- Email: [email protected]
Steady state and start-up performance characteristics of air source heat pump for cabin heating in an electric passenger vehicle
Accepted Manuscript Title: Steady state and start-up performance characteristics of air source heat pump for cabin heating in an electric passenger vehicle Author: Ho-Seong Lee, Moo-Yeon Lee PII: DOI: Reference:
S0140-7007(16)30185-2 http://dx.doi.org/doi: 10.1016/j.ijrefrig.2016.06.021 JIJR 3365
To appear in:
International Journal of Refrigeration
Received date: Revised date: Accepted date:
3-12-2015 11-5-2016 14-6-2016
Please cite this article as: Ho-Seong Lee, Moo-Yeon Lee, Steady state and start-up performance characteristics of air source heat pump for cabin heating in an electric passenger vehicle, International Journal of Refrigeration (2016), http://dx.doi.org/doi: 10.1016/j.ijrefrig.2016.06.021. This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting proof before it is published in its final form. Please note that during the production process errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.
Steady State and Start-up Performance Characteristics of Air Source Heat Pump for Cabin Heating in an Electric Passenger Vehicle Ho-Seong Lee1, Moo-Yeon Lee2, * 1
Thermal Management Research Center, KATECH, 74 Younjung-Ri, Pungse-Myun, Chonan, 330-912, Republic of Korea 2 School of Mechanical Engineering, Dong-A University, 37 Nakdong-Daero 550, Sahagu, Busan, Republic of Korea * Corresponding author. Tel.: +82-51-200-7659; fax +82-51-200-7656 E-mail address: [email protected] (M.-Y. Lee) Highlights
Many EVs suffer from a short driving range and shortage of cabin heating. The advanced air source heat pump system for cabin heating was investigated under various experimental conditions. The observed heating and capacity were 3.26 and 3.10 kW at an outdoor temperature of -10 oC. However, the observed transient performance of the considered system could be inferior to cabin heating load of electric passenger vehicles. Therefore, the suggested hybrid heating method under -10.0 oC could provide better thermal comfort. Abstract The heating COP and capacity for evaluating the steady state performance and the transient temperature for evaluating the cabin heating performance of an automotive air source heat pump system were investigated under various experimental conditions for an electric vehicle. To optimize refrigerant charge amount of tested heat pump system, the system was investigated under cooling modes and it was determined by 650g. The optimum EEV openings of heating modes were suggested with reasonable super-heating and heating capacity along with the ambient temperature. The heating COP and capacity were 3.26 and 3.10 kW at an outdoor temperature of -10 oC. The observed heating and transient performance of the automotive air source heat pump system mean it could be inferior to cabin heating load of electric passenger vehicles. Therefore, a hybrid heating method was suggested as a possible option for realization of longer driving ranges and 1 Page 1 of 36
thermal comfort for passengers.
Keywords: automotive heat pump, COP, heating capacity, hybrid heating, R-134a,
Nomenclature COP
coefficient of performance
EHX
exterior heat exchanger
i
enthalpy (kJ kg-1)
IC
inner condenser
Q
capacity (W)
s
entropy (kJ kg-1 K-1)
T
temperature (C, K)
compressor power (W)
W
Subscripts a
air
comp
compressor
in
inlet
out
outlet
ref
refrigerant
1. Introduction Fossil fuel energy has been used for a long time to power vehicles with internal combustion engines. Internal combustion engines should no longer be used for public transportation and private vehicles because of the world energy crisis and global warming (Kim et al, 2009a). Because of increasing international usage regulations on fossil fuels and environmental concerns to mitigate global warming and glacier melting, many automotive companies have developed zero-emission vehicles as an alternative to
2 Page 2 of 36
the internal combustion engine (ICE). So, many studies on the development of “green cars”, which do not use fossil fuels, have been conducted recently by company engineers and researchers for protection of the environment. Such vehicles have been developed by many automotive makers, although the classifications of green cars have not been officially defined internationally. In this study, green cars would be generally justified by their power source. Electric vehicles, fuel cell electric vehicles and hybrid electric vehicles could be widely classified to deal with environmental regulations. In order to minimize the power consumption for vehicle accessories such as air conditioner the considerable researches have been reported. Takahisa and katsuya (1996) developed dehumidifying heat pump system which consisted of an interior unit which separated recirculation air flow and fresh air flow, an exterior unit which controlled heat transfer and an electric-driven compressor. The developed system provided comfortable cabin air temperatures ranging from -10 oC to 40 oC. Brouwer et al. (2013) studied about the combined heat and power (CHP) installation under different EV (Electric Vehicle) charging patterns and household demands considering costs and emissions and concluded that centralized low-carbon power system may be best solution to meet the electricity demand of household with EVs in future compared to electricity from the grid. Consideration of high cost of EVs compared to ICEs has been also one of the hindrance to market growth of EVs. Alwai and Bradley (2013) developed model for plug-in hybrid electric vehicles based on total cost of ownership, payback and consumer preference. Although electric vehicles do not emit air pollutants, their limited driving range is a critical problem for commercialization and popularization. Doucette and McCulloch emphasized EVs’ (electric vehicles) represent one way to reduce the amount of GHG
3 Page 3 of 36
(greenhouse gas) emission compared with conventional vehicles based on ICE on the real road. One of the interesting finding from study conducted by Doucette and McCulloch (2011) showed that in countries with power generation mix with a high CO2 intensity like China, cumulative CO2 emissions averaged over the distance traveled for PHEVs are lesser than conventional vehicles and EVs. The EV and PHEV-20 achieved around 5% and 11% less CO2 emissions compared to conventional vehicles, respectively. Lee et al. (2011a) proposed mathematical model for predicting vehicle fuel consumption based on RPM and TPS (Throttle position sensor) with reasonable coefficient of determination. Wang et al. (2013) proposed the an extended Kalman filter method to estimate state of charge (SOC) for battery for EVs considering parameters like battery charge and discharge rate, temperature, self-discharge, aging. Kim and Lee (2014) suggested heat generation model for Li-ion battery system using the chemical reaction while charging and discharging using energy balance. The battery thermal management system under different operating conditions were discussed. The efficient heating of the cabin of an electric vehicle is a very important factor in preventing a reduction in driving range. In electric vehicles, heating devices are necessary to heat the cabin air because a high heat source like the engine of conventional vehicles does not exist. Generally, an electric heater of the PTC (positive temperature coefficient) type has been used in electric vehicles. An electric heater system has the advantage of low cost because it is not necessary to modify the established design and add additional devices. However, it could draw heavily from the battery due to the operation conversion characteristics of the electric PTC heater. This can result in a dramatically reduced driving range when the heater is operated. Therefore, effective heating systems for the cabins of electric vehicles are required, to
4 Page 4 of 36
minimize such reductions in range. The heat pump system has been considered as an alternative to the electric PTC heater for increasing heating efficiency. Jang et al. (2013) proposed continuous heating technology with high temperature and low temperature hot bypass method for defrosting which provided 17% higher heat capacity without any interruption as in reverse-cycle defrosting with increased cumulated energy efficiency of 8%. Aikins et al. (2013) reviewed recent works on two-stage heat pump systems for various applications. The review included two-stage cycle with intercooling, two-stage cycle with refrigerant injection and two-stage cascade system. Byrne et al. (2011a) conducted studies on performance of air-source heat pump system in heating, cooling and simultaneous modes and showed that performance increased by 16% when HPS is compared to standard reversible heat pump. In a subsequent study (2012b) by the same authors, a research was conducted on HPS operation with dynamic point of view. Lee et al. (2011b) investigated cooling performance of mobile air conditioning system using R744 instead of R-134a. The authors concluded that for system with R744, cooling capacity increased by 42.2% but COP decreased by 55.4% when compressor speed was increased from 900 rev/min to 1800 rev/min. Studies on the heat pump system for conventional vehicles were initially made. Antonijevic and Heckt reported that it was superior as an automotive heating unit to other heating solutions with respect to heating performance and fuel consumption (Antonijevic and Heckt, 2004). Hosoz and Direk investigated the operating characteristics of an R-134a heat pump system using an air source (Hosoz and Direk, 2006). The tested heat pump system provided sufficient heating performance in mild weather conditions but its heating capacity dropped rapidly with a decrease in outdoor
5 Page 5 of 36
temperature. The use of an additional heating device was required for the heating load (Kim et al, 2013). Also, few commercialized electric and hybrid electric vehicles of the VW e-Golf , Nissan Leaf and Toyota Prius PHEV have already implemented the heat pump system for cabin heating and cooling system. Also, the air source heat pump of few commercialized electric and hybrid electric vehicles enables to operate under ambient temperature of 0 oC. Previous studies on heat pump systems for heating vehicle cabins have focused on conventional vehicles using the internal combustion engine. However, there are few studies on heat pump systems for electric vehicles. Therefore, the automotive air source heat pump system for heating of the cabin in electric passenger vehicles was considered, with variations in driving conditions, including ambient temperature and compressor speeds. The performance characteristics based on heating capacity and the COP (coefficient of performance) of the heat pump system were analyzed and a hybrid heating method of the developed heat pump system was suggested to provide thermal comfort for passengers under extremely cold weather conditions. In addition, an exergy destruction analysis was simply performed to estimate the system efficiency approximately based on the measured data, taking into consideration the destruction rate of exergy for each component in the tested heat pump system.
2. Experimental method 2.1 Experimental setup Fig. 1 shows a schematic diagram of the test set-up for the heating performance and the exergy analysis of the automotive air source heat pump system for an electric vehicle. The psychrometric unit consists of two chambers to control temperature and humidity conditions for an exterior heat exchanger and interior heat exchanger (inner 6 Page 6 of 36
condenser), respectively. The psychrometric units were controlled by using a PID (proportional integral differential) control, and the two chambers were equipped with temperature and humidity devices which comprised a cooling coil, a heating coil and a humidifier, to control the temperature and humidity settings of the air with an accuracy of ± 0.2 oC. In the tested heat pump system, an electrically driven scroll compressor having internal oil separator with a displacement volume of 34.0 cm3 rev-1 at 350 V was installed and controlled by inverter driver. The POE (polyolester) oil was used. It was connected to the inner condenser, which was the parallel flow type. It was a louvered fin type aluminum heat exchanger of 37.0 mm, 156.2 mm and 232.0 mm in depth, height and width, respectively. The heat capacity and effectiveness of an interior heat exchanger were 3.1 kW and 0.656 at the volume flow rate of 250 m3 h-1 and 3.7 kW and 0.703 at the air flow rate of 400 m3 h-1, respectively, under the inlet air temperature of 0 o
C. The interior heat exchanger (inner condenser or IC) was located in the
psychrometric chamber to simulate the cabin of the electric vehicle. The parallel flow type interior heat exchanger was an aluminum louvered fin and tube of 37.0 mm, 156.2 mm, 232.0 mm in depth, height and width, respectively. The fin side specifications of the interior heat exchanger were 5.6 mm of the fin height and 0.91 mm of the fin pitch, respectively, with the overall heat transfer area of 3.83 m2. The tube side specifications of the interior heat exchanger were consisted of 2 distributed headers of 1.4 mm of tube height and 16.0 mm of tube width with the overall heat transfer area of 0.651 m2. Each header of the tube side had 13 channels with 0.9 mm in width and 0.9 mm in height, respectively. An electric expansion valve (EEV) with an orifice diameter of 1.6 mm was positioned in front of the exterior heat exchanger and controlled by EEV driver. The parallel flow type exterior heat exchanger was an aluminum louvered fin type of 21.0
7 Page 7 of 36
mm, 370.0 mm and 400.0 mm in depth, height and width, respectively. The fin side specifications of the exterior heat exchanger were 5.6 mm of the fin height and 0.91 mm of the fin pitch, respectively, with the overall heat transfer area of 6.22 m2. Outer size of the tube side of the exterior heat exchanger was 1.4 mm of tube height and 16.0 mm of tube width and the tube was consisted of 13 channels with 0.9 mm in width and 0.9 mm in height, respectively, with the overall heat transfer area of 0.936 m2. The heat capacity of an exterior heat exchanger was 15.6 kW at the air flow velocity of 5.0 m s-1 and an inlet air temperature of 37 oC. It was located in the psychrometric chamber to simulate the driving conditions of the vehicle. Finally, the exterior heat exchanger as an evaporator and the interior heat exchanger as an inner condenser in the considered automotive air source heat pump system for heating of the cabin in electric passenger vehicles were used and an evaporator was additionally considered with the cooling mode test for deciding the refrigerant charge amount of the tested system. The accumulator was located in front of the compressor. A sight glass was installed at the accumulator outlet and the gas and liquid phases of the refrigerant were checked before entering the compressor inlet. The installed PTC heater had a rated capacity of 6.0 kW. Table 1 shows the specifications of the automotive air source heat pump system. Table 2 shows the test conditions of the heating for the automotive air source heat pump system. The air temperatures for an interior side at a relative humidity of 50.0% was varied within a range of -10.0 °C to 5.0 °C with an interval of 5.0 °C. The air temperatures for an exterior side were ranged from -10.0 °C to 5.0 °C with an interval of 5.0 °C, and the air flow rate for an interior side was fixed at 300 m3 hr-1. Inlet air velocity of an interior side was set to 1.2 m s-1 and 2.2 m s-1. The compressor speeds were tested at both 2,000 and 3,500 rev min-1.
