- Email: [email protected]
Heating performance evaluation of a CO2 heat pump system for an electrical vehicle at cold ambient temperatures
Applied Thermal Engineering 142 (2018) 656–664
Contents lists available at ScienceDirect
Applied Thermal Engineering journal homepage: www.elsevier.com/locate/apthermeng
Research Paper
Heating performance evaluation of a CO2 heat pump system for an electrical vehicle at cold ambient temperatures Dandong Wanga, Binbin Yua, Wanyong Lia, Junye Shia,b, Jiangping Chena,b, a b
T
⁎
Institute of Refrigeration and Cryogenics, Shanghai Jiaotong University, Shanghai, China Shanghai High Efficient Cooling System Research Center, Shanghai, China
H I GH L IG H T S
CO heat pump system with Series Gas Cooler configuration was proposed. • AA novel developed mobile CO compressor was applied to EV CO heat pump system. • Thenewly of charge, gas cooler and return air ratio were investigated. • The impacts • CO heat pump achieved great heating capacity and COP at −20 °C temperature. 2
2
2
2
A R T I C LE I N FO
A B S T R A C T
Keywords: Electrical vehicle Heat pump CO2 refrigerant Trans-critical Gas cooler
A novel CO2 heat pump system with Series Gas Cooler (SGC) configuration was proposed for application in an Electrical Vehicle (EV). A newly developed mobile CO2 compressor with 6.8 ml displacement and 8000 RPM maximum speed was employed. The heating performance of this CO2 heat pump system was experimentally evaluated under low temperature conditions. Comparison experiments were conducted to reveal the effects of CO2 refrigerant charge, GC configuration and indoor return air ratio on heating performance characteristics. Results show that using SGC improves the heating capacity and COP by up to 33% and 32%, respectively, compared with One Gas Cooler (OGC). At the extreme cold temperature of −20 °C, the novel CO2 heat pump system achieved 5.6 kW maximum heating capacity and 1.8 COP, which greatly outperforms the conventional PTC heating supplement for an EV.
1. Introduction In recent years, mobile air source heat pump systems are becoming increasingly popular in the application of Electrical Vehicles (EV) due to the benefits of improved heating efficiency for cabin heating [1]. However, a mobile heat pump system with conventional refrigerant (HFC-134a or HFO-1234yf) suffers significantly diminishment of heating capacity and system efficiency as the ambient temperature decreases. To face the challenge of the performance attenuation in winter, there are two promising options. One is to develop a high efficient vapor injection heat pump system using conventional refrigerant, the other is to develop a mobile heat pump system using natural refrigerant CO2. Qin et al. [2] have tested an R134a vapor injection heat pump system for an EV under −20 °C ambient conditions and investigated the effect of injection portholes shapes on heating performance. It is found that the best system achieved 3.5 kW heating capacity at compressor ⁎
speed of 8500 RPM. When the inlet air temperature is lower than −10 °C, the performance of the three interlinked portholes system is similar as that of single porthole system. Higuchi et al. [3] provided a novel flash tank vapor injection heat pump system using R1234yf refrigerant, which is installed in 2017′ Toyota Prius. This heat pump system can achieve multi-functions by regulating several valves. And it can achieve a 55 °C air supply temperature under 5 °C ambient condition in heating defogging mode, with a heating COP of 2.7. Few data in heating mode was released in this paper. Jung et al. [4] developed a simulation model for an R134a vapor injection heat pump system and numerically analyzed the effect of injection port geometries. The optimum angle of the single-injection port is determined to be 440°. Although the vapor injection heat pump system significantly improves the heating capacity at low temperatures, the enhanced heating capacity is still insufficient for cabin heating under high heating loads conditions (such as 100% fresh air and −20 °C ambient temperature), thus a PTC with a small capacity of 2–3 kW must be equipped to provide a heat
Corresponding author at: Institute of Refrigeration and Cryogenics, Shanghai Jiaotong University, Shanghai, China. E-mail address: [email protected] (J. Chen).
https://doi.org/10.1016/j.applthermaleng.2018.07.062 Received 18 March 2018; Received in revised form 11 July 2018; Accepted 12 July 2018 1359-4311/ © 2018 Elsevier Ltd. All rights reserved.
Applied Thermal Engineering 142 (2018) 656–664
D. Wang et al.
Nomenclature
p PTC Q R RPM SGC T W*H*D W X
Roman AC COP Cp DSH DSC EV EXV GC h HVAC HX IHX m OGC
air conditioning coefficient of performance specific heat (J kg−1 K−1) degree of superheat (K) degree of sub-cooling (K) electrical vehicle electrical expansion valve gas cooler enthalpy (kJ kg−1) heating, ventilation and air conditioning heat exchanger internal heat exchanger mass flow rate (kg s−1) one gas cooler
pressure (MPa) positive temperature coefficient heating capacity (kW) result revolutions per minute series gas cooler temperature (K) Width * Height * Depth compressor input power (kW) measurement
Subscripts a i g o r
air inlet gas cooler outlet refrigerant
for conventional vehicles and concluded that its dehumidifying efficiency performance exceeds that of the HFC-134a system. Hammer and Audi [14] experimentally compared the vehicle warming-up time between a system with a CO2 heat pump system and with a conventional PTC heater for an Audi car with a 1.6 L gasoline engine. Kim et al. [15] found that use of a CO2 heat pump system with exhaust heat recovery for fuel cell vehicles could improve heating capacity by 100%, compared to conventional PTC heating. In conclusion, further research is needed to fully evaluate the heating performance characteristics of mobile CO2 heat pump system for EVs. When the CO2 heat pump system is operated in trans-critical cycle, the performance of gas cooler (GC) has been identified as a critical component as it affects the heat rejection pressure, refrigerant exit temperature and secondary fluid exit temperature. Therefore, the whole system performance relies heavily on the design of the GC. In addition, cabin return air utilization is a promising approach for EVs to reduce heating load and save electricity power [16]. The effect of this strategy on CO2 heat pump performance remains to be determined. An EV CO2 heat pump system has been designed and experimentally investigated in previous research [17], which concluded that the heat pump system using CO2 refrigerant exhibited good heating performance in a cold climate. However, the achieved heating capacity of 3.6 kW at −20 °C condition is insufficient for passenger cabin heating. In order to
supplement when using vapor injection heat pump system for an EV. Natural refrigerant CO2 offers no toxicity, no flammability, high volumetric capacity, and high heat transfer properties [5,6]. Based on the advantage of thermo-physical properties, CO2 refrigerant has been successfully applied in tap water heat pumps [7] and supermarket refrigeration systems [8]. As a countermeasure against global warming, regulations in the EU, Japan and the US are phasing out the use of HFC134a (GWP = 1430) in mobile air conditioning (AC) systems [9]. Natural refrigerant CO2 (GWP = 1) is considered as a promising alternative to HFC-134a. The European automotive manufacture DaimlerBenz has announced its first-ever plan to offer a CO2 AC system in a production vehicle [10]. Several researches have studied the cooling performance of mobile CO2 AC systems. Brown et al. [11] evaluated the performance merits of CO2 and R134a automotive AC systems. The result from semi-theoretical cycle models show that the COP of CO2 was lower by 21% at 32.2 °C and by 34% at 48.9 °C. The COP disparity was even greater at higher speeds and ambient temperatures. Kim et al. [12] studied the effects of operating parameters on the cooling performance of a CO2 automotive AC system and concluded that the CO2 refrigerant system exhibited good performance. However, there are only a few data available on mobile CO2 heat pump systems, especially for cold climate operation. Tamura et al. [13] developed a prototype CO2 automotive cooling and heating AC system
Fig. 1. Schematic diagram of systems with (a) conventional OGC and (b) novel SGC. 657
