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Heating performance characteristics of a dual source heat pump using air and waste heat in electric vehicles
Applied Energy 119 (2014) 1–9
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Applied Energy journal homepage: www.elsevier.com/locate/apenergy
Heating performance characteristics of a dual source heat pump using air and waste heat in electric vehicles Jae Hwan Ahn, Hoon Kang, Ho Seong Lee, Hae Won Jung, Changhyun Baek, Yongchan Kim ⇑ Department of Mechanical Engineering, Korea University, Anam-Dong, Sungbuk-Ku, Seoul 136-713, Republic of Korea
h i g h l i g h t s An R134a heat pump with the dual heat source was investigated. Performance characteristics were analyzed by varying heat source condition. Performance of the dual source system was higher than that of the single source. Optimum operation methods were suggested to improve the heating performance.
a r t i c l e
i n f o
Article history: Received 2 September 2013 Received in revised form 18 December 2013 Accepted 23 December 2013
Keywords: Heat pump R134a Automobile Waste heat Performance
a b s t r a c t A heat pump has been considered as an alternative to an electric heater to increase the efficiency of the heating unit in electric vehicles. However, the heating performance of a single source heat pump has yet to be improved at low outdoor temperatures. This study investigates the feasibility of a dual source heat pump using both air and waste heat in electric vehicles. The performance of the dual source heat pump was measured at various operation modes: air source-only, waste heat-only and dual heat source. The heating performance of the dual source heat pump was higher than those of the single source heat pumps. In addition, an alternating single mode operation of air source-only and waste heat-only modes was proposed to improve performance at low outdoor air temperatures. Ó 2014 Elsevier Ltd. All rights reserved.
1. Introduction Electric vehicles (EVs) such as plug-in and fuel-cell vehicles have been developed to meet environmental regulations [1–5]. Electric vehicles are eco-friendly and emit no air-pollutants, but their short driving range is a critical problem. The maximum driving range of recent-developed electric vehicles is approximately 200 km, but it decreases by about 40% with the operation of the heating and cooling units [6]. Therefore, efficient heating and cooling units for the cabin should be developed to increase the driving range of electric vehicles. In this study, heat pumps are considered as alternative heating units to the conventional electric heaters to increase the heating efficiency. The heat pump can be converted to a cooling system with the adoption of 4-way valve by changing the refrigerant flow direction. The conventional cooling system such as a vapor compression cycle has been successfully adopted in electric vehicles with the effort increasing the system efficiency.
⇑ Corresponding author. Tel.: +82 2 3290 3366; fax: +82 2 921 5439. E-mail address: [email protected] (Y. Kim). 0306-2619/$ - see front matter Ó 2014 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.apenergy.2013.12.065
Therefore, it is required to investigate the heating performance of the heat pump using all possible heat sources in electric vehicles. Heat pumps have been widely used as heating devices for commercial and residential buildings because of their high energy efficiency [7–12]. In addition, the application of an air-source heat pump to an internal combustion engine vehicle has been investigated. Antonijevic and Heckt [13] reported that the heat pump was superior to other heating solutions with respect to heating performance and fuel consumption as an automotive heating unit. Hosoz and Direk [14] investigated the operating characteristics of an R134a heat pump using the air-source. The tested heat pump provided sufficient heating performance in mild weather conditions, but its heating capacity dropped rapidly with the decrease of the outdoor temperature, necessitating the use of a supplemental heating device for proper heating operations. The use of waste heat from batteries and electric devices to increase the heating performance of a heat pump has been studied. Kim et al. [15,16] investigated a CO2 heat pump that used partial heat from the stack coolant in fuel cell vehicles. Cho et al. [17] investigated the performance of an R134a heat pump that used waste heat from the electric devices in an electric bus. They
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Nomenclature COP cp CR DB EEV h IDHX _ m OD-EEV ODHX P q_ Q_
coefficient of performance specific heat (kJ kg1 k1) compression ratio dry bulb electronic expansion valve enthalpy (kJ kg1) indoor heat exchanger mass flow rate (kg s1) EEV at the outdoor heat exchanger inlet outdoor heat exchanger pressure (kPa) heat transfer rate (kW) volumetric flow rate (m3 s1) R ratio T temperature (°C) v specific volume (m3 kg1) V velocity (m s1) _ W work (kW) WB wet bulb WH-EEV EEV at the waste heat exchanger inlet
suggested that the heat pump was applicable as a cabin heating device. Lee et al. [18] investigated the performance of a CO2 heat pump that used the stack coolant as the heat source for fuel cell electric vehicles under cold weather conditions. However, the study on the performance of a dual source heat pump that combined air and waste heat for electric vehicles is very limited in the open literature. Previous studies on heat pumps for internal combustion engines and electric vehicles focused on heat pumps that used a single heat source, either air or waste heat. However, all possible heat sources in an electric vehicle must be used to enhance the heating performance of a heat pump at low outdoor temperatures. In this study, an R134a heat pump using dual heat sources of both air and waste heat was investigated to replace the conventional electric heaters used in electric vehicles. The heating performance of the heat pump was measured at various heat source operation modes: air source-only, waste heat-only, and dual heat source. The operating characteristics of the heat pump were analyzed at various heat source conditions, such as air velocity, air temperature, waste heat amount, and coolant flow rate. In addition, system operation methods were suggested to increase the COP and heating capacity of the heat pump.