8 Page 8 of 36
2.2 Data analysis Table 3 shows the uncertainties of the experimental parameters and measured data. A mass flow meter was installed behind the interior heat exchanger to measure the mass flow rate of the refrigerant. The mass flow rate of the Coriolis type flow meter was measured with an accuracy of ± 0.2% of the reading. Thermocouples with an accuracy of ± 0.1 °C were used to measure the refrigerant-side and air-side temperatures, and digital pressure gauges with an accuracy of ± 0.1% of the full scale were installed to measure the refrigerant-side pressure. A data logger of a digital power meter type with an accuracy of ± 0.5% of the reading was used to measure the electric power consumed by both the compressor and the electric heater. In the exterior psychrometric a digital pressure transducer with an accuracy of ± 0.15% was installed to measure the pressure difference across the nozzle. The nozzle method (ASHRAE Standard 51, 1975) was utilized to measure the air flow rate. The precision limits and bias limits of all the parameters associated with heating capacity and COP were estimated in uncertainty analysis. The average uncertainties of the experimental data on heating capacity and the heating COP were 4.5% and 5.8%, respectively. Based on the measured data, the refrigerant-side heat transfer rate was calculated by equation (1) using the refrigerant enthalpy method (ASHRAE Standard 116, 1983). The air-side heat transfer rate was calculated by equation (2) using both the air flow rate and enthalpy difference suggested by (Kim et al, 2009b). The air-side heat transfer rate was validated against the refrigerant-side heat transfer rate. The heating COP of the tested air-source heat pump system was calculated by equation (3) using the compressor power consumed by compressor work. The compressor work was calculated using a power meter with an accuracy of ± 0.2% and the power input was measured exactly to evaluate the 9 Page 9 of 36
automotive air source heat pump system using R-134a. An exergy rate for the useful heat transfer rate of the inner condenser was calculated by equation (4), using measured data. Refrigerant flow rate, surrounding air-side temperatures, and refrigerant-side temperatures and pressures at inlet and outlet were considered.
Q ref m ref iref
(1)
Q a m a ( i a , in i a , o u t )
(2)
COP
Q
a
(3)
W e le c tr ic c o m p . ( i out i in ) E D m ref T o ( s out s in ) T a _ in
(4)
Fundamental heating performances for both steady state performance and the transient air temperatures for cabin heating of the designed automotive air source heat pump system were tested, and the exergy destruction rate for better effective design was considered for providing the warm-up speed and thermal comfort. Especially in electric vehicles using an electrically driven compressor for the cabin air conditioning system, the transient temperature and warm-up performance could be actively controlled to meet the cabin heating loads under various driving conditions, as mentioned in (Cho et al, 2012). The transient temperature performance during warm-up conditions for electric vehicles is an important factor in a designed heat pump system operated in cold weather for passenger comfort. In addition, the exergy destruction rate was calculated with assumptions of heat losses of the connecting lines due to the well insulation. Also, the exergy destruction rate of the evaporator was not considered because the heating 10 Page 10 of 36
performance characteristics of the air source heat pump for cabin heating in an electric passenger vehicle under heating mode were tested.
3. Results and discussion 3.1 Data accuracy and refrigerant charge test During the refrigerant charge test, the exterior and interior temperatures and air flow conditions of the automotive air source heat pump system for cabin heating were reflected in the normal operation and weather conditions, including extremely cold weather. A test set-up was built, to investigate heating performance and exergy destruction characteristics of an automotive air source heat pump system for cabin heating in electric vehicles. In order to evaluate the accuracy of the developed test equipment, an energy balance for both the air-side and refrigerant-side was performed using equation (3). The air-side heat transfer rates were compared with the results of the refrigerant-side heat transfer rate. The results showed a good agreement, within ± 5%, as shown in Fig. 2. Fig. 2 shows the validation data between the air-side heat transfer rate and refrigerant-side heat transfer rate of the automotive heat pump system with air source. Since the automotive heat pump system is supposed to operate in various driving and weather conditions, determination of the optimum refrigerant charge amount for the developed system is difficult. Theoretically, as the optimum refrigerant charge amount is determined, cooling capacity and heating capacity should be considered simultaneously, to minimize the unbalance problems between the heat capacity and the heat load under heating and cooling mode operations or switching, as mentioned in (Lee and Lee, 2013a). However, in a practical and experimental approach, the optimum refrigerant charge amount test is generally fulfilled under the cooling mode, because the refrigerant in the cooling mode generally flows more than in the heating 11 Page 11 of 36
mode due to operating conditions, as also mentioned in (Lee and Lee, 2013b). Fig. 3 shows the pressure and temperature test results of the automotive air source heat pump system with the refrigerant charge amounts. The discharge pressure of the compressor and outlet air averaged temperature of the evaporator of the tested heat pump system were implemented under compressor speeds of 2,000 and 3,500 rev min-1, respectively. The optimum refrigerant charge amount of the tested heat pump system for considering both low compressor speed (2,000 rev min-1) and high compressor speed (3,500 rev min-1) was determined at 650 g because of the proper sub-cooling and superheating temperatures as well as the effective cooling capacity, like vent temperature depicted in Fig. 3. This ensures sufficient heating capacity of the interior heat exchanger in the heat pump system because electric passenger vehicles are generally used in various driving and weather conditions as discussed in (Seo et al, 2013). The one optimized refrigerant charge amount of two data was selected based on providing the effective heating operations of all tested compressor operating ranges for the developed automotive air source heat pump system, although the refrigerant charge amount trends of two data were not exactly consistent, as shown in Figs. 3 (a) and (b). Fig. 4 shows the heating performance of the automotive air source heat pump system at the optimum EEV opening. The heating performance and superheating at the compressor suction were tested at an inlet temperature of the interior condenser of 5.0 o
C, an inlet temperature and air velocity on the exterior side of 5.0 oC and 2.2 m s-1,
respectively, and a compressor speed of 3,500 rev min-1, as shown in Fig. 4 (a). The heating capacity, the heating COP and the superheating were varied with the EEV openings. The EEV openings were determined with consideration of both the stable airside heating performance and the developed system reliability protections, such as
12 Page 12 of 36
preventing liquid compression of the compressor for the automotive air source heat pump system. In addition, the optimum EEV openings under various inlet temperatures of the interior condenser and compressor speeds were determined at an inlet temperature of the interior condenser of 5.0 oC. As shown in Fig. 4 (b), the optimum EEV openings were considered with the heating capacity, the heating COP, and reasonable super-heating for protecting the compressor. The optimum EEV openings with ambient temperatures were determined based on interior air temperature conditions yielding the maximum heating capacity and reasonable super-heating conditions over 5.0 °C at the compressor. Fig. 4 (b) shows the optimum openings of the EEV at compressor speeds of 2,000 and 3,500 rev min-1 under an inlet temperature of the interior condenser of 5.0 oC, the inlet temperature of the exterior heat exchanger of 5.0 o
C, and air velocities on the exterior side of 1.2 m s-1 and 2.2 m s-1, respectively.
3.2. Fundamental steady state and transient performances Fig. 5 shows the pressure and enthalpy diagram and air temperature of the interior heat exchanger of the air source heat pump system for cabin heating at compressor speeds of 2,000 and 3,500 rev min-1. As shown in Figs. 5(a) and 5(b), the vapor compression cycle for heating showed a comparatively stable performance with the air temperatures of the interior heat exchanger (inner condenser), compared with that of the conventional air conditioning system of an internal combustion engine, because the electric compressor driven by inverter was operated properly with variations of the heating loads. In addition, the developed heat pump system satisfied the requirement of super-heating over at least 5.0 oC between the compressor inlet and the outlet of the exterior heat exchanger for a stable vapor compression cycle operation (Lee et al, 2013).
13 Page 13 of 36
In addition, Fig. 5(c) shows the refrigerant mass flow rate of the automotive air source heat pump system with the inlet air temperatures of the inner condenser. The refrigerant mass flow rate with compressor speeds of 2,000 rev min-1 to 3,500 rev min-1 is related to the relative heat capacity. The refrigerant mass flow rate at the compressor speed of 3,500 rev min-1 increased 30.0% higher than that at the compressor speed of 2,000 rev min-1. It caused the heat capacity difference. As well as the stable performance of the pressure and enthalpy diagrams at Figs. 5(a) and 5(b), the transient temperature performances are also significant in providing the passengers with thermal comfort, as mentioned in (Cho et al, 2011). Fig. 5 (d) shows the temperature difference between inlet and outlet air of the interior heat exchanger, to represent interior heating capacity and find the satisfactory state to provide comfortable cabin heating with outdoor temperatures of -10.0 oC and 5.0 oC. Also in Fig. 5 (d), the interior air temperature performance for cabin heating was tested against outdoor temperatures. The interior air temperature was increased with time, and the warm-up speed at an outdoor temperature of 5.0 oC was better than that at -10.0 oC due to the decreased heating load and the increased heating capacity with an increase of the refrigerant flow rate. Also, the warm-up speed after five minutes was faster at an outdoor temperature of 5.0 oC due to increased refrigerant flow rate with higher refrigerant density at the compressor suction, which resulted in an exposed system line under a relatively higher temperature. In addition, as the designed heat pump system for cabin heating under an outdoor temperature of -10.0 oC is used, the additional heating source with the air source by interior heat exchanger (inner condenser) may be necessary to provide thermal comfort for passengers. Generally, as the heat pump system and air conditioning for providing cabin heating
14 Page 14 of 36
and cooling is designed, both the super-heating and the sub-cooling are critical parameters for stable evaporating performance and preventing liquid compression under various driving conditions. Therefore, the designed air source heat pump system for cabin heating could be used as a heating system for electric vehicles. Fig. 6 shows the heating performance and system operating characteristics of the automotive air source heat pump system with the compressor speeds of 2,000 and 3,500 rev min-1 under an interior air temperature of 5.0 oC and exterior air temperature of 5.0 o
C. In this study, the heating performances of the designed heat pump system increased
with the increase in compressor speeds, although the compressor work increased. This is because the rate of increase of compressor work was lower than that of the heating capacity. However, the heating COPs at all compressor speeds with given conditions decreased with the rise in ambient temperatures. As shown in Figs. 6 (b) and 6 (c), the heating capacity, compressor work, compressor discharge pressure and mass flow rate for analyzing the tested heat pump system characteristics at compressor speeds of 2,000 and 3,500 rev min-1 were measured with the variation of ambient temperatures. The heating capacities increased with the increase of ambient temperatures, but the heating COPs decreased due to the increased compressor work with the increase in both refrigerant mass flow rate and compressor discharge pressure, as shown in Fig. 6 (c). Also, the heating COPs at compressor speeds of 2,000 and 3,500 rev min-1 decreased by 29.4% and 20.0% as ambient temperature increased from -10.0 oC to 5.0 oC. This is because compressor efficiency decreased with the rise of ambient temperatures, as mentioned in (Lee et al, 2012). In addition, the refrigerant mass flow rates and the discharge pressures of the tested heat pump system with the rise of ambient
15 Page 15 of 36
temperatures from -10 oC to 5.0 oC increased by 76.5% and 108.2%, respectively, at a compressor speed of 2,000 rev min-1, and by 54.1% and 42.5%, respectively, at a compressor speed of 3,500 rev min-1, as shown in Fig. 6 (c).
