures of the coolant were 48.3 C and 32.9 C, respectively, for 850 rpm engine speed and 5 Nm load. These temperatures increased on rising engine speed and torque, reaching 88.1 C and 79.5 C for 2250 rpm and 60 Nm, respectively. In the tests of the AHP system with ambient air, the air temperature Fig. 6. The change of the heating capacity as a function of the compressor speed at the end of the five minute operation period.

Fig. 6 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 were maintained at 5 C. The coolant based AHP system had the highest heating capacity at all compressor speeds. At 850 rpm and 5 Nm conditions, the heating capacity of the AHP system with engine coolant, ambient air, exhaust gas and that of the baseline heating system were 4.09 kW, 2.15 kW, 2.51 kW and 1.65 kW, respectively. Additionally, the heating capacity at the end of the fifth minute increased on rising the compressor speed at the fixed engine load of 60 Nm. In all tests performed at 60 Nm, the coolant based AHP system had the highest fifth minute heating capacity, which was followed by the baseline heating system, the AHP system with exhaust gas and the AHP system with ambient air in descending order. Fig. 7 reports the changes in the heating capacity versus compressor speed after the steady-state was achieved. The AHP Fig. 8. The change of the steady-state heating capacity as a function of the compressor speed at the engine load of 60 Nm.

Fig. 7. The change of the steady-state heating capacity as a function of the compressor speed.

Fig. 9. The change of the coefficient of performance (COP) as a function of the compressor speed at the engine load of 60 Nm.

426


Fig. 10. The change of the coefficient of performance based on mechanical power input (COPm) as a function of the compressor speed at the engine load of 60 Nm.

Fig. 11. The change of the total exergy destruction as a function the compressor speed at the engine load of 60 Nm.

at the outlet of the outdoor unit changed between 2.3 C and 4.2 C, each corresponding to the engine speeds of 850 rpm and 2250 rpm, respectively. On the other hand, in the tests of the AHP system with exhaust gas, the air temperature at the outlet of the outdoor unit was 1.9 C for 850 rpm engine speed and 5 Nm load, while it reached to 10.5 C for 2250 rpm and 60 Nm. The effect of the compressor speed on some of the steady-state performance parameters of the AHP system is shown in Figs. 8e10. It can be seen in Fig. 8 that the steady-state heating capacity usually got higher on increasing the compressor speed or on decreasing the temperatures of the air streams entering the indoor and outdoor coil. In the AHP system with the engine coolant, the refrigerant absorbed a greater amount of heat from the coolant heat exchanger, thereby providing a considerably higher heating capacity compared with the AHP systems using ambient air and exhaust gas. Fig. 9 indicates that the COP for heating decreased by increasing the compressor speed since the compressor power increased faster than the heating capacity did with increasing speed. Moreover, the COP got higher on decreasing the air temperatures entering the evaporator and condenser. This is due to the increase in heating capacity and decrease in compressor power with decreasing Tevap, ain and Tcond, ain. On the other hand, the AHP system using engine coolant provided the highest COP compared with the AHP systems using ambient air and exhaust gas. Fig. 10 reports the variations in the coefficient of performance of the AHP system based on mechanical power input to the system (COPm) versus compressor speed. It indicates that an increase in the compressor speed caused a decrease in the COPm for the AHP systems using engine coolant and exhaust gas but caused a slight increase in the COPm for the AHP system using ambient air. The AHP system using ambient air experienced considerably low evaporating temperatures at high compressor speeds, which caused frost formation on the surfaces of the outdoor coil. As a result of restricted flow area, the air flow rate passing through the outdoor coil decreased, and the TXV reduced the refrigerant mass flow rate circulating through the circuit. Consequently, the AHP system using ambient air yielded a lower compressor power and slightly higher COP in expense of a lower heating capacity compared with the AHP systems using other heat sources. Fig. 11 indicates 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 increased on

rising the compressor speed. This is mainly due to the increased refrigerant mass flow rate and condensing pressure together with decreased evaporating pressure with increasing compressor speed. Furthermore, as the compressor speed increased, the heat transfer in the condenser and evaporator took place across a higher temperature difference, and the exergy destructions in these components increased. On the other hand, as the air temperature entering the evaporator in the AHP systems with ambient air and exhaust gas increased, the TXV opened up and caused the refrigerant to circulate at a higher rate, thus yielding higher exergy destructions in the components of the AHP system. Moreover, the condensing pressure and accompanying condensing temperature increased on rising the air temperature entering the condenser, thus causing a higher exergy destruction in the condenser. The exergy destruction rate in the AHP system with engine coolant was higher than those in the AHP systems with ambient air and exhaust gas mainly because of the higher refrigerant mass flow rate. 6. Conclusions The transient and steady-state performance of an experimental automotive heat pump system for the cases of using ambient air, exhaust gas and engine coolant as a heat source was evaluated, and compared with each other as well as with the performance of the baseline heating system. It was found that all tested AHP systems yielded higher conditioned air temperatures and heating capacities than the baseline heating system at the end of five minute operation period when the engine was operated at idling conditions (n ¼ 850 rpm and T ¼ 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 yielded 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. Moreover, it was determined that the AHP system with engine coolant yielded the highest COP and COPm values, while the AHP system with ambient air yielded the lowest ones. The AHP system with engine coolant resulted in the highest rates of exergy


destructions, usually followed by the AHP systems with exhaust gas and ambient air in descending order. These results reveal that, compared to the baseline heating system, the use of all considered AHP systems are advantageous in terms of heating capacity both at idling operation and in the first minutes of high speed and torque operations yielding insufficient coolant temperatures. Acknowledgements 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).