2. Experimental setup and test procedure 2.1. Experimental setup Fig. 1 shows a schematic diagram of the test setup. The experimental setup was composed of two sections: a vapor compression heating cycle and a coolant loop. The vapor compression heating cycle using R134a was composed of a twin rotary compressor, an outdoor heat exchanger (ODHX), an indoor heat exchanger (IDHX), a waste heat exchanger (WHX), two electric expansion valves (EEVs), and an accumulator. The designed heating capacity of the test setup was 2.5 kW. The hermetic twin rotary compressor had a displacement volume of 24.0 cm3 rev1. The IDHX was a crossflow type, plate-fin brazed-aluminum heat exchanger of 100 mm in depth, 198 mm in height, and 147 mm in width. The ODHX was a parallel-flow type, louvered-fin brazed-aluminum heat
WHO WHX
waste heat-only waste heat exchanger
Subscripts a air c coolant d discharge eva evaporation h heating i inlet ID indoor o outlet OD outdoor ODHX outdoor heat exchanger s suction T total WH waste heat WHX waste heat exchanger
exchanger of 17 mm in depth, 350 mm in height, and 670 mm in width. The maximum capacity and efficiency of the ODHX were 2.3 kW and 0.28, respectively, at the outdoor air temperature of 7 °C and air velocity of 4.5 m s1. The WHX was a plate heat exchanger of 60 mm in depth, 370 mm in height, and 140 mm in width. The maximum capacity and efficiency of the WHX were 2.5 kW and 0.20, respectively, at the waste heat amount of 2.5 kW and coolant flow rate of 0.009 m3 min1. The two EEVs, which were used as expansion devices, had an orifice diameter of 1.6 mm. The coolant loop consisted of a pump, an electric heater, and a water bath, and was designed to simulate the thermal load from the batteries or electric devices in an electric vehicle. The coolant was 50% ethylene–glycol water solution. The thermal load in the coolant loop was adjusted by controlling the electric power to the heater. The heat pump for an electric vehicle has three heat source operation modes: air source-only, waste heat-only, and dual heat source. In the air source-only mode (Fig. 1(a)), the discharged refrigerant from the compressor flows into the IDHX used as a condenser. After heat rejection in the IDHX, the condensed liquid refrigerant enters the ODHX through the EEV at the ODHX inlet (OD-EEV). Consequently, the refrigerant evaporated in the ODHX flows into the compressor. In the waste heat-only mode (Fig. 1(b)), the condensed liquid refrigerant in the IDHX flows into the WHX (coolant-to-refrigerant heat exchanger) through the EEV at the WHX inlet (WH-EEV). The refrigerant evaporated in the WHX using the waste heat enters the compressor and then rejects heat in the IDHX, which converts the waste heat into heating capacity. In the dual heat source mode (Fig. 1(c)), the condensed refrigerant in the IDHX splits into the WHX and ODHX. The heats gained by the refrigerant from the coolant-to-refrigerant heat transfer in the WHX and the air-to-refrigerant heat transfer in the ODHX convert to heating capacity through the IDHX. The heat pump was experimented in the air-enthalpy calorimeter having two chambers: one for the IDHX and the other for the ODHX. The indoor chamber was designed to control the air temperature from 0 °C to 50 °C and relative humidity (RH) from 20% to 90%. The outdoor chamber was designed to control the air temperature from 20 °C and 50 °C and relative humidity (RH) from 20% to 90%. A resistance temperature sensor with an accuracy of
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OD-EEV
M
Nozzle P
P T P
Nozzle
ODHX P T
T H
IDHX
T H
H
Distributor T H
M
Distributor
Pressure sensor Temperature sensor Mass flow meter ` Humidify sensor
T H
P T
P T P T
Fan
P T
Fan Accumulator Compressor
M
WH-EEV P T
P T
P
Nozzle
Heater T
Distributor T H T H
IDHX
M V H
Pressure sensor Temperature sensor Mass flow meter ` Flow meter Humidify sensor
WHX
Pump
P T
P
V
P T P T P T
P T
Fan Accumulator Compressor
WH-EEV M
P T
P T
M
Nozzle
P
P
Nozzle
OD-EEV
P T
T H
WHX Pump P T
T H
T H
P T
P T
P T
Fan
ODHX
`
T H
Distributor
Distributor
IDHX
Heater
Fan
V
P T
P T
Accumulator Compressor
Fig. 1. Schematic diagram of the experimental setup.
±0.2 °C was used to measure the refrigerant temperature, and a digital pressure transducer with an accuracy of ±0.5% was utilized to measure the refrigerant pressure. A mass flow meter with an accuracy of ±0.2% was used to measure the mass flow rate of the refrigerant. A digital power meter with an accuracy of ±0.5% was applied to measure the powers to the compressor and the electric
heater. A magnetic flow meter with an accuracy of ±0.35% was utilized to measure the coolant flow rate. A digital power meter with an accuracy of ±0.2% was used to measure the power to the pump. A digital pressure transducer with an accuracy of ±0.3% was applied to measure the pressure difference across the nozzle. The nozzle method [19] was utilized to measure the air flow rate.