3.3 Alternative hybrid heating system In order to simply analyze the irreversibility of the major components of the designed air source heat pump system, the exergy destruction rate was approximately analyzed at varied ambient temperatures under compressor speeds of 2,000 and 3,500 rev min-1 and exterior air velocity of 1.2, 2.2 m s-1. The exergy destruction rate of all components of the tested heat pump system increased with ambient temperatures, because increasing ambient temperature resulted in raising the refrigerant flow rate and decreasing the pressure ratio, from equation (4). As shown in Fig. 7 (a), the inner condenser had the largest exergy destruction compared with other components, and the rate of exergy destruction of the interior condenser increased by 61% as the ambient temperature increased from -10.0 oC to 5.0 o
C due to the increase of refrigerant flow rate. This result showed a good agreement
with the trend of heating capacity with the rise in ambient temperatures, as shown in Fig. 6 (b). Therefore, the exergy destruction rate of both the inner condenser and the compressor should be decreased for an effective automotive air source heat pump system. Especially, the system needs to improve the interior heat exchanger’s efficiency with consideration of the overall heat transfer area, the fin and tube configurations, and the fin efficiency. Generally, exergy destruction rate among the major components of the automotive air conditioning and the heat pump systems appears at the heat exchangers, including the evaporator and the condenser, because these were designed
16 Page 16 of 36
compactly and located at the limited space inside the HVAC system for vehicles, as mentioned in (Hosoz and Direk, 2006). As shown in Fig. 7 (b), the exergy destruction rates for the inner condenser with the compressor speeds of 2,000 and 3,500 rev min-1 were tested with the variation of ambient temperatures. The exergy destruction rate of the inner condenser at all compressor speeds increased with increasing ambient temperatures because of the combined effects of both the decreased enthalpy difference and the increased refrigerant flow rate of the inner condenser as the ambient temperature increases, considering equation (4). The exergy destruction rate at a compressor speed of 3,500 rev min-1 was on average 88.3% higher than at a compressor speed of 2,000 rev min-1. This is because the increased compressor speed leads to raised refrigerant flow and a higher condensing pressure with similar evaporating pressure. Higher condensing pressure at the inner condenser enlarges the temperature gap between the air and the refrigerant, so that the exergy destruction rate increases consequently. Because of the general lack of heating capacity for the air source heat pump system with a decrease in ambient temperature, an additional heat source is required to cope with the thermal comfort demand of the cabin heating for electric vehicles. The hybrid heating method with an electric PTC heater was carried out to achieve the required heating capacity and thermal comfort as the heat pump system operated under a heating load of 5.0 kW and ambient temperature of -10.0 o
C. Fig. 8 shows the hybrid heating method for the heat pump system operated at a
heating load of 5.0 kW and ambient temperature of -10 oC under the exterior air velocity of 1.2 m s-1 at a compressor speed of 2,000 rev min-1, and the exterior air velocity of 2.2 m s-1 at a compressor speed of 3,500 rev min-1. As shown in Fig. 8, the heating COPs of
17 Page 17 of 36
the air source heat pump system satisfying the given conditions decreased by 66.3 % and 47.5 % respectively at compressor speeds of 2,000 and 3,500 rev min-1, compared to when the heat pump system only was operated. However, the heating COPs for the hybrid heating system using both the air source and the electric PTC heater is deemed to have higher efficiency than only electric heater operation, which consumes the same amount of electric power to meet the heating load for cabin heating of the electric vehicle under severe cold conditions. Furthermore, even though the heating COP for the hybrid heating system at 2,000 rev min-1 decreased from 4.4 to 1.48 to satisfy the heating load at -10 oC, when compared to only electric heater usage with less than 1.0 of heat transfer efficiency, the hybrid heating system is considered the more effective due to its almost 50.0% higher efficiency. Therefore, the designed hybrid heating method is suitable for cabin heating by an automotive heat pump system in extremely cold weather conditions under 0 oC. In addition, the suggested hybrid heating method for electric vehicles, especially in cold weather conditions, has the advantage of providing heating to electrically driven vehicles without or less using of internal combustion engine and is widely considered as a possible option for the realization of longer driving ranges and thermal comfort for passengers. Also, advanced studies on the exergy analysis including exergy destruction rate between the refrigerant flow, the heat transfer process and the air flow for improving the automotive air source heat pump system efficiency would be reported in the next work.
4. Conclusions This study investigated the heating performances for evaluating the steady state
18 Page 18 of 36
performance and the transient temperature of an automotive air source heat pump system for electric vehicles under various experimental conditions, including variations of air temperature and air flow of both exterior and interior sides, and compressor speeds. Also, a hybrid heating method of the developed heat pump system was suggested to provide thermal comfort for passengers under extremely cold weather conditions. An experimental set-up was built and optimum charge amount with R-134a and EEV opening was investigated under appropriate operating conditions prior to performance analysis tests. The heating COPs at compressor speeds of 2,000 and 3,500 rev min-1 decreased by 29.4% and 20.0% as ambient temperatures increased from -10.0 o
C to 5.0 oC. The refrigerant mass flow rate and discharge pressure at the compressor of
the tested heat pump system increased by 76.5% and 108.2%, respectively, at a compressor speed of 2,000 rev min-1, and by 54.1% and 42.5%, respectively, at a compressor speed of 3,500 rev min-1, with the rise of the ambient temperatures from 10.0 oC to 5.0 oC. The exergy destruction rate was shortly analyzed and performed for decreasing the irreversibility of the major components to design an effective automotive air source heat pump system. The observed heating and transient performance characteristics of the automotive air source heat pump means it could be suitable for cabin heating of electric passenger vehicles. In addition, the suggested hybrid heating method for cabin heating of electric vehicles in cold weather conditions under -10.0 oC could be a possible option as a cabin heating system to provide better thermal comfort to passengers.
Acknowledgements This work was supported by the Dong-A University research fund.
19 Page 19 of 36
References Aikins, K.A., Lee, S.H., Choi, J.M. Technology review of two-stage vapor compression heat pump system. International Journal of Air-Conditioning and Refrigeration, (21) 2013, pp. 1330002. Al-Alawi, B.X., Bradley, T.H. Total cost of ownership, payback, and consumer preference modeling of plug-in hybrid electric vehicles. Applied Energy (103) 2013, pp. 488–506. Antonijevic, D., Heckt, R. Heat pump supplemental heating system for motor vehicles. Journal of Automobile Engineering, 218(10) 2004, pp.1111–1115. ASHRAE Standard 116. Methods of testing for seasonal efficiency of unitary airconditioners and heat pumps. American Society of Heating, Refrigerating and AirConditioning Engineers Atlanta (GA); 1983. ASHRAE Standard 51. Laboratory methods of testing fans for rating. Atlanta: American Society of Heating Refrigerating and Air-conditioning Engineers; 1975. Brouwer, A.S., Kuramochi, T., Broek, M., Faaij, A. Fulfilling the electricity demand of electric vehicles in the long term future: an evaluation of centralized and decentralized power supply systems. Applied Energy (107) 2013, pp. 33–51. Byrne, P., Miriel, J., Lent, Y. Experimental study of an air-source heat pump forsimultaneous
heating
and
cooling
–
Part
1:
Basic
concepts
and
performanceverification. Applied Energy (88) 2011a, pp. 1841–1847. Byrne, P., Miriel, J., Lent, Y. Experimental study of an air-source heat pump for simultaneous heating and cooling – Part 2: Dynamic behavior and two-phase thermosiphon defrosting technique. Applied Energy (88) 2011b, pp. 3072–3078. Cho, C.W., Lee, H.S., Won, J.P., Lee, M.Y. Measurement and evaluation of heating
20 Page 20 of 36
performance of heat pump system using wasted heat of electric devices for an electric bus. Energies (5) 2012, pp. 658-669. Cho, C.W., Lee, M.Y., Lee, H.S., Mo, J.H., Oh, S.T., Won, J.P. Performance characteristics of a heat pump system using coolant of electric devices for an online electric bus, 2011 society for design and process science SDPS-2011, 2011, pp. 1-6. Doucette, R.T., McCulloch, M.D., Modeling the prospects of plug-in hybrid electric vehicles to reduce CO2 emissions. Applied Energy (88) 2011, pp. 2315–2323. Hosoz, M., Direk, M. Performance evaluation of an integrated automotive air conditioning and heat pump system. Energy Convers Manage, 4 (47) 2006, pp. 545–549. Jang, J.Y., Bae, H.H., Lee, S.J., Ha, M.Y. Continuous heating of an air-source heat pump during defrosting and improvement of energy efficiency. Applied Energy, (110) 2013, pp. 9–16. Kim, S.C., Won, J.P., Park, Y.S., Lim, T.W., Kim, M.S. Performance evaluation of a stack cooling system using CO2 air conditioning system in fuel cell vehicle. International Journal of Refrigeration (32) 2009a, pp. 70-77. Kim, S.C., Won, J.P., Kim, M.S. Effects of operating parameters on the performance of a CO2 air conditioning system for vehicles. Applied Thermal Engineering (29) 2009b, pp. 2408-2416. Kim, J.H., Min, Y.K., Kim, B.S. Is the PMV Index an indicator of human thermal comfort sensation. International Journal of Smart Home, 7(1) 2013, pp. 27-34. Kim, D.W., Lee, M.Y. Theoretical approach on the heating and cooling system design for an effective operation of Li-ion batteries for electric vehicles. Journal of the Korea Academia-Industrial Cooperation Society (15) 2014, pp. 2545-2552.