Applied Thermal Engineering 142 (2018) 656–664
D. Wang et al.
buffer. Both the OGC and SGC heat pump system can be switched from cooling to heating mode operation via stop valves and EXVs. There is no cycle difference between these two systems at cooling mode. For OGC at heating mode, after being discharged from compressor, the refrigerant enters into the indoor HX1 (No.2 in Fig. 1a) and rejects heat to cold air, then directly passes through the EXV to the low pressure side of the cycle. The indoor HX2 (No.3 in Fig. 1a) is bypassed when all refrigerant flows from the outdoor HX (No.6 in Fig. 1a) to the accumulator. For SGC, however, after the heat rejection in the indoor HX1 (No.2 in Fig. 1b), refrigerant first passes through the full open EXV (NO.8–1 in Fig. 1b), which doesn’t produce pressure drop here, then enters into the indoor HX2 (No.3 in Fig. 1b) to release extra heat to cold inlet air. Thus the indoor HX2 (No.3 in Fig. 1b) is utilized as a second GC in the SGC system, and the cold inlet air is capable of absorbing the rejected heat from indoor HX1 (No.2 in Fig. 1b) and indoor HX2 (No.3 in Fig. 1b). After the heat rejection process in first GC and second GC, the refrigerant passes through the EXV (8–2 in Fig. 1b) and is throttled to low pressure state. After throttling, the two-phase refrigerant evaporates in outdoor HX (No.6 in Fig. 1b), which is used as an evaporator at heating mode. Then the refrigerant enters into the accumulator and goes back to the compressor. When air dehumidification is required to avoid windshield flash fogging, the EXV (NO.8–1 in Fig. 1b) can be regulated to throttle the refrigerant after indoor HX1 (No.2 in Fig. 1b), then the indoor HX2 (No.3 in Fig. 1b) is used as an evaporator and dehumidifies the inlet air. Figs. 2 and 3 show the p-h diagrams of CO2 cycles and the psychrometric chart of indoor air for two CO2 heat pump systems, respectively. As for the IHX, at cooling mode, the refrigerant after outdoor HX (GC, No.6 in Fig. 1) passes through the IHX, undergoing a heat transfer process to the low pressure refrigerant flowed from accumulator, which decreases the inlet quality of indoor HX (Evaporator, No.3 in Fig. 1) and thereby improves the cooling performance of system. At heating mode, both sides of the IHX are low pressure and low temperature refrigerants, so the heat transfer between the two sides of the IHX is negligible and the IHX can be regarded as not working. If IHX was able to work at both cooling mode and heating mode, more switching stop valves will be required and the refrigerant pipes need be extended, thus the working of IHX at heating mode is not considered in this study. The novel SGC system enhances the total heat rejection area of GC by utilizing the indoor HX2 (No.3 in Fig. 1b). One simple choice to achieve the same result of novel SGC in conventional OGC system is to directly enlarge the indoor HX1 (No.2 in Fig. 1a), but this will lead to the increase of air side pressure drop and HX installation space. In addition, as the indoor HX1 (No.2 in Fig. 1a) is installed inside the HVAC module (No.4 in Fig. 1) of the vehicle, the space increase of the
further improve the heating capacity of CO2 mobile heat pump system, this paper put forward a novel CO2 heat pump system using two indoor heat exchangers (HXs) as heat rejection GCs. The total GC heat transfer area can be increased without the change of HVAC module structure. In addition, the CO2 compressor used in previous study [17] was originally designed for domestic CO2 heat pump water heater application. The maximum allowable compressor discharge pressure and temperature were 12 MPa and 120 °C, respectively, which limit the heating performance of mobile CO2 heat pump system. A newly developed mobile rotary type CO2 compressor was adopted in this study. It can be operated up to 8000 RPM speed and 150 °C discharge temperature. With this compressor, comparison experiments were conducted to reveal the impacts of refrigerant charge, GC configuration and indoor return air ratio on CO2 heating performance at the ambient temperatures of −10 °C and −20 °C. The cycle characteristics and performance advantages of the novel CO2 heat pump system were further analyzed and discussed based on the experimental results.
2. Experimental setup 2.1. Experimental apparatus The novel EV CO2 heat pump system is provided considering the HX arrangements. Fig. 1 shows schematic diagram of systems with conventional OGC and novel SGC. Each system is composed of an electrical compressor, an outdoor HX, an indoor HX1, an indoor HX2, an internal heat exchanger (IHX), an accumulator, several stop valves and expansion valves. The CO2 mobile electrical compressor, integrated with a motor driver, is a newly developed rotary type compressor. It is driven by 300–400 V DC electricity power and uses PAG lubricant oil. It has a displacement of 6.8 ml and a weight of 6.7 kg. Besides, it can achieve the speed of up to 8000 RPM and the discharge temperature of up to 150 °C. The CO2 HXs are micro-channel type made of aluminum with high operating pressure resistance of 14 MPa. The sizes of indoor HX1 (No.2 in Fig. 1), indoor HX2 (No.3 in Fig. 1), and outdoor HX (No.6 in Fig. 1) are 180W * 190H * 32D (mm), 230W * 250H * 32D (mm), and 600W * 450H * 12D (mm), respectively. Their dimensions fit with the specifications for a commercial B-class vehicle. The expansion valve used here is a CO2 EXV integrated with a step motor driver, connected to the controller by Lin (Local Interconnect Network) communication. On the controller interface, the input 0–5 DC voltage could be manually adjusted to change the steps of the motor and then regulate the CO2 mass flow rate. This EXV is developed for mobile CO2 heat pump system application. It has throttling function, allows full flow in one direction, and can be closed bi-directionally. An accumulator 1.0 L in volume is installed at compressor suction to serve as a system refrigerant charge
Fig. 2. p-h diagrams of CO2 cycle for (a) OGC system and (b) SGC system under heating operation. 658
Applied Thermal Engineering 142 (2018) 656–664
D. Wang et al.
Fig. 3. Psychrometric charts of indoor air for (a) OGC system and (b) SGC system under heating operation.
indoor HX1 will cause the structure redesign and volume increase of whole HVAC module. Therefore, the proposed SGC system has the advantage of avoiding the change of indoor HX1 dimensions, air side pressure drop and HVAC module structure. But the drawback is that one stop valve is added and the refrigerant lines are extended, which can be seen from Fig. 1. The heating performance of the OGC and SGC system were tested in an automobile heat pump calorimeter facility, whose schematic diagram is shown in Fig. 4. The detailed facility introduction can be referred in a previous research paper [17]. It consisted of an outdoor chamber and an indoor chamber with different open wind tunnels. The outdoor GC and indoor HVAC module were installed before the outdoor wind tunnel and indoor wind tunnel, respectively.
Table 1 Uncertainties of the experimental parameters and measured data. Items
Uncertainties
Temperature sensors (RTD-type, Yokogawa) Pressure transducers (GE-Druck)
± 0.2 °C
Mass flow rate (Coriolis type, KROHNE) Digital power meter (WT210, Yokogawa) Data logger (34972A, Agilent) Heating capacity Heating COP
2.2. Measurement uncertainties Measurement devices were equipped to measure refrigerant and air
Fig. 4. Schematic diagram of calorimeter facility to measure CO2 heating performance. 659
± 0.5% of full scale, Max 10 MPa or 20 MPa, ± 0.15% of full scale, Max 600 kg h−1 ± 0.5% reading 0.004% dcV accuracy 5.5% 6.3%
Applied Thermal Engineering 142 (2018) 656–664
D. Wang et al.
different refrigerant charge conditions. The reduction of DSH at the compressor suction slightly enhanced refrigerant density and refrigerant mass flow rate, which accounted for the impact of DSH on heating capacity. Meanwhile, the DSH had small impact on COP, which was less than 6%. Based on the effects of DSH on heating performance, the DSH was controlled at 0 K by EXV adjustment to achieve a relatively greater heating capacity in later experiments. In addition, taking the maximum heating capacity under 0 K DSH condition as a comparison, the heating capacity increased by 16% from 4.5 kW to 5.2 kW, and the COP increased by 9% from 1.88 to 2.05, when refrigerant charge changed from 0.6 kg to 0.9 kg. It could be seen from Fig. 7c that adding refrigerant charge increased the discharge pressure and the enthalpy difference of GC, which leads to the increase of heating capacity. In addition, the compression pressure ratio and the enthalpy difference of compressor increased with the increase of refrigerant charge, which explains that the increase of COP is smaller than heating capacity. Besides, the results of air outlet temperature of GC (Tg,a,o) and EXV opening are shown in Fig. 7b. The decrease of DSH from 10 K to 0 K improved the Tg,a,o by 2–3 K, and the increase of refrigerant charge from 0.6 kg to 0.9 kg enhanced the highest Tg,a,o by 7.4 K from 37.2 °C to 44.6 °C. Thus the bigger refrigerant charge and smaller DSH control were beneficial to Tg,a,o and heating capacity improvements in this operation condition. Fig. 7b also shows that the EXV opening was decreased with the increase of refrigerant charge or DSH. It was noticed that the EXV opening changed from maximum 73% to minimum 14%, whose variation was related to the change of CO2 inlet pressure, inlet temperature, outlet pressure and refrigerant mass according to mass flow characteristics of CO2 EXV [19,20]. This phenomenon of large EXV opening variation in heating operation requires special consideration in future development and design of the system control. What’s more, the p-h diagram in Fig. 7c shows that the system discharge pressure became higher with increased refrigerant amount adding into the system. At first, when the refrigerant charge is 0.6 kg, the system exhibited CO2 sub-critical cycle, and the refrigerant outlet temperature of GC (Tg,r,o) was 24 °C with 1 K sub-cooling degree (DSC). Then, when refrigerant charge increased to 0.9 kg, the system cycle transformed to CO2 trans-critical cycle, Tg,r,o was lowered to 5 °C. Apart from the reduction of Tg,r,o, the enthalpy difference of GC increased by 107kJ kg-1 with the change of refrigerant charge. As refrigerant charge increased from 0.6 kg to 0.9 kg, the pressure drop of high pressure side GCs decreased from 0.68 MPa to 0.20 MPa, and the pressure drop of low pressure side evaporator and accumulator decreased from 0.66 MPa to 0.24 MPa. The reduction of refrigerant mass flow from 73.1 kg h−1 to 51.8 kg h−1 could account for the reduction of pressure drop. When the refrigerant charge was 0.9 kg, the
conditions. Table 1 shows the uncertainties of the experimental parameters and measured data. The location of sensors on the refrigerant side and air side are shown in Figs. 1 and 4. The heating capacity (Q) of the tested CO2 heat pump was determined by the air side heat transfer rate using Eq. (1). The overall system COP was determined by Eq. (2). The refrigerant side heat transfer rate was also calculated as a comparison by Eq. (3). The properties of the CO2 refrigerant were calculated according to the NIST REFPROP 8.0. Fig. 5 shows that the errors between the heat transfer rate for the air side and the refrigerant side were within ± 5%, which indicated tested system has a good heat balance between air side and refrigerant side.