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2.2. Test procedure
COP ¼
The optimum refrigerant charge amount was determined as 1700 g based on preliminary tests, yielding the maximum COP at the standard heating test condition [20] and compressor frequency of 60 Hz. Table 1 shows the test conditions for the heat pump. The indoor air conditions were fixed: dry bulb temperature of 20 °C, wet bulb temperature of 15 °C, and air flow rate of 7.0 m3 min1. In the air source-only mode, the operating characteristics of the heat pump were measured as the outdoor air temperature was varied from 10 °C to 7 °C, and the air velocity from 0.5 to 4.5 m s1. The maximum air velocity was determined as 4.5 m s1 because the effects of the air velocity on the heating capacity became negligible as the air velocity increased beyond 4.5 m s1. In the waste heat-only mode, the operating characteristics of the heat pump were measured as the waste heat amount was varied from 1.0 to 2.5 kW and the coolant flow rate from 0.0015 to 0.012 m3 min1. The waste heat amount is dependent on the power and driving conditions of an electric vehicle. In this study, the range of the waste heat amount was determined to cover wide driving conditions. The waste heat amount of 1.0 kW corresponded to the minimum value at low battery load condition, while the value of 2.5 kW corresponded to the target heating capacity. In the dual heat source mode, the operating characteristics of the heat pump were measured as the outdoor temperature was varied from -10 °C to 7 °C and the waste heat amount from 1.0 to 2.5 kW. The superheats at the outlets of the ODHX and WHX were controlled at 5 °C by modulating the EEV openings. As given in Eq. (1), the heating capacity of the heat pump was evaluated by the air flow rate and air temperature difference across the IDHX [21] where only sensible heat transfer occurred. As given in Eq. (2), the air-side heat transfer rate in the ODHX was evaluated by using the air flow rate and air enthalpy difference across the ODHX where both sensible and latent heat transfer occurred. The air enthalpies were evaluated by measuring the temperatures (DB, WB) and pressure of the humid air at the inlet and outlet of the ODHX. As given in Eq. (3), the heat transfer rate in the WHX was evaluated by using the coolant flow rate and temperature difference across the WHX. The COP of the heat pump was evaluated by Eq. (4). In addition, the uncertainties of the COP and heating capacity were estimated as ±3.7% and ±3.6%, respectively [22]. The estimated bias and precision errors for the COP were 2.35 and 2.87, respectively.
q_ h ¼
Q_ a
v
q_ ev a ¼
q_ w ¼
cp ðT a;o T a;i Þ
Q_ a
ðha;o ha;i Þ
ð2Þ
cp ðT c;o T c;i Þ
ð3Þ
v
Q_ c
vc
ð1Þ
q_ h _ W
ð4Þ
3. Results and discussion 3.1. Performance comparison in single heat source operation modes The operating characteristics of the heat pump were compared between single heat source modes: air source-only and waste heat-only modes. Fig. 2 represents the suction pressure, discharge pressure, and refrigerant flow rate of the heat pump in the air source-only mode. The suction and discharge pressures in the air source-only mode increased with increasing outdoor air velocity because of the increased overall heat transfer coefficient of the ODHX. It should be noted that the increasing rates of these pressures at air velocities below 2.0 m s1 were higher than those at higher air velocities because the increase of the convective heat transfer coefficient according to the air velocity was higher at lower air velocities [23]. In addition, the suction and discharge pressures in the air source-only mode increased with increasing outdoor air temperature due to the increased temperature difference between the refrigerant and air in the ODHX. The refrigerant mass flow rate through the compressor increased with increasing outdoor air velocity and air temperature because of the increased refrigerant density at the compressor inlet from the increased suction pressure. The increase of the refrigerant mass flow rate can provide higher COP and heating capacity in the air source-only mode due to the increased heat transfer rate in the ODHX. In the waste heat-only mode, the coolant was used to recover waste heat. The coolant was circulated by the pump in the closed loop, and this circulation consumed additional pump power. Fig. 3 represents the suction pressure, discharge pressure, and refrigerant flow rate of the heat pump in the waste heat-only mode. The discharge and suction pressures remained constant according to the coolant flow rate at a given waste heat amount because the refrigerant mass flow rate was constant regardless of the coolant flow rate through the WHX. Although the overall heat transfer coefficient in the WHX increased with increasing coolant flow rate, the heat transfer rate from the coolant remained constant because the temperature difference between the coolant and refrigerant decreased at a given waste heat amount. However, the discharge and suction pressures increased with increasing waste heat amount because the refrigerant mass flow rate increased linearly. Therefore, in the waste heat-only mode, the refrigerant mass flow rate and heat transfer rate in the WHX were strongly dependent on the amount of waste heat. Fig. 4 represents the work, heating capacity, and COP of the heat pump in the air source-only mode. The work and heating capacity in the air source-only mode increased with increasing outdoor air velocity and outdoor air temperature because of the increased refrigerant mass flow rate. The heating capacity satisfied the design value of 2.5 kW at the outdoor air temperature of 7 °C and outdoor air velocity of 4.5 m s1. The COP also increased with
Table 1 Test conditions. Parameters
Air source-only mode
Waste heat-only mode
Dual heat source mode
Indoor temperature (°C) Indoor air flow rate (m3 min1) Outdoor temperature (°C) Outdoor air velocity (m s1) Superheat at the ODHX outlet (°C) Waste heat transfer rate (kW) Coolant flow rate (m3 min1) Superheat at the WHX outlet (°C)
20 DB/15 WB 7.0 10, 0, 7 DB 0.5, 1.0, 2.0, 3.0, 4.5 5 – – –
20 DB/15 WB 7.0 7 – – 1.0, 1.5, 2.0, 2.5 0.0015, 0.003, 0.006, 0.009, 0.012 5
20 DB/15 WB 7.0 10, 0, 7 DB 4.5 5 1.0, 1.5, 2.0, 2.5 0.009 5
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350
1400
(a)
TOD=-10oC o
o
TOD= 0 C
TOD= 7oC
TOD= 0 C
TOD= 7oC
1200
(c)
TOD=-10oC
o
TOD= 0 C
300
80
(b)
TOD=-10oC
TOD= 7oC
60
200
-1
m (kg h )
Pd (kPa)
Ps (kPa)
250
1000
40
•
150 800
20
600
0
100 Air source-only mode
50 0
1
2
3
4
5
0
1
2
3
4
5
0
1
2
3
4
5
-1
Vair (ms ) Fig. 2. Variations of (a) suction pressure, (b) discharge pressure, and (c) refrigerant mass flow rate with air velocity and outdoor air temperature in the air source-only mode.