21 Page 21 of 36
Lee, M.G., Park, Y.K., Jung, K.K., Yoo, J.J. Estimation of fuel consumption using invehicle parameters. International Journal of u- and e- Service, Science and Technology, (4) 2011a, pp. 37-46. Lee, H.S., Cho, C.W., Won, J.P., Lee, M.Y. Study on cooling performance characteristics of air conditioning system using R744 for a passenger vehicle. Journal of the Korea Academia-Industrial Cooperation Society, (12) 2011b, pp. 1-7. Lee, H.S., Won, J.P., Cho, C.W., Kim, Y.C., Lee, M.Y. Heating performance characteristics of stack coolant source heat pump using R744 for fuel cell electric vehicles. Journal of Mechanical Science and Technology, (26) 2012, pp. 2065-2071. Lee, H.S., Lee, M.Y. Cooling performance characteristics on mobile air-conditioning system for hybrid electric vehicles. Advances in Mechanical Engineering 2013a, pp. 1-9. Lee, M.Y., Lee, D.Y. Review on conventional air conditioning, alternative refrigerants and CO2 heat pumps for vehicles. Advances in Mechanical Engineering 2013b, pp. 1-15. Lee, D.Y., Cho, C.W., Won, J.P., Park, Y.C., Lee, M.Y. Performance characteristics of mobile heat pump for a large passenger electric vehicle. Applied Thermal Engineering, 50(1) 2013, pp. 660-669. Seo, J.H., Kim, H.J., Jung, K.J., Kim, D.W., Lee, M.Y. Review of conventional air conditioning system for internal combustion engines. International Journal of AirConditioning and Refrigeration, (18) 2013, pp. 1-8. Takahisa, S., Katsuya, I. Air conditioning system for electric vehicle. Proceedings of SAE International Congress & Exposition 1996, pp. 960688. Wang, H., Liu, Y., Fu, H., Li, G. Estimation of state of charge of batteries for electric
22 Page 22 of 36
vehicles. International Journal of Control and Automation, (6) 2013, pp. 185-194.
Psychrometric
Psychrometric for cabin room
Pressure gage (P)
Conditioned air
Temperature gage (T) Ref. line for heating EEV Driver
Exterior heat exchanger
Inverter driver assembly
EEV Mass flow meter
Sight glass
Interior heat exchanger (Inner condenser)
Cabin room For electric vehicles
Accumulator
Power meter
Inverter driver
Compressor
Fig. 1 - Schematic diagram of the experimental set-up of automotive air source heat pump system.
23 Page 23 of 36
Heat capacity of the refrigerant side (kW)
4.0 Heat balance
+5.0%
3.5
-5.0%
3.0
2.5
2.0 2.0
2.5
3.0
3.5
4.0
Heat capacity of the air side (kW)
Fig. 2 - Validation data between air-side and refrigerant-side heat transfer rate.
24 Page 24 of 36
3
-1
Inlet air flow rate = 420 m hr
Pressure (bar)
Pressure at Comp.Dis Outlet air averaged temperature at evaporator Sub-cooling at TXVin
18
Super-heating at Comp.suc
50 45 40 35
16
30 25
14
o
20
-1
Temperature ( C)
Compressor Speed = 2,000 rev min
20 15
12
10 5
10 500
550
600
650
700
750
800
850
0 900
Refrigerant charge (g)
(a) Compressor speed of 2,000 rev min-1
3
-1
Inlet air flow rate = 420 m hr
-1
Compressor Speed = 3,500 rev min
Pressure (bar)
Super-heating at Comp.Suc
50 45 40 35 30
16
25 20
14
15
o
18
Pressure at Comp.Dis Outlet air averaged temperature at evaporator Sub-cooling at TXVIn
Temperature ( C)
20
10 12
5 0
10 500
550
600
650
700
750
800
-5 850
Refrigerant charge (g)
(b) Compressor speed of 3,500 rev min-1 Fig. 3 - Effects on the pressure and temperature of the automotive air source heat p ump system with refrigerant charge amounts.
25 Page 25 of 36
30
Heating capacity Coefficient of performance
o
Super-heating at Comp.Suc
4
20
3 10 2 o
Inlet air temperature of IC = 5 C
1
0
o
Inlet air temperature of EHX = 5 C -1
Inlet air velocity of EHX = 2.2 m s 0 20
Compressor speed = 3,500 rev min 25
30
35
40
Superheat at Comp. Suction ( C)
Heating Capacity (kW) Coefficient of Performance(COP)
5
45
-1
50
55
-10 65
60
EEV opening (%)
(a) Heating capacity and heating COP
50 Compressor speed -1
2,000 rev min
EEV opening (%)
40
-1
3,500 rev min
30
20
10 -1
Inlet air velocity of EHX = 1.2, 2.2 m s 0 -15
-10
-5
0
5
10
o
Air inlet temperature of IC ( C)
(b) Optimum EEV opening Fig. 4 - Relation between EEV opening and performance of automotive heat pu mp system.
26 Page 26 of 36
o
Inlet air temperature of EHX = 5, 0, -10 C -1
Inlet air velocity of EHX = 1.2 m s
Compressor speed = 2,000 rev min
Pressure (MPa)
10
-1
Inlet air temperature of IC o
5C o
0C o
-10 C 1
0.1 150
200
250
300
350
400
450
500
-1
Enthalpy (kJ kg )
(a) Compressor speed of 2,000 rev min-1
o
Inlet temperature of EHX = 5, 0, -10 C Inlet air velocity of EHX = 2.2 m s
Compressor rotational speed = 3,500 rev min
10
Pressure (MPa)
-1 -1
Inlet air temperature of IC o
5C o
0C o
-10 C 1
0.1 150
200
250
300
350
400
450
500
-1
Enthalpy (kJ kg )
(b) Compressor speed of 3,500 rev min-1
27 Page 27 of 36
-1
Refrigerant mass flow rate (kg hr )
100
o
Inlet temperature of EHX = 5, 0, -10 C -1
Inlet air velocity of EHX = 1.2, 2.2 m s Compressor speed
90 80
2,000 rev min
-1
3,500 rev min
-1
70 60 50 40 30 20 -15
-10
-5
0
5
10 o
Inlet air temperature of inner condenser ( C)
(c) Refrigerant mass flow rate with the compressor speed
70
o
Inlet air temperature of EHX = 5, 0, -10 C -1
Inlet air velocity of EHX = 1.2 m s
o
Temperature ( C)
60
Compressor speed = 2,000 rev min
-1
50
40
30 Inlet / outlet temperature difference of inner condenser
20
o
Outdoor temperature = -10 C o
Outdoor temperature = 5 C 10 0.0
1.0
2.0
3.0
4.0
5.0
6.0
7.0
8.0
9.0
10.0
Time (min)
(d) Temperature difference for cabin heating Fig. 5 - Pressure and enthalpy diagram and interior air temperature for automotive air source heat pump system.
28 Page 28 of 36
o
Inlet air temperature of IC = 5 C o
Inlet air temperature of EHX = 5 C -1
Inlet air velocity of EHX = 1.2, 2.2 m s
10
Pressure (MPa)
Compressor speed 2,000 rev min
-1
3,500 rev min
-1
1
0.1 150
200
250
300
350
400
450
500
-1
Enthalpy (kJ kg )
(a) Cycle operation characteristic
o
Inlet air temperature of EHX = 5, 0, -5, -10 C
Qheating(kW), Comp. Work(kW)
Heating capacity
5.0
2,000 rev min 3,500 rev min
-1
6.0
COP
Compressor work -1
2,000 rev min
-1
2,000 rev min
-1
3,500 rev min
-1
3,500 rev min
-1
5.0
4.0
4.0
3.0
3.0
2.0
2.0
1.0
1.0
0.0 -15
-10
-5
0
5
0.0 10
Coefficient of Performance(COP)
Inlet air velocity of EHX = 2.2 m s 6.0
-1
o
Ambient Temperature ( C)
(b) Heating performances with ambient temperature
29 Page 29 of 36
150
Compressor speed = 2,000 rev min
-1
30
Compressor speed = 3,500 rev min Refrigerant mass flow rate
-1
25
Compressor speed = 2,000 rev min
-1
Compressor speed = 3,500 rev min
-1
-1
Compressor discharge pressure
120
20
90
15 60 10 5
-10
-5
0
Refrigerant mass flow rate (kg h )
Compressor Discharge Pressure (bar)
35
5
30
o
Temperature ( C)
(c) Compressor discharge pressure and refrigerant mass flow rate Fig. 6 - Heating performances and system operating characteristics of the automotive air source heat pump system, with compressor speed.
30 Page 30 of 36
o
Inlet temperature of EHX = 5, 0, -10 C -1
Inlet air velocity of EHX = 2.2 m s
Compressor speed = 3,500 rev min
Rate of exergy destruction (W)
800
-1
Compressor Inner condenser EEV Exterior heat exchanger
700 600 500 400 300 200 100 0
-10
-5
0
5
o
Temperature ( C)
(a) Exergy destruction rate of the major components
Rate of exergy destruction (W)
800
Compressor speed -1
-1
-1
-1
2,000 rev min (Inlet air velocity of EHX = 1.2 m s ) 700
3,500 rev min (Inlet air velocity of EHX = 2.2 m s )
600 500 400 300 200 100 -10
-5
0
5
o
Temperature ( C)
(b) Exergy destruction rate with compressor speeds
31 Page 31 of 36
o
Inlet air temperature of inner-condenser : -10 C o
Inlet air temperature of EHX : -10 C -1
-1
-1
-1
Inlet air velocity of EHX : 1.2 m s @ 2,000 rev min
140
Inlet air velocity of EHX : 2.2 m s @ 3,500 rev min
o
Temperature ( C)
120
Compressor speed -1
2,000 rev min ,
3,500 rev min
-1
100 80 60 9.9 bar
40 7.9 bar
20 0 -20 -40 0.6
0.8
0.2
0.4
0.6
1.0
1.2
1.4
0.8
1.6 -1
1.8
2.0
-1
Entropy (kJ kg K )
(c) T-S diagrams with the compressor speeds Fig. 7 - Exergy destruction rate of the automotive air source heat pump system with variation of ambient temperature.
32 Page 32 of 36
o
Inlet air temperature of IC = -10 C o
Inlet air temperature of EHX = -10 C -1
-1
-1
-1
Inlet air velocity of EHX = 1.2 m s @ 2,000 rev min ,
Heating capacity (kW), Compressor Work(kW), Coefficient of Performance (COP)
2.2 m s @ 3,500 rev min 8.0
Qtotal (kW)
Qheating (kW)
Powertotal (kW)
Compressor Work (kW) COPheating
COPtotal 6.0
required Heating Capacity
4.0
2.0
0.0
2,000 RPM with PTC
3,500 RPM with PTC
Compressor speed
Fig. 8 - Hybrid heating method for the heat pump system operated under heating load of 5.0 kW and ambient temperature of -10.0 oC.
33 Page 33 of 36
Table 1 - Specifications of the automotive air source heat pump system Specifications
Conditions
Compressor (Displacement, cm3)
Scroll type (34.0)
Interior heat exchanger
PF(parallel flow) type louvered-fin brazed-aluminum heat exchanger
(Size, mm)
232W x 156.2H x 37D Serpentine type louvered-fin brazedaluminum heat exchanger
Evaporator (Size, mm)
250W x 259.3H x 58D PF(parallel flow) type louvered-fin brazed-aluminum heat exchanger
Exterior heat exchanger (Size, mm)
400W x 370H x 21D
Expansion devices
Electric expansion valve (EEV)
Accumulator (m3)
0.00095
34 Page 34 of 36
Table 2 - Test conditions Components
Conditions
Compressor speed (rev min-1)
2000/ 3500
Interior heat exchanger (Inner condenser)
Exterior heat exchanger
Air flow rate (m3 hr-1)
300
Air temperature (oC)
5/ 0/ -5/ -10
Air velocity (m s-1)
1.2/ 2.2
Air temperature (oC)
5/ 0/ -5/ -10
Relative humidity (%)
50
Refrigerant
R-134a
Working fluid
Air
35 Page 35 of 36
Table 3 - Uncertainties of the experimental parameters and measured data Items
Uncertainties
Thermocouples (T-type)
±0.1 oC
Pressure gage (Sensors, PI3H)
±0.1% Full scale, Max 25 Mpa
Different pressure transducer (Sensors, EJX110)
±0.15% Full scale
Mass flow rate (Coriolis type)
±0.2% Full scale, Max 650 kg/h
Digital power meter (WT210)
+0.5% Reading
Data logger (Gantner)
E. Gate IP (V3), 2.93W @ 12.06 V
Heating COP
5.8%
Heating capacity
4.5%
36 Page 36 of 36
S0140-7007(16)30185-2 http://dx.doi.org/doi: 10.1016/j.ijrefrig.2016.06.021 JIJR 3365
To appear in:
International Journal of Refrigeration
Received date: Revised date: Accepted date:
3-12-2015 11-5-2016 14-6-2016
Please cite this article as: Ho-Seong Lee, Moo-Yeon Lee, Steady state and start-up performance characteristics of air source heat pump for cabin heating in an electric passenger vehicle, International Journal of Refrigeration (2016), http://dx.doi.org/doi: 10.1016/j.ijrefrig.2016.06.021. This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting proof before it is published in its final form. Please note that during the production process errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.