Qa = ma Cp (Ta, out −Ta, in ) = Cp (Ta, out −Ta, in ) ρa v a A COP =
(1)
Q Q = W U ∗I
(2)
Qr = mr (hg, in−hg, out )
(3)
In order to verify the measured data of the heating capacity and COP, a single-sample uncertainty analysis was performed according to Moffat [18], using the equations presented in Eqs. (4) and (5). Using this analysis method, the overall relative uncertainty of the heating capacity and COP were calculated as 5.5% and 6.3%, respectively. 1/2
δCp 2 δρ δQ δ ΔTa 2 δv δA = ⎜⎛( ) +( ) + ( a )2 + ( a )2 + ( )2⎟⎞ Q C Δ T ρ v A ⎠ p a a a ⎝
(4)
1/2
δCOP δQ δU 2 δI ) + ( ) 2⎞ = ⎛( )2 + ( COP Q U I ⎠ ⎝ ⎜
⎟
(5)
2.3. Test method The impact of refrigerant charge on trans-critical cycle characteristics of EV CO2 heat pump system was evaluated in system experiments at −10 °C ambient temperature, and the degree of superheat (DSH) at compressor suction was varied to determine its effect on heating performance. Next, with the determined refrigerant charge, the heating performance of the OGC and SGC system were experimentally compared at the ambient temperatures of −10 °C and −20 °C and 4000/ 6000/8000 RPM compressor speeds. Besides, the effect of indoor return air utilization on the system performance was investigated with various return air ratios from 0% to 40%. Table 2 shows the details of the operation conditions of system experiments. 3. Results and discussion 3.1. Effect of CO2 refrigerant charge System experiments were first conducted with refrigerant charge varying from 0.6 kg to 0.9 kg at the ambient temperature of −10 °C. Unlike optimum high pressure control in cooling mode, here EXV opening was manually regulated to obtain various refrigerant DSHs at the compressor suction, including 0 K (saturation point), 5 K, and 10 K. The pictures in Fig. 6 shows the refrigerant flow behavior with different DSHs, which were captured by the camera in front of the sight glass. When DSH was higher than 5 K (Fig. 6b and c), the refrigerant flow was sequent steady flow, which represents the superheated vapor state of the refrigerant. When DSH was 0 K (Fig. 6a), intermittent liquid-gas spray flow occurred, representing the two-phase state of the refrigerant. Since the measured small DSH (less than 2 K) can’t be used to judge the saturation point between the two-phase and the superheated vapor state of the refrigerant, the appearance of spray flow was utilized to determine this 0 K DSH saturation condition during experiments. Fig. 7 shows the experimental results of heating performance with different refrigerant charges and DSHs. From Fig. 7a, with the decrease of DSH from 10 K to 0 K, the heating capacity increased by 4–7% under
Fig. 5. Comparison of heat transfer rate between refrigerant side and air side. 660
Applied Thermal Engineering 142 (2018) 656–664
D. Wang et al.
Table 2 Operation conditions of system experiments.
1 2 3
Refrigerant charge (kg)
Indoor air inlet temperature (°C)
Indoor air inlet mass flow (kg h−1)
Outdoor air inlet temperature (°C)
Outdoor air inlet velocity (m s−1)
Compressor speed (RPM)
0.6, 0.7, 0.8, 0.9 0.6, 0.9 0.9
−10 −10, −20 −20, −16, −12, −8, −4
337 337 337
−10 −10, −20 −20
4 4 4
4000 4000, 6000, 8000 8000, 7200, 6500, 5800, 5200
temperature from exceeding limitation, this CO2 SGC heat pump system has not been tested with refrigerant charge more than 0.9 kg. 0.9 kg was considered as the largest refrigerant charge at the ambient temperature of −10 °C. The condition of 0 K DSH was taken to consider the optimum refrigerant charge. The heating capacity and COP increased with the adding of refrigerant charge. At 0.9 kg, this system achieved a highest heating capacity of 5.2 kW and a best COP of 2.05. If more refrigerant are added into this system, the discharge temperature will exceed the limitation of 150 °C, which can be reduced by increasing the opening of EXV. However, some gas-liquid refrigerant will flow into compressor, which is harmful to the compressor reliability, and the COP of system will be decreased because of the increase in the power consumption of compressor. Therefore, 0.9 kg was then chose as the optimum system refrigerant charge.
Fig. 6. Refrigerant flow behavior at compressor suction with (a) 0 K DSH, (b) 5 K DSH and (c) 10 K DSH.
two phase equilibrium temperature was −22.2 °C at evaporator inlet, then it was dropped to −26.7 °C at compressor suction due to the pressure drop of 0.24 MPa. The huge pressure loss and equilibrium temperature reduction in low pressure side will result in the reduction of the suction mass flow of compressor, the heating capacity and the COP of system, which needs to be optimized in the future research. In addition, for 0.9 kg refrigerant charge, the discharge temperature increased to 145 °C, close to discharge limitation 150 °C, which is caused by the low compressor suction pressure of 1.9 MPa and the high discharge pressure of 9.0 MPa. In order to prevent the discharge
3.2. Effect of GC configuration In this study, the SGC CO2 heat pump system was put forward for EV, whose performance was experimentally compared with the OGC system at compressor speeds of 4000, 6000 and 8000 RPM. Two typical
Fig. 7. Heating performance of EV CO2 heat pump with different refrigerant charges (a) Tg,a,o and EXV opening, (b) heating capacity and COP, (c) p-h diagram. 661
Applied Thermal Engineering 142 (2018) 656–664
D. Wang et al.
amount and thus boosts the heating performance of EV CO2 heat pump system. Fig. 8a and b also shows the effect of compressor speed on heating performance. For example, at −20 °C for SGC system, the heating capacity and compressor input power increased by 74% and 185%, and the corresponding COP decreased by 41% as the compressor speed increased from 4000 RPM to 8000 PRM. The increase of compressor speed could greatly increase the refrigerant mass flow rate, which results in the heating capacity improvement. With the increased compressor speed, the compression pressure ratio increased from 3.8 to 6.3, and then the compressor enthalpy difference increased by 64% from 82 kJ kg−1 to 134 kJ kg−1, which accounts for the big increase of compressor input power and the reduction of COP. What’s more, at the highest compressor speed of 8000 RPM, the SGC system achieved 6.4 kW heating capacity, 59.4 °C Tg,a,o and 1.6 COP at mild cold temperature of −10 °C, and 5.6 kW heating capacity, 40.4 °C Tg,a,o and 1.8 COP at extremely cold temperature of −20 °C. This indicated that the SGC CO2 heat pump system could successfully provide comfortable warm air and sufficient heating capacity to passenger cabin even under the conditions of extreme cold weather (−20 °C) and severe heating loads (100% fresh air). Based on it, it is concluded that the utilization of the SGC CO2 heat pump system greatly outperforms the PTC heating supplement, which would significantly save electricity consumption and extend driving range for EVs.
cold outdoor weather conditions for heating mode operation of an EV, −10 °C and −20 °C, were studied. In the meanwhile, the indoor air inlet temperature was maintained the same as the outdoor temperature to simulate 100% fresh air (0% return air) operation condition. For SGC system, the optimum refrigerant charge was 0.9 kg, which was stated in the Section 3.1, whereas for OGC system, the optimum refrigerant charge was determined at 0.6 kg by the same experiment method as the SGC system. The increase of volume and the change of operation conditions can explain the refrigerant charge increase for SGC system. Firstly, the added GC in SGC system had an internal volume of 1.5 * 10−4 m3 . Secondly, under the optimum charge condition, the refrigerant density at GC outlet of SGC system was 896 kg m−3, 38% higher than 650 kg m−3 of OGC system, which leads to more refrigerants kept in HXs and pipes of SGC system. The heating performance of SGC and OGC system under varied operation conditions are shown in Fig. 8. Fig. 8a reveals that the heating capacity of the SGC system was 20–33% higher than the OGC system, and the COP of the SGC system was 17–32% better than the OGC system. The p-h diagram for results in 6000 RPM compressor speed is shown in Fig. 8c. A distinct difference from the OGC system is that Tg,r,o in the SGC system was much lower. For instance, at the temperature of −20 °C, Tg,r,o was −1.2 °C and 18.8 °C for SGC and OGC, respectively. This great reduction of Tg,r,o represents that the heat transfer temperature difference between the cold air and hot refrigerant was narrowed. Tg,r,o is a significant influencing parameter in CO2 heat pump system, whose decrease will produce heating capacity and COP enhancement. Thus it is concluded that the proposed method of SGC could greatly enhances GC rejection
3.3. Effect of indoor return air ratio The comparison experiments were conducted to investigate the
Fig. 8. Heating performance of EV CO2 heat pump with different GC configurations (a) heating capacity and COP, (b) Tg,a,o and compressor input power, (c) p-h diagram. 662
Applied Thermal Engineering 142 (2018) 656–664
D. Wang et al.
subcritical and trans-critical cycles for further application. 4.2. SGC heat pump system The method of SGC was proposed for EV CO2 heat pump system and experimentally validated. By comparing with OGC, SGC not only increases the heat transfer area of GC, but also strengthens the counter flow of hot refrigerant and cold air. One major feature for CO2 refrigerant is that it has a large gliding temperature in the supercritical region, and all of the heat is rejected from the GC without undergoing an isothermal condensation phase change. And one characteristic of EV heat pump system is that the indoor heated air experiences a huge temperature rise from cold ambient temperature to warm supply temperature. Thus it could be inferred that the stronger counter flow brought by SGC contributes to the matching of refrigerant glide curve and air temperature rise curve, which achieves heat transfer benefits in GC and performance improvements in system. In order to cope with the problem of insufficient heating capacity of R1234yf or R134a heat pump system at low ambient temperature, vapor injection heat pump system for EVs has been developed and investigated in recent years [2–4]. By comparing this up-to-date mobile heat pump technology, CO2 heat pump system still owns great performance advantages with much stronger capacity and warmer supply air. For instance, at the ambient temperature of −20 °C, the vapor injection system using R134a performs with 3.5 kW heating capacity and 20 °C Tg,a,o [2], whereas CO2 heat pump system could achieve 5.6 kW heating capacity and 40 °C Tg,a,o. The proposed SGC system has the advantage of avoiding the change of indoor HX1 dimensions and HVAC module structure. However, since the indoor HX2 is used as a secondary GC in heating mode for SGC system, it must meet the requirement of high pressure resistance of GC operation. In addition, the using of SGC system will add a stop valve, increase the refrigerant charge, and extend the refrigerant lines, thereby increasing cost and weight of the system. In order to reduce these negative effects, it is possible to optimize the layout of the components to shorten the refrigerant lines, and to develop multiple valve combination package to reduce the cost of the valve.