500
1600
(a)
•
qWH= 1.0 kW
•
qWH= 1.5 kW
•
qWH= 2.0 kW
qWH= 1.5 kW
•
1400
qWH= 2.0 kW
1300
qWH= 2.5 kW
•
80
qWH= 2.0 kW
70
qWH= 2.5 kW
•
Pd (kPa)
300
200
100
•
m (kg h-1)
•
qWH= 2.5 kW
Ps (kPa)
qWH= 1.0 kW
90
•
qWH= 1.5 kW
400
(c)
•
qWH= 1.0 kW
1500
•
100
(b)
•
1200 1100
60 50
•
1000
40
900
30
800
20
700
10
Waste heat-only mode
0 0
3
6
9
12
0
3
•
6
9
12
3
3
-1
0
3
6
9
12
Qcoolant x 10 (m min ) Fig. 3. Variations of (a) suction pressure, (b) discharge pressure, and (c) refrigerant mass flow rate with coolant flow rate and waste heat amount in the waste heat-only mode.
1.0 TOD=-10oC
(a)
3.0
TOD=-10oC
o
0.9
4.0
(b)
TOD=-10oC
o
TOD= 0 C
TOD= 0 C
TOD= 7oC
TOD= 7oC
2.5
(c)
o
TOD= 0 C
3.5
TOD= 7oC
0.7
•
0.6
(a)
0.5 0
1
2
2.0
COP
qh (kW)
•
W (kW)
3.0 0.8
1.5
2.5
2.0
1.0
1.5
0.5
1.0
Air source-only mode
3
4
5
0
1
2
3
4
5
0
1
2
3
4
5
Vair (ms -1) Fig. 4. Variations of (a) work, (b) heating capacity, and (c) COP with air velocity and outdoor air temperature in the air source-only mode.
increasing outdoor air velocity and outdoor air temperature because the increasing rate of the heating capacity was higher than that of the work. The COP increased by 19.3% with increasing
outdoor air velocity from 0.5 to 4.5 m s1 at the outdoor temperature of 0 °C. In addition, the COP increased by 49.6% with increasing outdoor air temperature from 10 °C to 0 °C at the outdoor air
J.H. Ahn et al. / Applied Energy 119 (2014) 1–9
velocity of 4.5 m s1. In the air source-only mode, the maximum COP was 3.1 at the outdoor air temperature of 7 °C and outdoor air velocity of 4.5 m s1. Fig. 5 shows the work, heating capacity, and COP of the heat pump in the waste heat-only mode. The work in the waste heatonly mode increased with increasing coolant flow rate at a given waste heat amount due to the increased pumping power, while the heating capacity remained constant. Therefore, the COP decreased slightly with increasing coolant flow rate at a given waste heat amount. In addition, both the work and heating capacity in the waste heat-only mode increased with increasing waste heat amount because of the increased refrigerant mass flow rate. The COP increased with increasing waste heat amount because the increasing rate of the work was lower than that of the heating capacity. At the coolant flow rate of 0.0015 m3 min1, the COP increased by 73% with increasing waste heat amount from 1.0 to 2.5 kW. In the waste heat-only mode, the maximum COP was 3.2 at the waste heat amount of 2.5 kW and coolant flow rate of 0.0015 m3 min1. Fig. 6(a) and (b) represents the heating capacities of the heat pump in the air source-only and waste heat-only modes according to the outdoor air temperature and waste heat amount, respectively. In the air source-only mode, the heating capacity was strongly dependent on the outdoor air temperature. As the outdoor air temperature decreased from 7 °C to 10 °C at the outdoor air velocity of 4.5 m s1, the heating capacity in the air source-only mode decreased from 2.5 to 0.9 kW, which was 36% of the design value. Therefore, in the air source-only mode, the heat pump may not operate properly at low outdoor air temperatures. In the waste heat-only mode, the heating capacity decreased linearly with decreasing waste heat amount. As the waste heat amount decreased from 2.5 to 1.0 kW, the heating capacity in the waste heat-only mode decreased from 2.8 to 1.3 kW, which was 52.4% of the design value. Therefore, the heating capacity should be increased by applying dual heat sources of both air and waste heat at low ambient temperature or low waste heat conditions.