Steady State and Start-up Performance Characteristics of Air Source Heat Pump for Cabin Heating in an Electric Passenger Vehicle Ho-Seong Lee1, Moo-Yeon Lee2, * 1
Thermal Management Research Center, KATECH, 74 Younjung-Ri, Pungse-Myun, Chonan, 330-912, Republic of Korea 2 School of Mechanical Engineering, Dong-A University, 37 Nakdong-Daero 550, Sahagu, Busan, Republic of Korea * Corresponding author. Tel.: +82-51-200-7659; fax +82-51-200-7656 E-mail address: [email protected] (M.-Y. Lee) Highlights
Many EVs suffer from a short driving range and shortage of cabin heating. The advanced air source heat pump system for cabin heating was investigated under various experimental conditions. The observed heating and capacity were 3.26 and 3.10 kW at an outdoor temperature of -10 oC. However, the observed transient performance of the considered system could be inferior to cabin heating load of electric passenger vehicles. Therefore, the suggested hybrid heating method under -10.0 oC could provide better thermal comfort. Abstract The heating COP and capacity for evaluating the steady state performance and the transient temperature for evaluating the cabin heating performance of an automotive air source heat pump system were investigated under various experimental conditions for an electric vehicle. To optimize refrigerant charge amount of tested heat pump system, the system was investigated under cooling modes and it was determined by 650g. The optimum EEV openings of heating modes were suggested with reasonable super-heating and heating capacity along with the ambient temperature. The heating COP and capacity were 3.26 and 3.10 kW at an outdoor temperature of -10 oC. The observed heating and transient performance of the automotive air source heat pump system mean it could be inferior to cabin heating load of electric passenger vehicles. Therefore, a hybrid heating method was suggested as a possible option for realization of longer driving ranges and 1 Page 1 of 36
thermal comfort for passengers.
Keywords: automotive heat pump, COP, heating capacity, hybrid heating, R-134a,
Nomenclature COP
coefficient of performance
EHX
exterior heat exchanger
i
enthalpy (kJ kg-1)
IC
inner condenser
Q
capacity (W)
s
entropy (kJ kg-1 K-1)
T
temperature (C, K)
compressor power (W)
W
Subscripts a
air
comp
compressor
in
inlet
out
outlet
ref
refrigerant
1. Introduction Fossil fuel energy has been used for a long time to power vehicles with internal combustion engines. Internal combustion engines should no longer be used for public transportation and private vehicles because of the world energy crisis and global warming (Kim et al, 2009a). Because of increasing international usage regulations on fossil fuels and environmental concerns to mitigate global warming and glacier melting, many automotive companies have developed zero-emission vehicles as an alternative to
2 Page 2 of 36
the internal combustion engine (ICE). So, many studies on the development of “green cars”, which do not use fossil fuels, have been conducted recently by company engineers and researchers for protection of the environment. Such vehicles have been developed by many automotive makers, although the classifications of green cars have not been officially defined internationally. In this study, green cars would be generally justified by their power source. Electric vehicles, fuel cell electric vehicles and hybrid electric vehicles could be widely classified to deal with environmental regulations. In order to minimize the power consumption for vehicle accessories such as air conditioner the considerable researches have been reported. Takahisa and katsuya (1996) developed dehumidifying heat pump system which consisted of an interior unit which separated recirculation air flow and fresh air flow, an exterior unit which controlled heat transfer and an electric-driven compressor. The developed system provided comfortable cabin air temperatures ranging from -10 oC to 40 oC. Brouwer et al. (2013) studied about the combined heat and power (CHP) installation under different EV (Electric Vehicle) charging patterns and household demands considering costs and emissions and concluded that centralized low-carbon power system may be best solution to meet the electricity demand of household with EVs in future compared to electricity from the grid. Consideration of high cost of EVs compared to ICEs has been also one of the hindrance to market growth of EVs. Alwai and Bradley (2013) developed model for plug-in hybrid electric vehicles based on total cost of ownership, payback and consumer preference. Although electric vehicles do not emit air pollutants, their limited driving range is a critical problem for commercialization and popularization. Doucette and McCulloch emphasized EVs’ (electric vehicles) represent one way to reduce the amount of GHG
3 Page 3 of 36
(greenhouse gas) emission compared with conventional vehicles based on ICE on the real road. One of the interesting finding from study conducted by Doucette and McCulloch (2011) showed that in countries with power generation mix with a high CO2 intensity like China, cumulative CO2 emissions averaged over the distance traveled for PHEVs are lesser than conventional vehicles and EVs. The EV and PHEV-20 achieved around 5% and 11% less CO2 emissions compared to conventional vehicles, respectively. Lee et al. (2011a) proposed mathematical model for predicting vehicle fuel consumption based on RPM and TPS (Throttle position sensor) with reasonable coefficient of determination. Wang et al. (2013) proposed the an extended Kalman filter method to estimate state of charge (SOC) for battery for EVs considering parameters like battery charge and discharge rate, temperature, self-discharge, aging. Kim and Lee (2014) suggested heat generation model for Li-ion battery system using the chemical reaction while charging and discharging using energy balance. The battery thermal management system under different operating conditions were discussed. The efficient heating of the cabin of an electric vehicle is a very important factor in preventing a reduction in driving range. In electric vehicles, heating devices are necessary to heat the cabin air because a high heat source like the engine of conventional vehicles does not exist. Generally, an electric heater of the PTC (positive temperature coefficient) type has been used in electric vehicles. An electric heater system has the advantage of low cost because it is not necessary to modify the established design and add additional devices. However, it could draw heavily from the battery due to the operation conversion characteristics of the electric PTC heater. This can result in a dramatically reduced driving range when the heater is operated. Therefore, effective heating systems for the cabins of electric vehicles are required, to
4 Page 4 of 36
minimize such reductions in range. The heat pump system has been considered as an alternative to the electric PTC heater for increasing heating efficiency. Jang et al. (2013) proposed continuous heating technology with high temperature and low temperature hot bypass method for defrosting which provided 17% higher heat capacity without any interruption as in reverse-cycle defrosting with increased cumulated energy efficiency of 8%. Aikins et al. (2013) reviewed recent works on two-stage heat pump systems for various applications. The review included two-stage cycle with intercooling, two-stage cycle with refrigerant injection and two-stage cascade system. Byrne et al. (2011a) conducted studies on performance of air-source heat pump system in heating, cooling and simultaneous modes and showed that performance increased by 16% when HPS is compared to standard reversible heat pump. In a subsequent study (2012b) by the same authors, a research was conducted on HPS operation with dynamic point of view. Lee et al. (2011b) investigated cooling performance of mobile air conditioning system using R744 instead of R-134a. The authors concluded that for system with R744, cooling capacity increased by 42.2% but COP decreased by 55.4% when compressor speed was increased from 900 rev/min to 1800 rev/min. Studies on the heat pump system for conventional vehicles were initially made. Antonijevic and Heckt reported that it was superior as an automotive heating unit to other heating solutions with respect to heating performance and fuel consumption (Antonijevic and Heckt, 2004). Hosoz and Direk investigated the operating characteristics of an R-134a heat pump system using an air source (Hosoz and Direk, 2006). The tested heat pump system provided sufficient heating performance in mild weather conditions but its heating capacity dropped rapidly with a decrease in outdoor
5 Page 5 of 36
temperature. The use of an additional heating device was required for the heating load (Kim et al, 2013). Also, few commercialized electric and hybrid electric vehicles of the VW e-Golf , Nissan Leaf and Toyota Prius PHEV have already implemented the heat pump system for cabin heating and cooling system. Also, the air source heat pump of few commercialized electric and hybrid electric vehicles enables to operate under ambient temperature of 0 oC. Previous studies on heat pump systems for heating vehicle cabins have focused on conventional vehicles using the internal combustion engine. However, there are few studies on heat pump systems for electric vehicles. Therefore, the automotive air source heat pump system for heating of the cabin in electric passenger vehicles was considered, with variations in driving conditions, including ambient temperature and compressor speeds. The performance characteristics based on heating capacity and the COP (coefficient of performance) of the heat pump system were analyzed and a hybrid heating method of the developed heat pump system was suggested to provide thermal comfort for passengers under extremely cold weather conditions. In addition, an exergy destruction analysis was simply performed to estimate the system efficiency approximately based on the measured data, taking into consideration the destruction rate of exergy for each component in the tested heat pump system.
2. Experimental method 2.1 Experimental setup Fig. 1 shows a schematic diagram of the test set-up for the heating performance and the exergy analysis of the automotive air source heat pump system for an electric vehicle. The psychrometric unit consists of two chambers to control temperature and humidity conditions for an exterior heat exchanger and interior heat exchanger (inner 6 Page 6 of 36
condenser), respectively. The psychrometric units were controlled by using a PID (proportional integral differential) control, and the two chambers were equipped with temperature and humidity devices which comprised a cooling coil, a heating coil and a humidifier, to control the temperature and humidity settings of the air with an accuracy of ± 0.2 oC. In the tested heat pump system, an electrically driven scroll compressor having internal oil separator with a displacement volume of 34.0 cm3 rev-1 at 350 V was installed and controlled by inverter driver. The POE (polyolester) oil was used. It was connected to the inner condenser, which was the parallel flow type. It was a louvered fin type aluminum heat exchanger of 37.0 mm, 156.2 mm and 232.0 mm in depth, height and width, respectively. The heat capacity and effectiveness of an interior heat exchanger were 3.1 kW and 0.656 at the volume flow rate of 250 m3 h-1 and 3.7 kW and 0.703 at the air flow rate of 400 m3 h-1, respectively, under the inlet air temperature of 0 o
C. The interior heat exchanger (inner condenser or IC) was located in the
psychrometric chamber to simulate the cabin of the electric vehicle. The parallel flow type interior heat exchanger was an aluminum louvered fin and tube of 37.0 mm, 156.2 mm, 232.0 mm in depth, height and width, respectively. The fin side specifications of the interior heat exchanger were 5.6 mm of the fin height and 0.91 mm of the fin pitch, respectively, with the overall heat transfer area of 3.83 m2. The tube side specifications of the interior heat exchanger were consisted of 2 distributed headers of 1.4 mm of tube height and 16.0 mm of tube width with the overall heat transfer area of 0.651 m2. Each header of the tube side had 13 channels with 0.9 mm in width and 0.9 mm in height, respectively. An electric expansion valve (EEV) with an orifice diameter of 1.6 mm was positioned in front of the exterior heat exchanger and controlled by EEV driver. The parallel flow type exterior heat exchanger was an aluminum louvered fin type of 21.0
7 Page 7 of 36
mm, 370.0 mm and 400.0 mm in depth, height and width, respectively. The fin side specifications of the exterior heat exchanger were 5.6 mm of the fin height and 0.91 mm of the fin pitch, respectively, with the overall heat transfer area of 6.22 m2. Outer size of the tube side of the exterior heat exchanger was 1.4 mm of tube height and 16.0 mm of tube width and the tube was consisted of 13 channels with 0.9 mm in width and 0.9 mm in height, respectively, with the overall heat transfer area of 0.936 m2. The heat capacity of an exterior heat exchanger was 15.6 kW at the air flow velocity of 5.0 m s-1 and an inlet air temperature of 37 oC. It was located in the psychrometric chamber to simulate the driving conditions of the vehicle. Finally, the exterior heat exchanger as an evaporator and the interior heat exchanger as an inner condenser in the considered automotive air source heat pump system for heating of the cabin in electric passenger vehicles were used and an evaporator was additionally considered with the cooling mode test for deciding the refrigerant charge amount of the tested system. The accumulator was located in front of the compressor. A sight glass was installed at the accumulator outlet and the gas and liquid phases of the refrigerant were checked before entering the compressor inlet. The installed PTC heater had a rated capacity of 6.0 kW. Table 1 shows the specifications of the automotive air source heat pump system. Table 2 shows the test conditions of the heating for the automotive air source heat pump system. The air temperatures for an interior side at a relative humidity of 50.0% was varied within a range of -10.0 °C to 5.0 °C with an interval of 5.0 °C. The air temperatures for an exterior side were ranged from -10.0 °C to 5.0 °C with an interval of 5.0 °C, and the air flow rate for an interior side was fixed at 300 m3 hr-1. Inlet air velocity of an interior side was set to 1.2 m s-1 and 2.2 m s-1. The compressor speeds were tested at both 2,000 and 3,500 rev min-1.