Fig. 9. Heating performance of EV CO2 heat pump with different indoor return air ratios.
effect of indoor return air ratio on heating performance of the SGC system at the ambient temperature of −20 °C, whose results are shown in Fig. 9. In this study, the indoor return air is assumed as 20 °C, which represents the typical return air condition in practical vehicle operation. Based on this reasonable assumption, four different indoor inlet air temperatures (−16 °C, −12 °C, −8°C, −4°C) were adopted to simulate various indoor return air ratios (10%, 20%, 30%, 40%). The compressor speed was manually adjusted to obtain the same Tg,a,o of 40 °C, which is a comfortably warm supply air level for passengers. When the return air ratio increased from 0% to 40%, the heating capacity decreased by 24% from 5.6 kW to 4.3 kW, and the compressor input power decreased by 30% from 3.2 kW to 2.2 kW. Besides, the compressor speed decreased from 8000 RPM at 0% return air ratio to 5200 RPM at 40% return air ratio, which accounts for 30% reduction of compressor input power. It indicated that the utilization of return air was an effective method to reduce heating loads and save electricity consumption for cabin heating when using CO2 heat pump system. Fig. 9 also shows that the COP was in the range of 1.8–1.9 when return air ratio changed, which indicated that return air method has little impact on the heating efficiency of EV CO2 heat pump system.
4.3. Other discussion Cabin return air utilization is a promising approach for EVs to reduce heating loads and save battery power, so a certain percentage of return air will be considered for mobile heat pump system. Under return air conditions, we predict that the SGC system is still superior to the OGC system, but the performance improvement may not be as good as that under 100% fresh air conditions. The increased heat transfer area of GC in SGC system could improve the heating performance all the time. But the utilization of return air will reduce the total heating loads, for example from 6.5 kW to 4 kW, which will lead to the reduction of performance benefit of SGC system. The CO2 heat pump system provides the air cooling and air heating functions for an EV. In this study, we pay emphasis on the heating performance improvement of the SGC heat pump system. In our future work, we will try to quantify the effects of cooling and heating on the energy consumption and driving range for an EV. The influencing parameters such as vehicle driving condition and weather conditions will be considered. Based on the quantification model, the comparison results of all year performance between a SGC CO2 heat pump system with an OGC CO2 heat pump system, and an R134a heat pump system could be further revealed. In addition, we obtained the experimental results of heating performance for the proposed SGC CO2 heat pump system at cold ambient temperatures of −10 °C and −20 °C in this study. Future studies are warranted to evaluate the steady and transient heating performance of this heat pump system under variable operation conditions for on-road vehicle applications. More experimental data to describe the performance of CO2 HXs and electrical compressor will be
4. Further discussion 4.1. System control of mobile CO2 heat pump system Quite a few literature have discussed the high pressure control by EXV adjustment in mobile CO2 AC system, and proposed experimental correlations between optimum pressure and temperature at GC outlet. However, the mobile CO2 heat pump system is significantly different from the mobile CO2 AC system in terms of operation conditions and cycle characteristics. Therefore, the correlations for calculating the optimum pressure in CO2 AC system can’t be directly applied to CO2 heat pump system. If the optimum pressure refers to COP maximization, the approach for mobile CO2 heat pump operation may be the same as mobile AC operation. This paper has revealed following important features in mobile CO2 heat pump system: the GC heat rejection amount is greatly improved with the increase of discharge pressure; the indoor air experiences a huge rise in temperature (for example, from −20 °C to 40 °C); and the compressor discharge temperature is close to limitation under extreme operation conditions. These features may help establish proper control strategy for mobile CO2 heat pump system in future research. Furthermore, the EXV opening needs to be regulated accurately both under subcritical and supercritical inlet pressure conditions. Different control strategies need to be carefully considered in 663
Applied Thermal Engineering 142 (2018) 656–664
D. Wang et al.
Acknowledgements
obtained, which helps to establish accurate prediction models of components and system. The experimental results in this paper would be employed to validate the model accuracy for steady operation. Using the prediction models, the heating results will be extended for other operation conditions, and the impacts of control parameters on system performance such as air flow rate and compressor speed will be deeply revealed.
Thanks to support by the Natural Science Foundation of China (No. 51776119) References [1] Z. Zhang, et al., Electric vehicle range extension strategies based on improved AC system in cold climate – a review, Int. J. Refrig. 88 (2018) 141–150. [2] F. Qin, et al., Experimental investigation and theoretical analysis of heat pump systems with two different injection portholes compressors for electric vehicles, Appl. Energy 185 (2017) 2085–2093. [3] Y. Higuchi et al., Efficient Heat Pump System for PHEV/BEV, SAE International, 2017. [4] J. Jung, et al., Numerical study of the effects of injection-port design on the heating performance of an R134a heat pump with vapor injection used in electric vehicles, Appl. Therm. Eng. 127 (2017) 800–811. [5] G. Lorentzen, J. Pettersen, A new, efficient and environmentally benign system for car air-conditioning, Int. J. Refrig. 16 (1) (1993) 4–12. [6] M.-H. Kim, J. Pettersen, C.W. Bullard, Fundamental process and system design issues in CO2 vapor compression systems, Prog. Energy Combust. Sci. 30 (2) (2004) 119–174. [7] K. Nawaz, et al., Performance optimization of CO2 heat pump water heater, Int. J. Refrig. 85 (2018) 213–228. [8] P. Gullo, B. Elmegaard, G. Cortella, Energy and environmental performance assessment of R744 booster supermarket refrigeration systems operating in warm climates, Int. J. Refrig. 64 (2016) 61–79. [9] S. Andersen et al., Comparative Manufacturing and Ownership Cost Estimates for Secondary Loop Mobile Air Conditioning Systems (SL-MACs), SAE International, 2017. [10] Daimler, https://www.daimler.com/sustainability/product/further-environmentaltechnologies/co2-air-conditioning-system.html (Daimler official website), 2017. [11] J.S. Brown, S.F. Yana-Motta, P.A. Domanski, Comparitive analysis of an automotive air conditioning systems operating with CO 2 and R134a, Int. J. Refrig. 25 (1) (2002) 19–32. [12] S.C. Kim, J.P. Won, M.S. Kim, Effects of operating parameters on the performance of a CO2 air conditioning system for vehicles, Appl. Therm. Eng. 29 (11–12) (2009) 2408–2416. [13] 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 (8) (2005) 1302–1307. [14] H. Hammer, A. Audi, Comparative study of AC-and HP-systems using the refrigerants R134a and R744, in: VDA Alternate RefrigerantWinter Meeting, 2002. [15] S.C. Kim, et al., Heating performance enhancement of a CO2 heat pump system recovering stack exhaust thermal energy in fuel cell vehicles, Int. J. Refrig. 30 (7) (2007) 1215–1226. [16] J. Liu, et al., Experimental study and numerical simulation Concerning fogging characteristics and Improvement of return air utilization for electric vehicles, Appl. Therm. Eng. 129 (2018) 1115–1123. [17] D. Wang, et al., Heating performance characteristics of CO2 heat pump system for electrical vehicle in a cold climate, Int. J. Refrig. 85 (2018) 27–41. [18] R.J. Moffat, Describing the uncertainties in experimental results, Exp. Therm. Fluid Sci. 1 (1) (1988) 3–17. [19] Y. Hou, et al., Experimental investigation on the influence of EEV opening on the performance of transcritical CO2 refrigeration system, Appl. Therm. Eng. 65 (1) (2014) 51–56. [20] C. Liu, et al., Mass flow characteristics and empirical modeling of R744 flow through electronic expansion device, Int. J. Refrig. 86 (2018) 82–88.
5. Conclusions This paper studied the heating performance of an EV CO2 heat pump system with a novel GC configuration. A newly developed mobile CO2 compressor was adopted to set up this CO2 heat pump system. The impacts of CO2 refrigerant charge, GC configuration and indoor return air ratio on heating characteristics of EV CO2 heat pump system have been revealed by experiments. The conclusions are summarized as follows. 1. At the temperature of −10 °C, the bigger refrigerant charge and smaller DSH control were beneficial to Tg,a,o and heating capacity improvements. The increased discharge pressure enhanced the heat rejection amount of CO2 GC. The refrigerant charge should be less than a certain value for EV CO2 heat pump system to prevent exceeding discharge temperature limitation. 2. The method of SGC greatly enhanced the heat rejection amount of GC. The improvements of heating capacity and COP for the EV CO2 heat pump system were up to 33% and 32%, respectively, when using SGC instead of OGC. At the extreme cold temperature of −20 °C, the novel EV CO2 heat pump system achieved 5.6 kW maximum heating capacity and 1.8 COP, which is much better than vapor injection heat pump system using conventional refrigerant. 3. The SGC CO2 heat pump system could successfully provide comfortable warm air and sufficient heating capacity to passenger cabin even under the conditions of extreme cold temperature (−20 °C) and severe heating loads (100% fresh air). The utilization of the SGC CO2 heat pump system greatly outperforms the conventional PTC heating supplement, which would significantly save electricity consumption and extend driving range of EVs. 4. The utilization of return air was an effective method to reduce heating loads and save electricity consumption for cabin heating when using CO2 heat pump system. And the return air method has little impact on the heating efficiency of EV CO2 heat pump system. In conclusion, this paper has proved CO2 mobile heat pump system is a promising heat pump technology for EVs with great heating advantages in a cold climate, and established a strong foundation for the future application of natural refrigerant CO2 in EVs.