3.5 3.0
•
1.5
1.0
1.0
-8
-4
0
1.5
5
qh (kW)
(b)
•
qWH= 1.5 kW •
qWH= 2.0 kW
4
•
qWH= 2.5 kW
qWH= 2.5 kW
COP
3.0
(c)
•
qWH= 1.0 kW
2.5
3
•
0.8
2.0
0.7
1.5
0.6
1.0
2
Waste heat-only mode
0
3
6
9
2.0
2.5
•
qWH (kW)
than those in the air source-only and waste heat-only modes. The increased suction pressure in the dual heat source mode led to increase the refrigerant mass flow rate. Fig. 8 represents the suction pressure, discharge pressure, and compression ratio of the dual source heat pump according to the waste heat amount and
•
qh (kW)
1.0
8
Fig. 6. Variations of heating capacities (a) with outdoor air temperature in the air source-only mode and (b) with waste heat amount in the waste heat-only mode.
•
0.9
4
TOD (o C)
•
•
W (kW)
0.5
qWH= 2.0 kW
•
Waste heat-only mode
Air source-only mode
0.5 -12
qWH= 1.5 kW
3.5
•
qWH= 2.5 kW
2.0
1.5
qWH= 1.0 kW
qWH= 2.0 kW
1.0
2.5
•
•
qWH= 1.5 kW
3
Fig. 7. Pressure-enthalpy diagrams of the heat pump cycles.
qWH= 1.0 kW
1.1
-1 3.0 Qcoolant= 0.009 m min
Vair= 4.5 m s
2.0
4.0
(a)
•
•
The performance of the heat pump in the dual heat source mode was analyzed as a function of the outdoor air temperature and waste heat amount. Fig. 7 shows the pressure-enthalpy diagram for the heat pump cycle in each operation mode at the outdoor temperature of 0 °C and waste heat amount of 1.5 kW. The suction and discharge pressures in the dual heat source mode were higher
•
(b)
TID= DB 20oC DB
-1
2.5
3.2. Performance improvement in the dual heat source mode
1.2
3.5
(a)
TID= 20oC DB
qh (kW)
6
12
1 0
•
3
6 3
9 3
12
0
3
6
9
12
-1
Qcoolantx 10 (m min ) Fig. 5. Variations of (a) work, (b) heating capacity, and (c) COP with coolant flow rate and waste heat amount in the waste heat-only mode.
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500
Ps (kPa)
350
TOD= 0oC
1400
TOD=-10 C
TOD=-10oC
TOD= 0oC
TOD= 0oC
TOD= 7 C
TOD= 7oC
WHO
WHO
o
300 250
(c)
TOD= 7oC
8
WHO
1200
CR (-)
400
TOD=-10 C
10
(b)
o
Pd (kPa)
450
1600
(a) o
1000
6
200 4
150
800
100 Dual heat source mode
600 0.0 0.5 1.0 1.5 2.0 2.5
50 0.0 0.5 1.0 1.5 2.0 2.5
2 0.0 0.5 1.0 1.5 2.0 2.5
•
qWH (kW) Fig. 8. Variations of (a) suction pressure, (b) discharge pressure, and (c) compression ratio with waste heat amount and outdoor air temperature in the dual heat source mode.
outdoor air temperature at the outdoor air velocity of 4.5 m s1 and coolant flow rate of 0.009 m3 min1. The suction and discharge pressures in the dual heat source mode increased with increasing waste heat amount at a given outdoor air temperature. At outdoor air temperatures of 0 °C and 7 °C, the suction and discharge pressures in the dual heat source mode were significantly higher than those in the waste heat-only mode because of the use of air source in the ODHX. However, at the outdoor air temperature of 10 °C, the suction and discharge pressures in the dual heat source mode were very close to those in the waste heat-only mode because the heat absorption from the air source was negligible at this lower outdoor air temperature. The evaporating temperature in the dual heat source mode increased from 16.3 °C to 8.9 °C with increasing waste heat amount from 1.0 to 2.5 kW at the outdoor air temperature of 10 °C. Therefore, the heat transfer rate in the ODHX (air source) decreased with increasing waste heat amount because of the decreased temperature difference between the outdoor air and refrigerant. In addition, the compression ratio of the compressor in the dual heat source mode decreased with increasing waste heat amount at a given outdoor air temperature because the increasing rate of the suction pressure was higher than that of the discharge pressure. Fig. 9 shows the refrigerant mass flow rates of the dual source heat pump according to the outdoor air temperature and waste
100 TOD=-10oC
100
(a)
TOD=-10oC
o
100
(b)
TOD= 0 C
TOD= 7oC
80
80
TOD= 7oC
60
-1
40
mT (kg h )
-1
mODHX (kg h )
-1
mWHX (kg h )
•
(c)
TOD= 7oC WHO
WHO
60
TOD=-10oC TOD= 0oC
o
TOD= 0 C
80
heat amount. The refrigerant mass flow rate through the WHX increased significantly with increasing waste heat amount at all outdoor air temperatures. The mass flow rate in the dual heat source mode was higher than that in the waste heat-only mode at outdoor air temperatures of 0 °C and 7 °C due to the increased suction pressure, although it was close to that in the waste heatonly mode at the outdoor air temperature of 10 °C. However, the mass flow rate through the ODHX decreased with the increase in the waste heat amount because the superheat had to be controlled at 5 °C with the increased evaporating temperature in the ODHX. Especially, at the outdoor air temperature of 10 °C, the refrigerant mass flow rate became zero when the waste heat amount increased by over 1.5 kW. The total refrigerant mass flow rate increased with increasing waste heat amount at all outdoor air temperatures due to the simultaneous use of air and water heat sources. At the outdoor air temperature of 10 °C, the total refrigerant mass flow rate with increasing waste heat amount became very close to that in the waste heat-only mode because of the negligible heat absorption from the air source at this low outdoor air temperature. Fig. 10 represents the evaporating capacity in the ODHX and the total evaporating capacity of the dual source heat pump according to the waste heat amount and outdoor air temperature. The trends of the evaporating capacities according to the waste heat amount