cond cv d evap hc in ind m out outd r v

427

condenser control volume destruction evaporator heater core inlet or input indoor coil mechanic outlet outdoor coil refrigerant water vapour

Nomenclature

References

AAC AHP COP cp,a cp,v ex _ Ex

[1] H.W. Wienbolt, C.D. Augenstein, Visco Heater for Low Consumption Vehicles, SAE Word Congress, Detroit, Michigan, USA, 2003. Paper Code 2003-01-0738. [2] J. Meyer, G. Yang, E. Papoulis, R134a Heat Pump for Improved Passenger Comfort, SAE Technical Papers, 2004. Paper Code 2004-01-1379. [3] J.V.C. Vargas, J.A.R. Parise, Simulation in transient regime of a heat pump with closed-loop and oneoff control, Int. J. Refrig. 18 (1995) 235e243. [4] D. Antonijevic, R. Heckt, Heat pump supplemental heating system for motor vehicles, Proc. Inst. Mech. Eng. D. J. Automob. Eng. 218 (2004) 1111e1115. [5] E.R. Domitrovic, V.C. Mei, F.C. Chen, Simulation of an automotive heat pump, ASHRAE Trans. 103 (1997) 291e296. [6] M. Hosoz, M. Direk, Performance evaluation of an integrated automotive air conditioning and heat pump system, Energy Convers. Manag. 47 (2006) 545e559. [7] J. Rongstam, F.A. Mingrino, A coolant-based automotive heat pump system, in: Vehicle Thermal Management Systems Conference (VTMS6) Proceedings, 2003. SAE Paper Code C599/067/2003. [8] K.S. Yigit, Experimental investigation of a comfort heating system for a passenger vehicle with an air-cooled engine, Appl. Therm. Eng. 25 (2005) 2790e2799. [9] D.Y. Lee, C.W. Cho, J.P. Won, Y.C. Park, M.Y. Lee, Performance characteristics of mobile heat pump for a large passenger electric vehicle, Appl. Therm. Eng. 50 (2013) 660e669. [10] L.P. Scherer, M. Ghodbane, J.A. Baker, P.S. Kadle, On-vehicle Performance Comparison of an R-152a and R-134a Heat Pump System, SAE Technical Papers, 2003. Paper code 2003-01-0733. [11] M. Direk, M. Hosoz, Energy and exergy analysis of an automobile heat pump system, Int. J. Exergy 5 (2008) 556e566. [12] T. Tamura, Y. Yakumaru, F. Nishiwaki, Experimental study on automotive cooling and heating air conditioning system using CO2 as a refrigerant, Int. J. Refrig. 28 (2005) 1302e1307. [13] S.C. Kim, M.S. Kim, I.C. Hwang, T.W. Lim, Heating performance enhancement of a CO2 heat pump system recovering stack exhaust thermal energy in fuel cell vehicles, Int. J. Refrig. 30 (2007) 1215e1226. [14] M. Hosoz, H.M. Ertunc, Artificial neural network analysis of an automobile air conditioning system, Energy Convers. Manag. 47 (2006) 1574e1587. [15] H.M. Kamar, R. Ahmad, N.B. Kamsah, A.F.M. Mustafa, Artificial neural networks for automotive air-conditioning systems performance prediction, Appl. Therm. Eng. 50 (2013) 63e70. [16] O. Ozgener, A. Hepbasli, Modelling and performance evaluation of ground source (geothermal) heat pump systems, Energy Build 39 (2007) 66e75.

d

h m_ n p Pe Q_ R s T T0 Tj _ W _ cv W

automotive air conditioning automotive heat pump coefficient of performance specific heat of air (kJ kg1 K1) specific heat of water vapour (kJ kg1 K1) specific flow exergy (kJ kg1) the rate of exergy destruction (W) enthalpy (kJ kg1) mass flow rate (g s1) engine speed (rpm) pressure (Pa) engine effective power (W) time rate of heat transfer (W) ideal gas constant (kJ kg1K1) specific entropy (kJ kg1K1) temperature (K or C) or torque (Nm) environmental temperature representing the dead state (K) instantaneous temperature (K) power (W) power produced in the control volume (W)

Greek symbols humidity ratio

u

Subscripts 0 reference (dead) state a air c coolant comp compressor