664
Contents lists available at ScienceDirect
Applied Thermal Engineering journal homepage: www.elsevier.com/locate/apthermeng
Research Paper
Heating performance evaluation of a CO2 heat pump system for an electrical vehicle at cold ambient temperatures Dandong Wanga, Binbin Yua, Wanyong Lia, Junye Shia,b, Jiangping Chena,b, a b
T
⁎
Institute of Refrigeration and Cryogenics, Shanghai Jiaotong University, Shanghai, China Shanghai High Efficient Cooling System Research Center, Shanghai, China
H I GH L IG H T S
CO heat pump system with Series Gas Cooler configuration was proposed. • AA novel developed mobile CO compressor was applied to EV CO heat pump system. • Thenewly of charge, gas cooler and return air ratio were investigated. • The impacts • CO heat pump achieved great heating capacity and COP at −20 °C temperature. 2
2
2
2
A R T I C LE I N FO
A B S T R A C T
Keywords: Electrical vehicle Heat pump CO2 refrigerant Trans-critical Gas cooler
A novel CO2 heat pump system with Series Gas Cooler (SGC) configuration was proposed for application in an Electrical Vehicle (EV). A newly developed mobile CO2 compressor with 6.8 ml displacement and 8000 RPM maximum speed was employed. The heating performance of this CO2 heat pump system was experimentally evaluated under low temperature conditions. Comparison experiments were conducted to reveal the effects of CO2 refrigerant charge, GC configuration and indoor return air ratio on heating performance characteristics. Results show that using SGC improves the heating capacity and COP by up to 33% and 32%, respectively, compared with One Gas Cooler (OGC). At the extreme cold temperature of −20 °C, the novel CO2 heat pump system achieved 5.6 kW maximum heating capacity and 1.8 COP, which greatly outperforms the conventional PTC heating supplement for an EV.
1. Introduction In recent years, mobile air source heat pump systems are becoming increasingly popular in the application of Electrical Vehicles (EV) due to the benefits of improved heating efficiency for cabin heating [1]. However, a mobile heat pump system with conventional refrigerant (HFC-134a or HFO-1234yf) suffers significantly diminishment of heating capacity and system efficiency as the ambient temperature decreases. To face the challenge of the performance attenuation in winter, there are two promising options. One is to develop a high efficient vapor injection heat pump system using conventional refrigerant, the other is to develop a mobile heat pump system using natural refrigerant CO2. Qin et al. [2] have tested an R134a vapor injection heat pump system for an EV under −20 °C ambient conditions and investigated the effect of injection portholes shapes on heating performance. It is found that the best system achieved 3.5 kW heating capacity at compressor ⁎
speed of 8500 RPM. When the inlet air temperature is lower than −10 °C, the performance of the three interlinked portholes system is similar as that of single porthole system. Higuchi et al. [3] provided a novel flash tank vapor injection heat pump system using R1234yf refrigerant, which is installed in 2017′ Toyota Prius. This heat pump system can achieve multi-functions by regulating several valves. And it can achieve a 55 °C air supply temperature under 5 °C ambient condition in heating defogging mode, with a heating COP of 2.7. Few data in heating mode was released in this paper. Jung et al. [4] developed a simulation model for an R134a vapor injection heat pump system and numerically analyzed the effect of injection port geometries. The optimum angle of the single-injection port is determined to be 440°. Although the vapor injection heat pump system significantly improves the heating capacity at low temperatures, the enhanced heating capacity is still insufficient for cabin heating under high heating loads conditions (such as 100% fresh air and −20 °C ambient temperature), thus a PTC with a small capacity of 2–3 kW must be equipped to provide a heat
Corresponding author at: Institute of Refrigeration and Cryogenics, Shanghai Jiaotong University, Shanghai, China. E-mail address: [email protected] (J. Chen).
https://doi.org/10.1016/j.applthermaleng.2018.07.062 Received 18 March 2018; Received in revised form 11 July 2018; Accepted 12 July 2018 1359-4311/ © 2018 Elsevier Ltd. All rights reserved.
Applied Thermal Engineering 142 (2018) 656–664
D. Wang et al.
Nomenclature
p PTC Q R RPM SGC T W*H*D W X
Roman AC COP Cp DSH DSC EV EXV GC h HVAC HX IHX m OGC
air conditioning coefficient of performance specific heat (J kg−1 K−1) degree of superheat (K) degree of sub-cooling (K) electrical vehicle electrical expansion valve gas cooler enthalpy (kJ kg−1) heating, ventilation and air conditioning heat exchanger internal heat exchanger mass flow rate (kg s−1) one gas cooler
pressure (MPa) positive temperature coefficient heating capacity (kW) result revolutions per minute series gas cooler temperature (K) Width * Height * Depth compressor input power (kW) measurement
Subscripts a i g o r
air inlet gas cooler outlet refrigerant
for conventional vehicles and concluded that its dehumidifying efficiency performance exceeds that of the HFC-134a system. Hammer and Audi [14] experimentally compared the vehicle warming-up time between a system with a CO2 heat pump system and with a conventional PTC heater for an Audi car with a 1.6 L gasoline engine. Kim et al. [15] found that use of a CO2 heat pump system with exhaust heat recovery for fuel cell vehicles could improve heating capacity by 100%, compared to conventional PTC heating. In conclusion, further research is needed to fully evaluate the heating performance characteristics of mobile CO2 heat pump system for EVs. When the CO2 heat pump system is operated in trans-critical cycle, the performance of gas cooler (GC) has been identified as a critical component as it affects the heat rejection pressure, refrigerant exit temperature and secondary fluid exit temperature. Therefore, the whole system performance relies heavily on the design of the GC. In addition, cabin return air utilization is a promising approach for EVs to reduce heating load and save electricity power [16]. The effect of this strategy on CO2 heat pump performance remains to be determined. An EV CO2 heat pump system has been designed and experimentally investigated in previous research [17], which concluded that the heat pump system using CO2 refrigerant exhibited good heating performance in a cold climate. However, the achieved heating capacity of 3.6 kW at −20 °C condition is insufficient for passenger cabin heating. In order to
supplement when using vapor injection heat pump system for an EV. Natural refrigerant CO2 offers no toxicity, no flammability, high volumetric capacity, and high heat transfer properties [5,6]. Based on the advantage of thermo-physical properties, CO2 refrigerant has been successfully applied in tap water heat pumps [7] and supermarket refrigeration systems [8]. As a countermeasure against global warming, regulations in the EU, Japan and the US are phasing out the use of HFC134a (GWP = 1430) in mobile air conditioning (AC) systems [9]. Natural refrigerant CO2 (GWP = 1) is considered as a promising alternative to HFC-134a. The European automotive manufacture DaimlerBenz has announced its first-ever plan to offer a CO2 AC system in a production vehicle [10]. Several researches have studied the cooling performance of mobile CO2 AC systems. Brown et al. [11] evaluated the performance merits of CO2 and R134a automotive AC systems. The result from semi-theoretical cycle models show that the COP of CO2 was lower by 21% at 32.2 °C and by 34% at 48.9 °C. The COP disparity was even greater at higher speeds and ambient temperatures. Kim et al. [12] studied the effects of operating parameters on the cooling performance of a CO2 automotive AC system and concluded that the CO2 refrigerant system exhibited good performance. However, there are only a few data available on mobile CO2 heat pump systems, especially for cold climate operation. Tamura et al. [13] developed a prototype CO2 automotive cooling and heating AC system
Fig. 1. Schematic diagram of systems with (a) conventional OGC and (b) novel SGC. 657
Applied Thermal Engineering 142 (2018) 656–664
D. Wang et al.
buffer. Both the OGC and SGC heat pump system can be switched from cooling to heating mode operation via stop valves and EXVs. There is no cycle difference between these two systems at cooling mode. For OGC at heating mode, after being discharged from compressor, the refrigerant enters into the indoor HX1 (No.2 in Fig. 1a) and rejects heat to cold air, then directly passes through the EXV to the low pressure side of the cycle. The indoor HX2 (No.3 in Fig. 1a) is bypassed when all refrigerant flows from the outdoor HX (No.6 in Fig. 1a) to the accumulator. For SGC, however, after the heat rejection in the indoor HX1 (No.2 in Fig. 1b), refrigerant first passes through the full open EXV (NO.8–1 in Fig. 1b), which doesn’t produce pressure drop here, then enters into the indoor HX2 (No.3 in Fig. 1b) to release extra heat to cold inlet air. Thus the indoor HX2 (No.3 in Fig. 1b) is utilized as a second GC in the SGC system, and the cold inlet air is capable of absorbing the rejected heat from indoor HX1 (No.2 in Fig. 1b) and indoor HX2 (No.3 in Fig. 1b). After the heat rejection process in first GC and second GC, the refrigerant passes through the EXV (8–2 in Fig. 1b) and is throttled to low pressure state. After throttling, the two-phase refrigerant evaporates in outdoor HX (No.6 in Fig. 1b), which is used as an evaporator at heating mode. Then the refrigerant enters into the accumulator and goes back to the compressor. When air dehumidification is required to avoid windshield flash fogging, the EXV (NO.8–1 in Fig. 1b) can be regulated to throttle the refrigerant after indoor HX1 (No.2 in Fig. 1b), then the indoor HX2 (No.3 in Fig. 1b) is used as an evaporator and dehumidifies the inlet air. Figs. 2 and 3 show the p-h diagrams of CO2 cycles and the psychrometric chart of indoor air for two CO2 heat pump systems, respectively. As for the IHX, at cooling mode, the refrigerant after outdoor HX (GC, No.6 in Fig. 1) passes through the IHX, undergoing a heat transfer process to the low pressure refrigerant flowed from accumulator, which decreases the inlet quality of indoor HX (Evaporator, No.3 in Fig. 1) and thereby improves the cooling performance of system. At heating mode, both sides of the IHX are low pressure and low temperature refrigerants, so the heat transfer between the two sides of the IHX is negligible and the IHX can be regarded as not working. If IHX was able to work at both cooling mode and heating mode, more switching stop valves will be required and the refrigerant pipes need be extended, thus the working of IHX at heating mode is not considered in this study. The novel SGC system enhances the total heat rejection area of GC by utilizing the indoor HX2 (No.3 in Fig. 1b). One simple choice to achieve the same result of novel SGC in conventional OGC system is to directly enlarge the indoor HX1 (No.2 in Fig. 1a), but this will lead to the increase of air side pressure drop and HX installation space. In addition, as the indoor HX1 (No.2 in Fig. 1a) is installed inside the HVAC module (No.4 in Fig. 1) of the vehicle, the space increase of the
further improve the heating capacity of CO2 mobile heat pump system, this paper put forward a novel CO2 heat pump system using two indoor heat exchangers (HXs) as heat rejection GCs. The total GC heat transfer area can be increased without the change of HVAC module structure. In addition, the CO2 compressor used in previous study [17] was originally designed for domestic CO2 heat pump water heater application. The maximum allowable compressor discharge pressure and temperature were 12 MPa and 120 °C, respectively, which limit the heating performance of mobile CO2 heat pump system. A newly developed mobile rotary type CO2 compressor was adopted in this study. It can be operated up to 8000 RPM speed and 150 °C discharge temperature. With this compressor, comparison experiments were conducted to reveal the impacts of refrigerant charge, GC configuration and indoor return air ratio on CO2 heating performance at the ambient temperatures of −10 °C and −20 °C. The cycle characteristics and performance advantages of the novel CO2 heat pump system were further analyzed and discussed based on the experimental results.