60
40
•
40
•
20
20
20 Dual heat source mode
0 0.0 0.5 1.0 1.5 2.0 2.5
0 0.0 0.5 1.0 1.5 2.0 2.5
0 0.0 0.5 1.0 1.5 2.0 2.5
•
qWH (kW) Fig. 9. Variations of (a) refrigerant flow rate through the WHX, (b) refrigerant flow rate through the ODHX, and (c) total refrigerant flow rate with waste heat amount and outdoor air temperature in the dual heat source mode.
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corresponded to those of the refrigerant mass flow rates (Fig. 8). The evaporating capacity in the ODHX in the dual heat source mode decreased linearly with increasing waste heat amount at all outdoor air temperatures because the temperature difference between the refrigerant and air in the ODHX decreased from the increased evaporating temperature with increasing waste heat amount. At the outdoor air temperature of 10 °C, the evaporating capacity in the ODHX became zero when the waste heat amount increased by over 1.5 kW because the evaporating temperature in the ODHX was higher than the outdoor air temperature. In addition, the total evaporating capacity in the ODHX and in the WHX increased with increasing waste heat amount and outdoor air temperature in the dual heat source mode. The total evaporating capacity in the dual heat source mode at the outdoor air temperature of 0 °C increased by 10.7% with increasing waste heat amount from 1.0 to 2.5 kW. At the outdoor air temperature of 10 °C, the total evaporating capacity in the dual heat source mode became very close to that in the waste heat-only mode with increasing waste heat amount, indicating no heat absorption from the air source. Fig. 11 shows the work, heating capacity, and COP of the dual source heat pump according to the waste heat amount and outdoor air temperature. The work in the dual heat source mode increased 4 4
TOD=-10oC
(a)
TOD=-10oC
(b)
TOD= 0oC
TOD= 0oC
TOD= 7oC
TOD= 7oC
3
qeva,T (kW)
3
qeva,ODHX (kW)
with increasing waste heat amount and outdoor air temperature due to the increased total mass flow rate. The work in the dual heat source mode was higher than those in the air source-only mode (qWH = 0) and waste heat-only mode (WHO) because of the higher total mass flow rate. However, at the outdoor air temperature of 10 °C, the work in the dual heat source mode approached that in the waste heat-only mode with increasing waste heat amount. The COP and heating capacity in the dual heat source mode also increased with increasing waste heat amount and outdoor air temperature due to the increased total mass flow rate. The COP and heating capacity in the dual heat source mode at the outdoor air temperature of 0 °C increased by 9.3% and 31.5%, respectively, with increasing waste heat amount from 0 to 2.5 kW. In addition, the COP and heating capacity in the dual heat source mode at the waste heat amount of 1.0 kW increased by 35.1% and 55.5%, respectively, with increasing outdoor air temperature from 10 °C to 0 °C. It should be noted that the COP and heating capacity in the dual heat source mode were higher than those in the air source-only mode (qWH = 0) and waste heat-only mode (WHO) because of the use of both air and waste heat as heat sources. However, at the outdoor air temperature of 10 °C, the COP and heating capacity in the dual heat source mode approached those in the waste heat-only mode with increasing waste heat amount because the evaporating capacity in the ODHX approached zero. Therefore, at low outdoor air temperatures, the heating performance of the dual source heat pump was strongly dependent on the waste heat amount. An alternating single mode operation of air source-only and waste heat-only modes was proposed to use the air source more effectively at low outdoor air temperatures because the heat absorption from the ODHX in the dual heat source mode was negligible. During the operation in the air source-only mode, the waste heat is stored in the coolant loop without any heat transfer to the heat pump. Consequently, in the waste heat-only mode, the heat pump receives more heat from the coolant loop at an elevated coolant temperature. Fig. 12(a) shows the air temperature at the IDHX outlet and coolant temperature at the WHX inlet according to time. In the air source-only mode, the coolant temperature increased with time, while the air temperature remained constant. Generally, an auxiliary heating such as a PTC heater is required to provide thermal comfort for the passengers at low outdoor temperatures because the heating capacity in the air source-only mode is not sufficient. As the heat pump started to operate in the waste heat-only mode, the air temperature increased due to the higher coolant temperature, even though the coolant temperature decreased with time. The average air temperature at the IDHX out-
2
WHO
2
•
•
1
1
0 0.0
0 0.0
Dual heat source mode
0.5
1.0
1.5
2.0
2.5
0.5
1.0
1.5
2.0
2.5
•
qWH (kW) Fig. 10. Variations of (a) evaporating capacity in the ODHX and (b) total evaporating capacity with waste heat amount and outdoor air temperature in the dual heat source mode.