8 Page 8 of 36
2.2 Data analysis Table 3 shows the uncertainties of the experimental parameters and measured data. A mass flow meter was installed behind the interior heat exchanger to measure the mass flow rate of the refrigerant. The mass flow rate of the Coriolis type flow meter was measured with an accuracy of ± 0.2% of the reading. Thermocouples with an accuracy of ± 0.1 °C were used to measure the refrigerant-side and air-side temperatures, and digital pressure gauges with an accuracy of ± 0.1% of the full scale were installed to measure the refrigerant-side pressure. A data logger of a digital power meter type with an accuracy of ± 0.5% of the reading was used to measure the electric power consumed by both the compressor and the electric heater. In the exterior psychrometric a digital pressure transducer with an accuracy of ± 0.15% was installed to measure the pressure difference across the nozzle. The nozzle method (ASHRAE Standard 51, 1975) was utilized to measure the air flow rate. The precision limits and bias limits of all the parameters associated with heating capacity and COP were estimated in uncertainty analysis. The average uncertainties of the experimental data on heating capacity and the heating COP were 4.5% and 5.8%, respectively. Based on the measured data, the refrigerant-side heat transfer rate was calculated by equation (1) using the refrigerant enthalpy method (ASHRAE Standard 116, 1983). The air-side heat transfer rate was calculated by equation (2) using both the air flow rate and enthalpy difference suggested by (Kim et al, 2009b). The air-side heat transfer rate was validated against the refrigerant-side heat transfer rate. The heating COP of the tested air-source heat pump system was calculated by equation (3) using the compressor power consumed by compressor work. The compressor work was calculated using a power meter with an accuracy of ± 0.2% and the power input was measured exactly to evaluate the 9 Page 9 of 36
automotive air source heat pump system using R-134a. An exergy rate for the useful heat transfer rate of the inner condenser was calculated by equation (4), using measured data. Refrigerant flow rate, surrounding air-side temperatures, and refrigerant-side temperatures and pressures at inlet and outlet were considered.
Q ref m ref iref
(1)
Q a m a ( i a , in i a , o u t )
(2)
COP
Q
a
(3)
W e le c tr ic c o m p . ( i out i in ) E D m ref T o ( s out s in ) T a _ in
(4)
Fundamental heating performances for both steady state performance and the transient air temperatures for cabin heating of the designed automotive air source heat pump system were tested, and the exergy destruction rate for better effective design was considered for providing the warm-up speed and thermal comfort. Especially in electric vehicles using an electrically driven compressor for the cabin air conditioning system, the transient temperature and warm-up performance could be actively controlled to meet the cabin heating loads under various driving conditions, as mentioned in (Cho et al, 2012). The transient temperature performance during warm-up conditions for electric vehicles is an important factor in a designed heat pump system operated in cold weather for passenger comfort. In addition, the exergy destruction rate was calculated with assumptions of heat losses of the connecting lines due to the well insulation. Also, the exergy destruction rate of the evaporator was not considered because the heating 10 Page 10 of 36
performance characteristics of the air source heat pump for cabin heating in an electric passenger vehicle under heating mode were tested.
3. Results and discussion 3.1 Data accuracy and refrigerant charge test During the refrigerant charge test, the exterior and interior temperatures and air flow conditions of the automotive air source heat pump system for cabin heating were reflected in the normal operation and weather conditions, including extremely cold weather. A test set-up was built, to investigate heating performance and exergy destruction characteristics of an automotive air source heat pump system for cabin heating in electric vehicles. In order to evaluate the accuracy of the developed test equipment, an energy balance for both the air-side and refrigerant-side was performed using equation (3). The air-side heat transfer rates were compared with the results of the refrigerant-side heat transfer rate. The results showed a good agreement, within ± 5%, as shown in Fig. 2. Fig. 2 shows the validation data between the air-side heat transfer rate and refrigerant-side heat transfer rate of the automotive heat pump system with air source. Since the automotive heat pump system is supposed to operate in various driving and weather conditions, determination of the optimum refrigerant charge amount for the developed system is difficult. Theoretically, as the optimum refrigerant charge amount is determined, cooling capacity and heating capacity should be considered simultaneously, to minimize the unbalance problems between the heat capacity and the heat load under heating and cooling mode operations or switching, as mentioned in (Lee and Lee, 2013a). However, in a practical and experimental approach, the optimum refrigerant charge amount test is generally fulfilled under the cooling mode, because the refrigerant in the cooling mode generally flows more than in the heating 11 Page 11 of 36
mode due to operating conditions, as also mentioned in (Lee and Lee, 2013b). Fig. 3 shows the pressure and temperature test results of the automotive air source heat pump system with the refrigerant charge amounts. The discharge pressure of the compressor and outlet air averaged temperature of the evaporator of the tested heat pump system were implemented under compressor speeds of 2,000 and 3,500 rev min-1, respectively. The optimum refrigerant charge amount of the tested heat pump system for considering both low compressor speed (2,000 rev min-1) and high compressor speed (3,500 rev min-1) was determined at 650 g because of the proper sub-cooling and superheating temperatures as well as the effective cooling capacity, like vent temperature depicted in Fig. 3. This ensures sufficient heating capacity of the interior heat exchanger in the heat pump system because electric passenger vehicles are generally used in various driving and weather conditions as discussed in (Seo et al, 2013). The one optimized refrigerant charge amount of two data was selected based on providing the effective heating operations of all tested compressor operating ranges for the developed automotive air source heat pump system, although the refrigerant charge amount trends of two data were not exactly consistent, as shown in Figs. 3 (a) and (b). Fig. 4 shows the heating performance of the automotive air source heat pump system at the optimum EEV opening. The heating performance and superheating at the compressor suction were tested at an inlet temperature of the interior condenser of 5.0 o
C, an inlet temperature and air velocity on the exterior side of 5.0 oC and 2.2 m s-1,
respectively, and a compressor speed of 3,500 rev min-1, as shown in Fig. 4 (a). The heating capacity, the heating COP and the superheating were varied with the EEV openings. The EEV openings were determined with consideration of both the stable airside heating performance and the developed system reliability protections, such as
12 Page 12 of 36
preventing liquid compression of the compressor for the automotive air source heat pump system. In addition, the optimum EEV openings under various inlet temperatures of the interior condenser and compressor speeds were determined at an inlet temperature of the interior condenser of 5.0 oC. As shown in Fig. 4 (b), the optimum EEV openings were considered with the heating capacity, the heating COP, and reasonable super-heating for protecting the compressor. The optimum EEV openings with ambient temperatures were determined based on interior air temperature conditions yielding the maximum heating capacity and reasonable super-heating conditions over 5.0 °C at the compressor. Fig. 4 (b) shows the optimum openings of the EEV at compressor speeds of 2,000 and 3,500 rev min-1 under an inlet temperature of the interior condenser of 5.0 oC, the inlet temperature of the exterior heat exchanger of 5.0 o
C, and air velocities on the exterior side of 1.2 m s-1 and 2.2 m s-1, respectively.
3.2. Fundamental steady state and transient performances Fig. 5 shows the pressure and enthalpy diagram and air temperature of the interior heat exchanger of the air source heat pump system for cabin heating at compressor speeds of 2,000 and 3,500 rev min-1. As shown in Figs. 5(a) and 5(b), the vapor compression cycle for heating showed a comparatively stable performance with the air temperatures of the interior heat exchanger (inner condenser), compared with that of the conventional air conditioning system of an internal combustion engine, because the electric compressor driven by inverter was operated properly with variations of the heating loads. In addition, the developed heat pump system satisfied the requirement of super-heating over at least 5.0 oC between the compressor inlet and the outlet of the exterior heat exchanger for a stable vapor compression cycle operation (Lee et al, 2013).
13 Page 13 of 36
In addition, Fig. 5(c) shows the refrigerant mass flow rate of the automotive air source heat pump system with the inlet air temperatures of the inner condenser. The refrigerant mass flow rate with compressor speeds of 2,000 rev min-1 to 3,500 rev min-1 is related to the relative heat capacity. The refrigerant mass flow rate at the compressor speed of 3,500 rev min-1 increased 30.0% higher than that at the compressor speed of 2,000 rev min-1. It caused the heat capacity difference. As well as the stable performance of the pressure and enthalpy diagrams at Figs. 5(a) and 5(b), the transient temperature performances are also significant in providing the passengers with thermal comfort, as mentioned in (Cho et al, 2011). Fig. 5 (d) shows the temperature difference between inlet and outlet air of the interior heat exchanger, to represent interior heating capacity and find the satisfactory state to provide comfortable cabin heating with outdoor temperatures of -10.0 oC and 5.0 oC. Also in Fig. 5 (d), the interior air temperature performance for cabin heating was tested against outdoor temperatures. The interior air temperature was increased with time, and the warm-up speed at an outdoor temperature of 5.0 oC was better than that at -10.0 oC due to the decreased heating load and the increased heating capacity with an increase of the refrigerant flow rate. Also, the warm-up speed after five minutes was faster at an outdoor temperature of 5.0 oC due to increased refrigerant flow rate with higher refrigerant density at the compressor suction, which resulted in an exposed system line under a relatively higher temperature. In addition, as the designed heat pump system for cabin heating under an outdoor temperature of -10.0 oC is used, the additional heating source with the air source by interior heat exchanger (inner condenser) may be necessary to provide thermal comfort for passengers. Generally, as the heat pump system and air conditioning for providing cabin heating
14 Page 14 of 36
and cooling is designed, both the super-heating and the sub-cooling are critical parameters for stable evaporating performance and preventing liquid compression under various driving conditions. Therefore, the designed air source heat pump system for cabin heating could be used as a heating system for electric vehicles. Fig. 6 shows the heating performance and system operating characteristics of the automotive air source heat pump system with the compressor speeds of 2,000 and 3,500 rev min-1 under an interior air temperature of 5.0 oC and exterior air temperature of 5.0 o
C. In this study, the heating performances of the designed heat pump system increased
with the increase in compressor speeds, although the compressor work increased. This is because the rate of increase of compressor work was lower than that of the heating capacity. However, the heating COPs at all compressor speeds with given conditions decreased with the rise in ambient temperatures. As shown in Figs. 6 (b) and 6 (c), the heating capacity, compressor work, compressor discharge pressure and mass flow rate for analyzing the tested heat pump system characteristics at compressor speeds of 2,000 and 3,500 rev min-1 were measured with the variation of ambient temperatures. The heating capacities increased with the increase of ambient temperatures, but the heating COPs decreased due to the increased compressor work with the increase in both refrigerant mass flow rate and compressor discharge pressure, as shown in Fig. 6 (c). Also, the heating COPs at compressor speeds of 2,000 and 3,500 rev min-1 decreased by 29.4% and 20.0% as ambient temperature increased from -10.0 oC to 5.0 oC. This is because compressor efficiency decreased with the rise of ambient temperatures, as mentioned in (Lee et al, 2012). In addition, the refrigerant mass flow rates and the discharge pressures of the tested heat pump system with the rise of ambient
15 Page 15 of 36
temperatures from -10 oC to 5.0 oC increased by 76.5% and 108.2%, respectively, at a compressor speed of 2,000 rev min-1, and by 54.1% and 42.5%, respectively, at a compressor speed of 3,500 rev min-1, as shown in Fig. 6 (c).