2. Experimental setup 2.1. Experimental apparatus The novel EV CO2 heat pump system is provided considering the HX arrangements. Fig. 1 shows schematic diagram of systems with conventional OGC and novel SGC. Each system is composed of an electrical compressor, an outdoor HX, an indoor HX1, an indoor HX2, an internal heat exchanger (IHX), an accumulator, several stop valves and expansion valves. The CO2 mobile electrical compressor, integrated with a motor driver, is a newly developed rotary type compressor. It is driven by 300–400 V DC electricity power and uses PAG lubricant oil. It has a displacement of 6.8 ml and a weight of 6.7 kg. Besides, it can achieve the speed of up to 8000 RPM and the discharge temperature of up to 150 °C. The CO2 HXs are micro-channel type made of aluminum with high operating pressure resistance of 14 MPa. The sizes of indoor HX1 (No.2 in Fig. 1), indoor HX2 (No.3 in Fig. 1), and outdoor HX (No.6 in Fig. 1) are 180W * 190H * 32D (mm), 230W * 250H * 32D (mm), and 600W * 450H * 12D (mm), respectively. Their dimensions fit with the specifications for a commercial B-class vehicle. The expansion valve used here is a CO2 EXV integrated with a step motor driver, connected to the controller by Lin (Local Interconnect Network) communication. On the controller interface, the input 0–5 DC voltage could be manually adjusted to change the steps of the motor and then regulate the CO2 mass flow rate. This EXV is developed for mobile CO2 heat pump system application. It has throttling function, allows full flow in one direction, and can be closed bi-directionally. An accumulator 1.0 L in volume is installed at compressor suction to serve as a system refrigerant charge
Fig. 2. p-h diagrams of CO2 cycle for (a) OGC system and (b) SGC system under heating operation. 658
Applied Thermal Engineering 142 (2018) 656–664
D. Wang et al.
Fig. 3. Psychrometric charts of indoor air for (a) OGC system and (b) SGC system under heating operation.
indoor HX1 will cause the structure redesign and volume increase of whole HVAC module. Therefore, the proposed SGC system has the advantage of avoiding the change of indoor HX1 dimensions, air side pressure drop and HVAC module structure. But the drawback is that one stop valve is added and the refrigerant lines are extended, which can be seen from Fig. 1. The heating performance of the OGC and SGC system were tested in an automobile heat pump calorimeter facility, whose schematic diagram is shown in Fig. 4. The detailed facility introduction can be referred in a previous research paper [17]. It consisted of an outdoor chamber and an indoor chamber with different open wind tunnels. The outdoor GC and indoor HVAC module were installed before the outdoor wind tunnel and indoor wind tunnel, respectively.
Table 1 Uncertainties of the experimental parameters and measured data. Items
Uncertainties
Temperature sensors (RTD-type, Yokogawa) Pressure transducers (GE-Druck)
± 0.2 °C
Mass flow rate (Coriolis type, KROHNE) Digital power meter (WT210, Yokogawa) Data logger (34972A, Agilent) Heating capacity Heating COP
2.2. Measurement uncertainties Measurement devices were equipped to measure refrigerant and air
Fig. 4. Schematic diagram of calorimeter facility to measure CO2 heating performance. 659
± 0.5% of full scale, Max 10 MPa or 20 MPa, ± 0.15% of full scale, Max 600 kg h−1 ± 0.5% reading 0.004% dcV accuracy 5.5% 6.3%
Applied Thermal Engineering 142 (2018) 656–664
D. Wang et al.
different refrigerant charge conditions. The reduction of DSH at the compressor suction slightly enhanced refrigerant density and refrigerant mass flow rate, which accounted for the impact of DSH on heating capacity. Meanwhile, the DSH had small impact on COP, which was less than 6%. Based on the effects of DSH on heating performance, the DSH was controlled at 0 K by EXV adjustment to achieve a relatively greater heating capacity in later experiments. In addition, taking the maximum heating capacity under 0 K DSH condition as a comparison, the heating capacity increased by 16% from 4.5 kW to 5.2 kW, and the COP increased by 9% from 1.88 to 2.05, when refrigerant charge changed from 0.6 kg to 0.9 kg. It could be seen from Fig. 7c that adding refrigerant charge increased the discharge pressure and the enthalpy difference of GC, which leads to the increase of heating capacity. In addition, the compression pressure ratio and the enthalpy difference of compressor increased with the increase of refrigerant charge, which explains that the increase of COP is smaller than heating capacity. Besides, the results of air outlet temperature of GC (Tg,a,o) and EXV opening are shown in Fig. 7b. The decrease of DSH from 10 K to 0 K improved the Tg,a,o by 2–3 K, and the increase of refrigerant charge from 0.6 kg to 0.9 kg enhanced the highest Tg,a,o by 7.4 K from 37.2 °C to 44.6 °C. Thus the bigger refrigerant charge and smaller DSH control were beneficial to Tg,a,o and heating capacity improvements in this operation condition. Fig. 7b also shows that the EXV opening was decreased with the increase of refrigerant charge or DSH. It was noticed that the EXV opening changed from maximum 73% to minimum 14%, whose variation was related to the change of CO2 inlet pressure, inlet temperature, outlet pressure and refrigerant mass according to mass flow characteristics of CO2 EXV [19,20]. This phenomenon of large EXV opening variation in heating operation requires special consideration in future development and design of the system control. What’s more, the p-h diagram in Fig. 7c shows that the system discharge pressure became higher with increased refrigerant amount adding into the system. At first, when the refrigerant charge is 0.6 kg, the system exhibited CO2 sub-critical cycle, and the refrigerant outlet temperature of GC (Tg,r,o) was 24 °C with 1 K sub-cooling degree (DSC). Then, when refrigerant charge increased to 0.9 kg, the system cycle transformed to CO2 trans-critical cycle, Tg,r,o was lowered to 5 °C. Apart from the reduction of Tg,r,o, the enthalpy difference of GC increased by 107kJ kg-1 with the change of refrigerant charge. As refrigerant charge increased from 0.6 kg to 0.9 kg, the pressure drop of high pressure side GCs decreased from 0.68 MPa to 0.20 MPa, and the pressure drop of low pressure side evaporator and accumulator decreased from 0.66 MPa to 0.24 MPa. The reduction of refrigerant mass flow from 73.1 kg h−1 to 51.8 kg h−1 could account for the reduction of pressure drop. When the refrigerant charge was 0.9 kg, the
conditions. Table 1 shows the uncertainties of the experimental parameters and measured data. The location of sensors on the refrigerant side and air side are shown in Figs. 1 and 4. The heating capacity (Q) of the tested CO2 heat pump was determined by the air side heat transfer rate using Eq. (1). The overall system COP was determined by Eq. (2). The refrigerant side heat transfer rate was also calculated as a comparison by Eq. (3). The properties of the CO2 refrigerant were calculated according to the NIST REFPROP 8.0. Fig. 5 shows that the errors between the heat transfer rate for the air side and the refrigerant side were within ± 5%, which indicated tested system has a good heat balance between air side and refrigerant side.