TOD=-10oC
5
(a)
TOD=-10oC
TOD= 0 C
4
TOD= 7oC
•
0.8
q h (kW)
W (kW)
1.0
TOD= 0 C
TOD= 0 C
TOD= 7oC
3
2
(c)
o
4
WHO
WHO
•
TOD=-10oC
o
o
1.2
5
(b)
TOD= 7oC WHO
COP (-)
1.4
3
2 0.6
1
0.4
0 0.0 0.5 1.0 1.5 2.0 2.5
Dual heat source mode
0.0 0.5 1.0 1.5 2.0 2.5
1 0.0 0.5 1.0 1.5 2.0 2.5
•
qWH (kW) Fig. 11. Variations of (a) work, (b) heating capacity, and (c) COP with waste heat amount and outdoor air temperature in the dual heat source mode.
J.H. Ahn et al. / Applied Energy 119 (2014) 1–9
50
TOD= -10 C • qWH= 1.5 kW
30
T (oC)
130
(a)
(b)
o
20
Ta,o Tc,i
10 0
120
Performance ratio (%)
Air Waste source heat 40 -only -only
Acknowledgements
100
This research was supported by the Korean Evaluation Institute of Industrial Technology (No. 10035530) and the Human Resources Program in Energy Technology of the Korea Institute of Energy Technology Evaluation and Planning (KETEP) grant financial resource from the Ministry of Trade, Industry & Energy, Republic of Korea (No. 20124010203250).
90
80
70 10
20
30
40
50
peratures. At the outdoor air temperature of 10 °C and waste heat amount of 1.5 kW, the COP and heating capacity in the alternating single mode operation were 4.3% and 10.5%, respectively, higher than those in the dual heat source mode.
110
-10 0
9
•
qh
•
W
COP
Time (minute) Fig. 12. (a) Variations of air temperature at the IDHX outlet and coolant temperature at the WHX inlet with time and (b) ratio of heat pump performance in the alternating single mode operation to that in the dual heat source mode.
let was 1.2 °C higher than that in the dual heat source mode at a given waste heat amount. Fig. 12(b) shows the ratio of the heat pump performance in the alternating single mode operation to that in the dual heat source mode at the outdoor air temperature of 10 °C and waste heat amount of 1.5 kW. The heating capacity in the alternating single mode operation was 10.5% higher than that in the dual heat source mode. Even though the operation in the air-source only mode did not provide any benefit in the viewpoint of the total work in the heat pump and heater, the increased heating capacity in the waste heat-only mode led to reduce the total work. In addition, the COP was 4.3% higher than that in the dual heat source mode because the increase in the heating capacity was higher than the increase in the work. 4. Conclusions In this study, an R134a heat pump with dual heat sources of both air and waste heat was investigated experimentally to replace the conventional electric heaters used in electric vehicles. In the air source-only mode, the COP and heating capacity increased with increasing outdoor air velocity and outdoor air temperature. The heat pump in the air source-only mode yielded the design heating capacity at the outdoor temperature of 7 °C, but it provided approximately 36% of the design value at the outdoor air temperature of 10 °C. In the waste heat-only mode, the COP and heating capacity increased with increasing waste heat amount. The heating capacity in the waste heat-only mode satisfied the design value at the waste heat amount of 2.5 kW, but it was 52.4% of the design value at the waste heat amount of 1.0 kW. In the dual heat source mode, the COP and heating capacity increased with increasing waste heat amount and outdoor air temperature because of the increased total mass flow rate. The COP and heating capacity at the outdoor air temperature of 0 °C increased by 9.3% and 31.5%, respectively, with increasing waste heat amount from 0 to 2.5 kW. The heating performance in the dual heat source mode was higher than those in the air source-only and waste heat-only modes because of the use of both air and waste heat as heat sources. However, at the outdoor air temperature of 10 °C, the heating performance in the dual heat source mode was strongly dependent on the waste heat amount because the heat absorption from the air source was negligible. An alternating single mode operation of air source-only and waste heat-only modes was proposed to use the air source more effectively at low outdoor air tem-
References [1] Al-Alawi BM, Bradley TH. Total cost of ownership, payback, and consumer preference modeling of plug-in hybrid electric vehicles. Appl Energy 2013;103:488–506. [2] Brouwer AS, 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. Appl Energy 2013;107:33–51. [3] Foley A, Tyther B, Calnan P, Gallachoir BO. Impacts of electric vehicle charging under electricity market operations. Appl Energy 2013;101:93–102. [4] Kristoffersen TK, Capion K, Meibom P. Optimal charging of electric drive vehicles in a market environment. Appl Energy 2011;88:1940–8. [5] Doucette RT, McCulloch MD. Modeling the prospects of plug-in hybrid electric vehicles to reduce CO2 emissions. Appl Energy 2011;88:2315–23. [6] Farrington R, Rugh J. Impact of vehicle air-conditioning on fuel economy, tailpipe emissions, and electric vehicle range. NREL/CP-540-28960; 2000. [7] Jang JY, Bae HH, Lee SJ, Ha MY. Continuous heating of an air-source heat pump during defrosting and improvement of energy efficiency. Appl Energy 2013;110:9–16. [8] De Melo Reis RV, Oliveira RN, Machado L, Nassar Koury RN. Using a heat pump as an alternative to support solar collector for water heating in Brazil. Int J Air Cond Refrig 2012:20. 