3.3 Alternative hybrid heating system In order to simply analyze the irreversibility of the major components of the designed air source heat pump system, the exergy destruction rate was approximately analyzed at varied ambient temperatures under compressor speeds of 2,000 and 3,500 rev min-1 and exterior air velocity of 1.2, 2.2 m s-1. The exergy destruction rate of all components of the tested heat pump system increased with ambient temperatures, because increasing ambient temperature resulted in raising the refrigerant flow rate and decreasing the pressure ratio, from equation (4). As shown in Fig. 7 (a), the inner condenser had the largest exergy destruction compared with other components, and the rate of exergy destruction of the interior condenser increased by 61% as the ambient temperature increased from -10.0 oC to 5.0 o
C due to the increase of refrigerant flow rate. This result showed a good agreement
with the trend of heating capacity with the rise in ambient temperatures, as shown in Fig. 6 (b). Therefore, the exergy destruction rate of both the inner condenser and the compressor should be decreased for an effective automotive air source heat pump system. Especially, the system needs to improve the interior heat exchanger’s efficiency with consideration of the overall heat transfer area, the fin and tube configurations, and the fin efficiency. Generally, exergy destruction rate among the major components of the automotive air conditioning and the heat pump systems appears at the heat exchangers, including the evaporator and the condenser, because these were designed
16 Page 16 of 36
compactly and located at the limited space inside the HVAC system for vehicles, as mentioned in (Hosoz and Direk, 2006). As shown in Fig. 7 (b), the exergy destruction rates for the inner condenser with the compressor speeds of 2,000 and 3,500 rev min-1 were tested with the variation of ambient temperatures. The exergy destruction rate of the inner condenser at all compressor speeds increased with increasing ambient temperatures because of the combined effects of both the decreased enthalpy difference and the increased refrigerant flow rate of the inner condenser as the ambient temperature increases, considering equation (4). The exergy destruction rate at a compressor speed of 3,500 rev min-1 was on average 88.3% higher than at a compressor speed of 2,000 rev min-1. This is because the increased compressor speed leads to raised refrigerant flow and a higher condensing pressure with similar evaporating pressure. Higher condensing pressure at the inner condenser enlarges the temperature gap between the air and the refrigerant, so that the exergy destruction rate increases consequently. Because of the general lack of heating capacity for the air source heat pump system with a decrease in ambient temperature, an additional heat source is required to cope with the thermal comfort demand of the cabin heating for electric vehicles. The hybrid heating method with an electric PTC heater was carried out to achieve the required heating capacity and thermal comfort as the heat pump system operated under a heating load of 5.0 kW and ambient temperature of -10.0 o
C. Fig. 8 shows the hybrid heating method for the heat pump system operated at a
heating load of 5.0 kW and ambient temperature of -10 oC under the exterior air velocity of 1.2 m s-1 at a compressor speed of 2,000 rev min-1, and the exterior air velocity of 2.2 m s-1 at a compressor speed of 3,500 rev min-1. As shown in Fig. 8, the heating COPs of
17 Page 17 of 36
the air source heat pump system satisfying the given conditions decreased by 66.3 % and 47.5 % respectively at compressor speeds of 2,000 and 3,500 rev min-1, compared to when the heat pump system only was operated. However, the heating COPs for the hybrid heating system using both the air source and the electric PTC heater is deemed to have higher efficiency than only electric heater operation, which consumes the same amount of electric power to meet the heating load for cabin heating of the electric vehicle under severe cold conditions. Furthermore, even though the heating COP for the hybrid heating system at 2,000 rev min-1 decreased from 4.4 to 1.48 to satisfy the heating load at -10 oC, when compared to only electric heater usage with less than 1.0 of heat transfer efficiency, the hybrid heating system is considered the more effective due to its almost 50.0% higher efficiency. Therefore, the designed hybrid heating method is suitable for cabin heating by an automotive heat pump system in extremely cold weather conditions under 0 oC. In addition, the suggested hybrid heating method for electric vehicles, especially in cold weather conditions, has the advantage of providing heating to electrically driven vehicles without or less using of internal combustion engine and is widely considered as a possible option for the realization of longer driving ranges and thermal comfort for passengers. Also, advanced studies on the exergy analysis including exergy destruction rate between the refrigerant flow, the heat transfer process and the air flow for improving the automotive air source heat pump system efficiency would be reported in the next work.
4. Conclusions This study investigated the heating performances for evaluating the steady state
18 Page 18 of 36
performance and the transient temperature of an automotive air source heat pump system for electric vehicles under various experimental conditions, including variations of air temperature and air flow of both exterior and interior sides, and compressor speeds. Also, a hybrid heating method of the developed heat pump system was suggested to provide thermal comfort for passengers under extremely cold weather conditions. An experimental set-up was built and optimum charge amount with R-134a and EEV opening was investigated under appropriate operating conditions prior to performance analysis tests. The heating COPs at compressor speeds of 2,000 and 3,500 rev min-1 decreased by 29.4% and 20.0% as ambient temperatures increased from -10.0 o
C to 5.0 oC. The refrigerant mass flow rate and discharge pressure at the compressor of
the tested heat pump system increased by 76.5% and 108.2%, respectively, at a compressor speed of 2,000 rev min-1, and by 54.1% and 42.5%, respectively, at a compressor speed of 3,500 rev min-1, with the rise of the ambient temperatures from 10.0 oC to 5.0 oC. The exergy destruction rate was shortly analyzed and performed for decreasing the irreversibility of the major components to design an effective automotive air source heat pump system. The observed heating and transient performance characteristics of the automotive air source heat pump means it could be suitable for cabin heating of electric passenger vehicles. In addition, the suggested hybrid heating method for cabin heating of electric vehicles in cold weather conditions under -10.0 oC could be a possible option as a cabin heating system to provide better thermal comfort to passengers.
Acknowledgements This work was supported by the Dong-A University research fund.
19 Page 19 of 36
References Aikins, K.A., Lee, S.H., Choi, J.M. Technology review of two-stage vapor compression heat pump system. International Journal of Air-Conditioning and Refrigeration, (21) 2013, pp. 1330002. Al-Alawi, B.X., Bradley, T.H. Total cost of ownership, payback, and consumer preference modeling of plug-in hybrid electric vehicles. Applied Energy (103) 2013, pp. 488–506. Antonijevic, D., Heckt, R. Heat pump supplemental heating system for motor vehicles. Journal of Automobile Engineering, 218(10) 2004, pp.1111–1115. ASHRAE Standard 116. Methods of testing for seasonal efficiency of unitary airconditioners and heat pumps. American Society of Heating, Refrigerating and AirConditioning Engineers Atlanta (GA); 1983. ASHRAE Standard 51. Laboratory methods of testing fans for rating. Atlanta: American Society of Heating Refrigerating and Air-conditioning Engineers; 1975. Brouwer, A.S., Kuramochi, T., Broek, M., Faaij, A. Fulfilling the electricity demand of electric vehicles in the long term future: an evaluation of centralized and decentralized power supply systems. Applied Energy (107) 2013, pp. 33–51. Byrne, P., Miriel, J., Lent, Y. Experimental study of an air-source heat pump forsimultaneous
heating
and
cooling
–
Part
1:
Basic
concepts
and
performanceverification. Applied Energy (88) 2011a, pp. 1841–1847. Byrne, P., Miriel, J., Lent, Y. Experimental study of an air-source heat pump for simultaneous heating and cooling – Part 2: Dynamic behavior and two-phase thermosiphon defrosting technique. Applied Energy (88) 2011b, pp. 3072–3078. Cho, C.W., Lee, H.S., Won, J.P., Lee, M.Y. Measurement and evaluation of heating
20 Page 20 of 36
performance of heat pump system using wasted heat of electric devices for an electric bus. Energies (5) 2012, pp. 658-669. Cho, C.W., Lee, M.Y., Lee, H.S., Mo, J.H., Oh, S.T., Won, J.P. Performance characteristics of a heat pump system using coolant of electric devices for an online electric bus, 2011 society for design and process science SDPS-2011, 2011, pp. 1-6. Doucette, R.T., McCulloch, M.D., Modeling the prospects of plug-in hybrid electric vehicles to reduce CO2 emissions. Applied Energy (88) 2011, pp. 2315–2323. Hosoz, M., Direk, M. Performance evaluation of an integrated automotive air conditioning and heat pump system. Energy Convers Manage, 4 (47) 2006, pp. 545–549. Jang, J.Y., Bae, H.H., Lee, S.J., Ha, M.Y. Continuous heating of an air-source heat pump during defrosting and improvement of energy efficiency. Applied Energy, (110) 2013, pp. 9–16. Kim, S.C., Won, J.P., Park, Y.S., Lim, T.W., Kim, M.S. Performance evaluation of a stack cooling system using CO2 air conditioning system in fuel cell vehicle. International Journal of Refrigeration (32) 2009a, pp. 70-77. Kim, S.C., Won, J.P., Kim, M.S. Effects of operating parameters on the performance of a CO2 air conditioning system for vehicles. Applied Thermal Engineering (29) 2009b, pp. 2408-2416. Kim, J.H., Min, Y.K., Kim, B.S. Is the PMV Index an indicator of human thermal comfort sensation. International Journal of Smart Home, 7(1) 2013, pp. 27-34. Kim, D.W., Lee, M.Y. Theoretical approach on the heating and cooling system design for an effective operation of Li-ion batteries for electric vehicles. Journal of the Korea Academia-Industrial Cooperation Society (15) 2014, pp. 2545-2552.
21 Page 21 of 36
Lee, M.G., Park, Y.K., Jung, K.K., Yoo, J.J. Estimation of fuel consumption using invehicle parameters. International Journal of u- and e- Service, Science and Technology, (4) 2011a, pp. 37-46. Lee, H.S., Cho, C.W., Won, J.P., Lee, M.Y. Study on cooling performance characteristics of air conditioning system using R744 for a passenger vehicle. Journal of the Korea Academia-Industrial Cooperation Society, (12) 2011b, pp. 1-7. Lee, H.S., Won, J.P., Cho, C.W., Kim, Y.C., Lee, M.Y. Heating performance characteristics of stack coolant source heat pump using R744 for fuel cell electric vehicles. Journal of Mechanical Science and Technology, (26) 2012, pp. 2065-2071. Lee, H.S., Lee, M.Y. Cooling performance characteristics on mobile air-conditioning system for hybrid electric vehicles. Advances in Mechanical Engineering 2013a, pp. 1-9. Lee, M.Y., Lee, D.Y. Review on conventional air conditioning, alternative refrigerants and CO2 heat pumps for vehicles. Advances in Mechanical Engineering 2013b, pp. 1-15. Lee, D.Y., Cho, C.W., Won, J.P., Park, Y.C., Lee, M.Y. Performance characteristics of mobile heat pump for a large passenger electric vehicle. Applied Thermal Engineering, 50(1) 2013, pp. 660-669. Seo, J.H., Kim, H.J., Jung, K.J., Kim, D.W., Lee, M.Y. Review of conventional air conditioning system for internal combustion engines. International Journal of AirConditioning and Refrigeration, (18) 2013, pp. 1-8. Takahisa, S., Katsuya, I. Air conditioning system for electric vehicle. Proceedings of SAE International Congress & Exposition 1996, pp. 960688. Wang, H., Liu, Y., Fu, H., Li, G. Estimation of state of charge of batteries for electric
22 Page 22 of 36
vehicles. International Journal of Control and Automation, (6) 2013, pp. 185-194.