Qa = ma Cp (Ta, out −Ta, in ) = Cp (Ta, out −Ta, in ) ρa v a A COP =
(1)
Q Q = W U ∗I
(2)
Qr = mr (hg, in−hg, out )
(3)
In order to verify the measured data of the heating capacity and COP, a single-sample uncertainty analysis was performed according to Moffat [18], using the equations presented in Eqs. (4) and (5). Using this analysis method, the overall relative uncertainty of the heating capacity and COP were calculated as 5.5% and 6.3%, respectively. 1/2
δCp 2 δρ δQ δ ΔTa 2 δv δA = ⎜⎛( ) +( ) + ( a )2 + ( a )2 + ( )2⎟⎞ Q C Δ T ρ v A ⎠ p a a a ⎝
(4)
1/2
δCOP δQ δU 2 δI ) + ( ) 2⎞ = ⎛( )2 + ( COP Q U I ⎠ ⎝ ⎜
⎟
(5)
2.3. Test method The impact of refrigerant charge on trans-critical cycle characteristics of EV CO2 heat pump system was evaluated in system experiments at −10 °C ambient temperature, and the degree of superheat (DSH) at compressor suction was varied to determine its effect on heating performance. Next, with the determined refrigerant charge, the heating performance of the OGC and SGC system were experimentally compared at the ambient temperatures of −10 °C and −20 °C and 4000/ 6000/8000 RPM compressor speeds. Besides, the effect of indoor return air utilization on the system performance was investigated with various return air ratios from 0% to 40%. Table 2 shows the details of the operation conditions of system experiments. 3. Results and discussion 3.1. Effect of CO2 refrigerant charge System experiments were first conducted with refrigerant charge varying from 0.6 kg to 0.9 kg at the ambient temperature of −10 °C. Unlike optimum high pressure control in cooling mode, here EXV opening was manually regulated to obtain various refrigerant DSHs at the compressor suction, including 0 K (saturation point), 5 K, and 10 K. The pictures in Fig. 6 shows the refrigerant flow behavior with different DSHs, which were captured by the camera in front of the sight glass. When DSH was higher than 5 K (Fig. 6b and c), the refrigerant flow was sequent steady flow, which represents the superheated vapor state of the refrigerant. When DSH was 0 K (Fig. 6a), intermittent liquid-gas spray flow occurred, representing the two-phase state of the refrigerant. Since the measured small DSH (less than 2 K) can’t be used to judge the saturation point between the two-phase and the superheated vapor state of the refrigerant, the appearance of spray flow was utilized to determine this 0 K DSH saturation condition during experiments. Fig. 7 shows the experimental results of heating performance with different refrigerant charges and DSHs. From Fig. 7a, with the decrease of DSH from 10 K to 0 K, the heating capacity increased by 4–7% under
Fig. 5. Comparison of heat transfer rate between refrigerant side and air side. 660
Applied Thermal Engineering 142 (2018) 656–664
D. Wang et al.
Table 2 Operation conditions of system experiments.
1 2 3
Refrigerant charge (kg)
Indoor air inlet temperature (°C)
Indoor air inlet mass flow (kg h−1)
Outdoor air inlet temperature (°C)
Outdoor air inlet velocity (m s−1)
Compressor speed (RPM)
0.6, 0.7, 0.8, 0.9 0.6, 0.9 0.9
−10 −10, −20 −20, −16, −12, −8, −4
337 337 337
−10 −10, −20 −20
4 4 4
4000 4000, 6000, 8000 8000, 7200, 6500, 5800, 5200
temperature from exceeding limitation, this CO2 SGC heat pump system has not been tested with refrigerant charge more than 0.9 kg. 0.9 kg was considered as the largest refrigerant charge at the ambient temperature of −10 °C. The condition of 0 K DSH was taken to consider the optimum refrigerant charge. The heating capacity and COP increased with the adding of refrigerant charge. At 0.9 kg, this system achieved a highest heating capacity of 5.2 kW and a best COP of 2.05. If more refrigerant are added into this system, the discharge temperature will exceed the limitation of 150 °C, which can be reduced by increasing the opening of EXV. However, some gas-liquid refrigerant will flow into compressor, which is harmful to the compressor reliability, and the COP of system will be decreased because of the increase in the power consumption of compressor. Therefore, 0.9 kg was then chose as the optimum system refrigerant charge.
Fig. 6. Refrigerant flow behavior at compressor suction with (a) 0 K DSH, (b) 5 K DSH and (c) 10 K DSH.
two phase equilibrium temperature was −22.2 °C at evaporator inlet, then it was dropped to −26.7 °C at compressor suction due to the pressure drop of 0.24 MPa. The huge pressure loss and equilibrium temperature reduction in low pressure side will result in the reduction of the suction mass flow of compressor, the heating capacity and the COP of system, which needs to be optimized in the future research. In addition, for 0.9 kg refrigerant charge, the discharge temperature increased to 145 °C, close to discharge limitation 150 °C, which is caused by the low compressor suction pressure of 1.9 MPa and the high discharge pressure of 9.0 MPa. In order to prevent the discharge
3.2. Effect of GC configuration In this study, the SGC CO2 heat pump system was put forward for EV, whose performance was experimentally compared with the OGC system at compressor speeds of 4000, 6000 and 8000 RPM. Two typical
Fig. 7. Heating performance of EV CO2 heat pump with different refrigerant charges (a) Tg,a,o and EXV opening, (b) heating capacity and COP, (c) p-h diagram. 661
Applied Thermal Engineering 142 (2018) 656–664
D. Wang et al.
amount and thus boosts the heating performance of EV CO2 heat pump system. Fig. 8a and b also shows the effect of compressor speed on heating performance. For example, at −20 °C for SGC system, the heating capacity and compressor input power increased by 74% and 185%, and the corresponding COP decreased by 41% as the compressor speed increased from 4000 RPM to 8000 PRM. The increase of compressor speed could greatly increase the refrigerant mass flow rate, which results in the heating capacity improvement. With the increased compressor speed, the compression pressure ratio increased from 3.8 to 6.3, and then the compressor enthalpy difference increased by 64% from 82 kJ kg−1 to 134 kJ kg−1, which accounts for the big increase of compressor input power and the reduction of COP. What’s more, at the highest compressor speed of 8000 RPM, the SGC system achieved 6.4 kW heating capacity, 59.4 °C Tg,a,o and 1.6 COP at mild cold temperature of −10 °C, and 5.6 kW heating capacity, 40.4 °C Tg,a,o and 1.8 COP at extremely cold temperature of −20 °C. This indicated that the SGC CO2 heat pump system could successfully provide comfortable warm air and sufficient heating capacity to passenger cabin even under the conditions of extreme cold weather (−20 °C) and severe heating loads (100% fresh air). Based on it, it is concluded that the utilization of the SGC CO2 heat pump system greatly outperforms the PTC heating supplement, which would significantly save electricity consumption and extend driving range for EVs.
cold outdoor weather conditions for heating mode operation of an EV, −10 °C and −20 °C, were studied. In the meanwhile, the indoor air inlet temperature was maintained the same as the outdoor temperature to simulate 100% fresh air (0% return air) operation condition. For SGC system, the optimum refrigerant charge was 0.9 kg, which was stated in the Section 3.1, whereas for OGC system, the optimum refrigerant charge was determined at 0.6 kg by the same experiment method as the SGC system. The increase of volume and the change of operation conditions can explain the refrigerant charge increase for SGC system. Firstly, the added GC in SGC system had an internal volume of 1.5 * 10−4 m3 . Secondly, under the optimum charge condition, the refrigerant density at GC outlet of SGC system was 896 kg m−3, 38% higher than 650 kg m−3 of OGC system, which leads to more refrigerants kept in HXs and pipes of SGC system. The heating performance of SGC and OGC system under varied operation conditions are shown in Fig. 8. Fig. 8a reveals that the heating capacity of the SGC system was 20–33% higher than the OGC system, and the COP of the SGC system was 17–32% better than the OGC system. The p-h diagram for results in 6000 RPM compressor speed is shown in Fig. 8c. A distinct difference from the OGC system is that Tg,r,o in the SGC system was much lower. For instance, at the temperature of −20 °C, Tg,r,o was −1.2 °C and 18.8 °C for SGC and OGC, respectively. This great reduction of Tg,r,o represents that the heat transfer temperature difference between the cold air and hot refrigerant was narrowed. Tg,r,o is a significant influencing parameter in CO2 heat pump system, whose decrease will produce heating capacity and COP enhancement. Thus it is concluded that the proposed method of SGC could greatly enhances GC rejection
3.3. Effect of indoor return air ratio The comparison experiments were conducted to investigate the
Fig. 8. Heating performance of EV CO2 heat pump with different GC configurations (a) heating capacity and COP, (b) Tg,a,o and compressor input power, (c) p-h diagram. 662
Applied Thermal Engineering 142 (2018) 656–664
D. Wang et al.
subcritical and trans-critical cycles for further application. 4.2. SGC heat pump system The method of SGC was proposed for EV CO2 heat pump system and experimentally validated. By comparing with OGC, SGC not only increases the heat transfer area of GC, but also strengthens the counter flow of hot refrigerant and cold air. One major feature for CO2 refrigerant is that it has a large gliding temperature in the supercritical region, and all of the heat is rejected from the GC without undergoing an isothermal condensation phase change. And one characteristic of EV heat pump system is that the indoor heated air experiences a huge temperature rise from cold ambient temperature to warm supply temperature. Thus it could be inferred that the stronger counter flow brought by SGC contributes to the matching of refrigerant glide curve and air temperature rise curve, which achieves heat transfer benefits in GC and performance improvements in system. In order to cope with the problem of insufficient heating capacity of R1234yf or R134a heat pump system at low ambient temperature, vapor injection heat pump system for EVs has been developed and investigated in recent years [2–4]. By comparing this up-to-date mobile heat pump technology, CO2 heat pump system still owns great performance advantages with much stronger capacity and warmer supply air. For instance, at the ambient temperature of −20 °C, the vapor injection system using R134a performs with 3.5 kW heating capacity and 20 °C Tg,a,o [2], whereas CO2 heat pump system could achieve 5.6 kW heating capacity and 40 °C Tg,a,o. The proposed SGC system has the advantage of avoiding the change of indoor HX1 dimensions and HVAC module structure. However, since the indoor HX2 is used as a secondary GC in heating mode for SGC system, it must meet the requirement of high pressure resistance of GC operation. In addition, the using of SGC system will add a stop valve, increase the refrigerant charge, and extend the refrigerant lines, thereby increasing cost and weight of the system. In order to reduce these negative effects, it is possible to optimize the layout of the components to shorten the refrigerant lines, and to develop multiple valve combination package to reduce the cost of the valve.
Fig. 9. Heating performance of EV CO2 heat pump with different indoor return air ratios.
effect of indoor return air ratio on heating performance of the SGC system at the ambient temperature of −20 °C, whose results are shown in Fig. 9. In this study, the indoor return air is assumed as 20 °C, which represents the typical return air condition in practical vehicle operation. Based on this reasonable assumption, four different indoor inlet air temperatures (−16 °C, −12 °C, −8°C, −4°C) were adopted to simulate various indoor return air ratios (10%, 20%, 30%, 40%). The compressor speed was manually adjusted to obtain the same Tg,a,o of 40 °C, which is a comfortably warm supply air level for passengers. When the return air ratio increased from 0% to 40%, the heating capacity decreased by 24% from 5.6 kW to 4.3 kW, and the compressor input power decreased by 30% from 3.2 kW to 2.2 kW. Besides, the compressor speed decreased from 8000 RPM at 0% return air ratio to 5200 RPM at 40% return air ratio, which accounts for 30% reduction of compressor input power. It indicated that the utilization of return air was an effective method to reduce heating loads and save electricity consumption for cabin heating when using CO2 heat pump system. Fig. 9 also shows that the COP was in the range of 1.8–1.9 when return air ratio changed, which indicated that return air method has little impact on the heating efficiency of EV CO2 heat pump system.
4.3. Other discussion Cabin return air utilization is a promising approach for EVs to reduce heating loads and save battery power, so a certain percentage of return air will be considered for mobile heat pump system. Under return air conditions, we predict that the SGC system is still superior to the OGC system, but the performance improvement may not be as good as that under 100% fresh air conditions. The increased heat transfer area of GC in SGC system could improve the heating performance all the time. But the utilization of return air will reduce the total heating loads, for example from 6.5 kW to 4 kW, which will lead to the reduction of performance benefit of SGC system. The CO2 heat pump system provides the air cooling and air heating functions for an EV. In this study, we pay emphasis on the heating performance improvement of the SGC heat pump system. In our future work, we will try to quantify the effects of cooling and heating on the energy consumption and driving range for an EV. The influencing parameters such as vehicle driving condition and weather conditions will be considered. Based on the quantification model, the comparison results of all year performance between a SGC CO2 heat pump system with an OGC CO2 heat pump system, and an R134a heat pump system could be further revealed. In addition, we obtained the experimental results of heating performance for the proposed SGC CO2 heat pump system at cold ambient temperatures of −10 °C and −20 °C in this study. Future studies are warranted to evaluate the steady and transient heating performance of this heat pump system under variable operation conditions for on-road vehicle applications. More experimental data to describe the performance of CO2 HXs and electrical compressor will be
4. Further discussion 4.1. System control of mobile CO2 heat pump system Quite a few literature have discussed the high pressure control by EXV adjustment in mobile CO2 AC system, and proposed experimental correlations between optimum pressure and temperature at GC outlet. However, the mobile CO2 heat pump system is significantly different from the mobile CO2 AC system in terms of operation conditions and cycle characteristics. Therefore, the correlations for calculating the optimum pressure in CO2 AC system can’t be directly applied to CO2 heat pump system. If the optimum pressure refers to COP maximization, the approach for mobile CO2 heat pump operation may be the same as mobile AC operation. This paper has revealed following important features in mobile CO2 heat pump system: the GC heat rejection amount is greatly improved with the increase of discharge pressure; the indoor air experiences a huge rise in temperature (for example, from −20 °C to 40 °C); and the compressor discharge temperature is close to limitation under extreme operation conditions. These features may help establish proper control strategy for mobile CO2 heat pump system in future research. Furthermore, the EXV opening needs to be regulated accurately both under subcritical and supercritical inlet pressure conditions. Different control strategies need to be carefully considered in 663
Applied Thermal Engineering 142 (2018) 656–664
D. Wang et al.
Acknowledgements
obtained, which helps to establish accurate prediction models of components and system. The experimental results in this paper would be employed to validate the model accuracy for steady operation. Using the prediction models, the heating results will be extended for other operation conditions, and the impacts of control parameters on system performance such as air flow rate and compressor speed will be deeply revealed.
Thanks to support by the Natural Science Foundation of China (No. 51776119) References [1] Z. Zhang, et al., Electric vehicle range extension strategies based on improved AC system in cold climate – a review, Int. J. Refrig. 88 (2018) 141–150. [2] F. Qin, et al., Experimental investigation and theoretical analysis of heat pump systems with two different injection portholes compressors for electric vehicles, Appl. Energy 185 (2017) 2085–2093. [3] Y. Higuchi et al., Efficient Heat Pump System for PHEV/BEV, SAE International, 2017. [4] J. Jung, et al., Numerical study of the effects of injection-port design on the heating performance of an R134a heat pump with vapor injection used in electric vehicles, Appl. Therm. Eng. 127 (2017) 800–811. [5] G. Lorentzen, J. Pettersen, A new, efficient and environmentally benign system for car air-conditioning, Int. J. Refrig. 16 (1) (1993) 4–12. [6] M.-H. Kim, J. Pettersen, C.W. Bullard, Fundamental process and system design issues in CO2 vapor compression systems, Prog. Energy Combust. Sci. 30 (2) (2004) 119–174. [7] K. Nawaz, et al., Performance optimization of CO2 heat pump water heater, Int. J. Refrig. 85 (2018) 213–228. [8] P. Gullo, B. Elmegaard, G. Cortella, Energy and environmental performance assessment of R744 booster supermarket refrigeration systems operating in warm climates, Int. J. Refrig. 64 (2016) 61–79. [9] S. Andersen et al., Comparative Manufacturing and Ownership Cost Estimates for Secondary Loop Mobile Air Conditioning Systems (SL-MACs), SAE International, 2017. [10] Daimler, https://www.daimler.com/sustainability/product/further-environmentaltechnologies/co2-air-conditioning-system.html (Daimler official website), 2017. [11] J.S. Brown, S.F. Yana-Motta, P.A. Domanski, Comparitive analysis of an automotive air conditioning systems operating with CO 2 and R134a, Int. J. Refrig. 25 (1) (2002) 19–32. [12] S.C. Kim, J.P. Won, M.S. Kim, Effects of operating parameters on the performance of a CO2 air conditioning system for vehicles, Appl. Therm. Eng. 29 (11–12) (2009) 2408–2416. [13] 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 (8) (2005) 1302–1307. [14] H. Hammer, A. Audi, Comparative study of AC-and HP-systems using the refrigerants R134a and R744, in: VDA Alternate RefrigerantWinter Meeting, 2002. [15] S.C. Kim, et al., Heating performance enhancement of a CO2 heat pump system recovering stack exhaust thermal energy in fuel cell vehicles, Int. J. Refrig. 30 (7) (2007) 1215–1226. [16] J. Liu, et al., Experimental study and numerical simulation Concerning fogging characteristics and Improvement of return air utilization for electric vehicles, Appl. Therm. Eng. 129 (2018) 1115–1123. [17] D. Wang, et al., Heating performance characteristics of CO2 heat pump system for electrical vehicle in a cold climate, Int. J. Refrig. 85 (2018) 27–41. [18] R.J. Moffat, Describing the uncertainties in experimental results, Exp. Therm. Fluid Sci. 1 (1) (1988) 3–17. [19] Y. Hou, et al., Experimental investigation on the influence of EEV opening on the performance of transcritical CO2 refrigeration system, Appl. Therm. Eng. 65 (1) (2014) 51–56. [20] C. Liu, et al., Mass flow characteristics and empirical modeling of R744 flow through electronic expansion device, Int. J. Refrig. 86 (2018) 82–88.
5. Conclusions This paper studied the heating performance of an EV CO2 heat pump system with a novel GC configuration. A newly developed mobile CO2 compressor was adopted to set up this CO2 heat pump system. The impacts of CO2 refrigerant charge, GC configuration and indoor return air ratio on heating characteristics of EV CO2 heat pump system have been revealed by experiments. The conclusions are summarized as follows. 1. At the temperature of −10 °C, the bigger refrigerant charge and smaller DSH control were beneficial to Tg,a,o and heating capacity improvements. The increased discharge pressure enhanced the heat rejection amount of CO2 GC. The refrigerant charge should be less than a certain value for EV CO2 heat pump system to prevent exceeding discharge temperature limitation. 2. The method of SGC greatly enhanced the heat rejection amount of GC. The improvements of heating capacity and COP for the EV CO2 heat pump system were up to 33% and 32%, respectively, when using SGC instead of OGC. At the extreme cold temperature of −20 °C, the novel EV CO2 heat pump system achieved 5.6 kW maximum heating capacity and 1.8 COP, which is much better than vapor injection heat pump system using conventional refrigerant. 3. The SGC CO2 heat pump system could successfully provide comfortable warm air and sufficient heating capacity to passenger cabin even under the conditions of extreme cold temperature (−20 °C) and severe heating loads (100% fresh air). The utilization of the SGC CO2 heat pump system greatly outperforms the conventional PTC heating supplement, which would significantly save electricity consumption and extend driving range of EVs. 4. The utilization of return air was an effective method to reduce heating loads and save electricity consumption for cabin heating when using CO2 heat pump system. And the return air method has little impact on the heating efficiency of EV CO2 heat pump system. In conclusion, this paper has proved CO2 mobile heat pump system is a promising heat pump technology for EVs with great heating advantages in a cold climate, and established a strong foundation for the future application of natural refrigerant CO2 in EVs.
664