Psychrometric
Psychrometric for cabin room
Pressure gage (P)
Conditioned air
Temperature gage (T) Ref. line for heating EEV Driver
Exterior heat exchanger
Inverter driver assembly
EEV Mass flow meter
Sight glass
Interior heat exchanger (Inner condenser)
Cabin room For electric vehicles
Accumulator
Power meter
Inverter driver
Compressor
Fig. 1 - Schematic diagram of the experimental set-up of automotive air source heat pump system.
23 Page 23 of 36
Heat capacity of the refrigerant side (kW)
4.0 Heat balance
+5.0%
3.5
-5.0%
3.0
2.5
2.0 2.0
2.5
3.0
3.5
4.0
Heat capacity of the air side (kW)
Fig. 2 - Validation data between air-side and refrigerant-side heat transfer rate.
24 Page 24 of 36
3
-1
Inlet air flow rate = 420 m hr
Pressure (bar)
Pressure at Comp.Dis Outlet air averaged temperature at evaporator Sub-cooling at TXVin
18
Super-heating at Comp.suc
50 45 40 35
16
30 25
14
o
20
-1
Temperature ( C)
Compressor Speed = 2,000 rev min
20 15
12
10 5
10 500
550
600
650
700
750
800
850
0 900
Refrigerant charge (g)
(a) Compressor speed of 2,000 rev min-1
3
-1
Inlet air flow rate = 420 m hr
-1
Compressor Speed = 3,500 rev min
Pressure (bar)
Super-heating at Comp.Suc
50 45 40 35 30
16
25 20
14
15
o
18
Pressure at Comp.Dis Outlet air averaged temperature at evaporator Sub-cooling at TXVIn
Temperature ( C)
20
10 12
5 0
10 500
550
600
650
700
750
800
-5 850
Refrigerant charge (g)
(b) Compressor speed of 3,500 rev min-1 Fig. 3 - Effects on the pressure and temperature of the automotive air source heat p ump system with refrigerant charge amounts.
25 Page 25 of 36
30
Heating capacity Coefficient of performance
o
Super-heating at Comp.Suc
4
20
3 10 2 o
Inlet air temperature of IC = 5 C
1
0
o
Inlet air temperature of EHX = 5 C -1
Inlet air velocity of EHX = 2.2 m s 0 20
Compressor speed = 3,500 rev min 25
30
35
40
Superheat at Comp. Suction ( C)
Heating Capacity (kW) Coefficient of Performance(COP)
5
45
-1
50
55
-10 65
60
EEV opening (%)
(a) Heating capacity and heating COP
50 Compressor speed -1
2,000 rev min
EEV opening (%)
40
-1
3,500 rev min
30
20
10 -1
Inlet air velocity of EHX = 1.2, 2.2 m s 0 -15
-10
-5
0
5
10
o
Air inlet temperature of IC ( C)
(b) Optimum EEV opening Fig. 4 - Relation between EEV opening and performance of automotive heat pu mp system.
26 Page 26 of 36
o
Inlet air temperature of EHX = 5, 0, -10 C -1
Inlet air velocity of EHX = 1.2 m s
Compressor speed = 2,000 rev min
Pressure (MPa)
10
-1
Inlet air temperature of IC o
5C o
0C o
-10 C 1
0.1 150
200
250
300
350
400
450
500
-1
Enthalpy (kJ kg )
(a) Compressor speed of 2,000 rev min-1
o
Inlet temperature of EHX = 5, 0, -10 C Inlet air velocity of EHX = 2.2 m s
Compressor rotational speed = 3,500 rev min
10
Pressure (MPa)
-1 -1
Inlet air temperature of IC o
5C o
0C o
-10 C 1
0.1 150
200
250
300
350
400
450
500
-1
Enthalpy (kJ kg )
(b) Compressor speed of 3,500 rev min-1
27 Page 27 of 36
-1
Refrigerant mass flow rate (kg hr )
100
o
Inlet temperature of EHX = 5, 0, -10 C -1
Inlet air velocity of EHX = 1.2, 2.2 m s Compressor speed
90 80
2,000 rev min
-1
3,500 rev min
-1
70 60 50 40 30 20 -15
-10
-5
0
5
10 o
Inlet air temperature of inner condenser ( C)
(c) Refrigerant mass flow rate with the compressor speed
70
o
Inlet air temperature of EHX = 5, 0, -10 C -1
Inlet air velocity of EHX = 1.2 m s
o
Temperature ( C)
60
Compressor speed = 2,000 rev min
-1
50
40
30 Inlet / outlet temperature difference of inner condenser
20
o
Outdoor temperature = -10 C o
Outdoor temperature = 5 C 10 0.0
1.0
2.0
3.0
4.0
5.0
6.0
7.0
8.0
9.0
10.0
Time (min)
(d) Temperature difference for cabin heating Fig. 5 - Pressure and enthalpy diagram and interior air temperature for automotive air source heat pump system.
28 Page 28 of 36
o
Inlet air temperature of IC = 5 C o
Inlet air temperature of EHX = 5 C -1
Inlet air velocity of EHX = 1.2, 2.2 m s
10
Pressure (MPa)
Compressor speed 2,000 rev min
-1
3,500 rev min
-1
1
0.1 150
200
250
300
350
400
450
500
-1
Enthalpy (kJ kg )
(a) Cycle operation characteristic
o
Inlet air temperature of EHX = 5, 0, -5, -10 C
Qheating(kW), Comp. Work(kW)
Heating capacity
5.0
2,000 rev min 3,500 rev min
-1
6.0
COP
Compressor work -1
2,000 rev min
-1
2,000 rev min
-1
3,500 rev min
-1
3,500 rev min
-1
5.0
4.0
4.0
3.0
3.0
2.0
2.0
1.0
1.0
0.0 -15
-10
-5
0
5
0.0 10
Coefficient of Performance(COP)
Inlet air velocity of EHX = 2.2 m s 6.0
-1
o
Ambient Temperature ( C)
(b) Heating performances with ambient temperature
29 Page 29 of 36
150
Compressor speed = 2,000 rev min
-1
30
Compressor speed = 3,500 rev min Refrigerant mass flow rate
-1
25
Compressor speed = 2,000 rev min
-1
Compressor speed = 3,500 rev min
-1
-1
Compressor discharge pressure
120
20
90
15 60 10 5
-10
-5
0
Refrigerant mass flow rate (kg h )
Compressor Discharge Pressure (bar)
35
5
30
o
Temperature ( C)
(c) Compressor discharge pressure and refrigerant mass flow rate Fig. 6 - Heating performances and system operating characteristics of the automotive air source heat pump system, with compressor speed.
30 Page 30 of 36
o
Inlet temperature of EHX = 5, 0, -10 C -1
Inlet air velocity of EHX = 2.2 m s
Compressor speed = 3,500 rev min
Rate of exergy destruction (W)
800
-1
Compressor Inner condenser EEV Exterior heat exchanger
700 600 500 400 300 200 100 0
-10
-5
0
5
o
Temperature ( C)
(a) Exergy destruction rate of the major components
Rate of exergy destruction (W)
800
Compressor speed -1
-1
-1
-1
2,000 rev min (Inlet air velocity of EHX = 1.2 m s ) 700
3,500 rev min (Inlet air velocity of EHX = 2.2 m s )
600 500 400 300 200 100 -10
-5
0
5
o
Temperature ( C)
(b) Exergy destruction rate with compressor speeds
31 Page 31 of 36
o
Inlet air temperature of inner-condenser : -10 C o
Inlet air temperature of EHX : -10 C -1
-1
-1
-1
Inlet air velocity of EHX : 1.2 m s @ 2,000 rev min
140
Inlet air velocity of EHX : 2.2 m s @ 3,500 rev min
o
Temperature ( C)
120
Compressor speed -1
2,000 rev min ,
3,500 rev min
-1
100 80 60 9.9 bar
40 7.9 bar
20 0 -20 -40 0.6
0.8
0.2
0.4
0.6
1.0
1.2
1.4
0.8
1.6 -1
1.8
2.0
-1
Entropy (kJ kg K )
(c) T-S diagrams with the compressor speeds Fig. 7 - Exergy destruction rate of the automotive air source heat pump system with variation of ambient temperature.
32 Page 32 of 36
o
Inlet air temperature of IC = -10 C o
Inlet air temperature of EHX = -10 C -1
-1
-1
-1
Inlet air velocity of EHX = 1.2 m s @ 2,000 rev min ,
Heating capacity (kW), Compressor Work(kW), Coefficient of Performance (COP)
2.2 m s @ 3,500 rev min 8.0
Qtotal (kW)
Qheating (kW)
Powertotal (kW)
Compressor Work (kW) COPheating
COPtotal 6.0
required Heating Capacity
4.0
2.0
0.0
2,000 RPM with PTC
3,500 RPM with PTC
Compressor speed
Fig. 8 - Hybrid heating method for the heat pump system operated under heating load of 5.0 kW and ambient temperature of -10.0 oC.
33 Page 33 of 36
Table 1 - Specifications of the automotive air source heat pump system Specifications
Conditions
Compressor (Displacement, cm3)
Scroll type (34.0)
Interior heat exchanger
PF(parallel flow) type louvered-fin brazed-aluminum heat exchanger
(Size, mm)
232W x 156.2H x 37D Serpentine type louvered-fin brazedaluminum heat exchanger
Evaporator (Size, mm)
250W x 259.3H x 58D PF(parallel flow) type louvered-fin brazed-aluminum heat exchanger
Exterior heat exchanger (Size, mm)
400W x 370H x 21D
Expansion devices
Electric expansion valve (EEV)
Accumulator (m3)
0.00095
34 Page 34 of 36
Table 2 - Test conditions Components
Conditions
Compressor speed (rev min-1)
2000/ 3500
Interior heat exchanger (Inner condenser)
Exterior heat exchanger
Air flow rate (m3 hr-1)
300
Air temperature (oC)
5/ 0/ -5/ -10
Air velocity (m s-1)
1.2/ 2.2
Air temperature (oC)
5/ 0/ -5/ -10
Relative humidity (%)
50
Refrigerant
R-134a
Working fluid
Air
35 Page 35 of 36
Table 3 - Uncertainties of the experimental parameters and measured data Items
Uncertainties
Thermocouples (T-type)
±0.1 oC
Pressure gage (Sensors, PI3H)
±0.1% Full scale, Max 25 Mpa
Different pressure transducer (Sensors, EJX110)
±0.15% Full scale
Mass flow rate (Coriolis type)
±0.2% Full scale, Max 650 kg/h
Digital power meter (WT210)
+0.5% Reading
Data logger (Gantner)
E. Gate IP (V3), 2.93W @ 12.06 V
Heating COP
5.8%
Heating capacity
4.5%
36 Page 36 of 36