1250013. [9] Aikins KA, Lee SH, Choi JM. Technology review of two-stage vapor compression heat pump system. Int J Air Cond Refrig 2013;21:1330002. [10] Yang SH, Rhee JY. Utilization and performance evaluation of a surplus air heat pump system for greenhouse cooling and heating. Appl Energy 2013;105:244–51. [11] Byrne P, Miriel J, Lent Y. Experimental study of an air-source heat pump for simultaneous heating and cooling – Part 1: Basic concepts and performance verification. Appl Energy 2011;88:1841–7. [12] 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. Appl Energy 2011;88:3072–8. [13] Antonijevic D, Heckt R. Heat pump supplemental heating system for motor vehicles. Proc Inst Mech Eng 2004;218(1):1111–5. [14] Hosoz M, Direk M. Performance evaluation of an integrated automotive air conditioning and heat pump system. Energy Convers Manage 2006;47:545–59. [15] Kim SC, Kim MS, Hwang IC, Lim TW. Heating performance enhancement of a CO2 heat pump system recovering stack exhaust thermal energy in fuel cell vehicles. Int J Refrig 2007;30:1215–26. [16] Kim SC, Kim MS, Hwang IC, Lim TW. Performance evaluation of a CO2 heat pump system for fuel cell vehicles considering the heat exchanger arrangements. Int J Refrig 2007;30:1195–206. [17] Cho CW, Lee HS, Won JP, Lee MY. Measurement and evaluation of heating performance of heat pump system using wasted heat of electric devices for an electric bus. Int J Energies 2012;5:658–69. [18] Lee HS, Won JP, Cho CW, Kim YC, Lee MY. Heating performance characteristics of stack coolant source heat pump using R744 for fuel cell electric vehicles. J Mech Sci Technol 2012;26:2065–71. [19] ASHRAE Standard 51. Laboratory methods of testing fans for rating. Atlanta: American Society of Heating Refrigerating and Air-conditioning Engineers; 1975. [20] ISO/DIS 15042. Multiple split-system air-conditioners and air-to-air heat pumps – testing and rating for performance. International Standards Organization; 2005. [21] ASHRAE Standard 116. Methods of testing for seasonal efficiency of unitary air-conditioner and heat pumps. Atlanta: American Society of Heating Refrigerating and Air-conditioning Engineers; 1993. [22] ASHRAE Guideline 2. Engineering analysis of experimental data. Atlanta: American Society of Heating Refrigerating and Air-conditioning Engineers; 1986. [23] Chang YJ, Wang CC. A generalized heat transfer correlation for louver fin geometry. Int J Heat Mass Transf 1997;40:533–44.
Contents lists available at ScienceDirect
Applied Energy journal homepage: www.elsevier.com/locate/apenergy
Heating performance characteristics of a dual source heat pump using air and waste heat in electric vehicles Jae Hwan Ahn, Hoon Kang, Ho Seong Lee, Hae Won Jung, Changhyun Baek, Yongchan Kim ⇑ Department of Mechanical Engineering, Korea University, Anam-Dong, Sungbuk-Ku, Seoul 136-713, Republic of Korea
h i g h l i g h t s An R134a heat pump with the dual heat source was investigated. Performance characteristics were analyzed by varying heat source condition. Performance of the dual source system was higher than that of the single source. Optimum operation methods were suggested to improve the heating performance.
a r t i c l e
i n f o
Article history: Received 2 September 2013 Received in revised form 18 December 2013 Accepted 23 December 2013
Keywords: Heat pump R134a Automobile Waste heat Performance
a b s t r a c t A heat pump has been considered as an alternative to an electric heater to increase the efficiency of the heating unit in electric vehicles. However, the heating performance of a single source heat pump has yet to be improved at low outdoor temperatures. This study investigates the feasibility of a dual source heat pump using both air and waste heat in electric vehicles. The performance of the dual source heat pump was measured at various operation modes: air source-only, waste heat-only and dual heat source. The heating performance of the dual source heat pump was higher than those of the single source heat pumps. In addition, an alternating single mode operation of air source-only and waste heat-only modes was proposed to improve performance at low outdoor air temperatures. Ó 2014 Elsevier Ltd. All rights reserved.
1. Introduction Electric vehicles (EVs) such as plug-in and fuel-cell vehicles have been developed to meet environmental regulations [1–5]. Electric vehicles are eco-friendly and emit no air-pollutants, but their short driving range is a critical problem. The maximum driving range of recent-developed electric vehicles is approximately 200 km, but it decreases by about 40% with the operation of the heating and cooling units [6]. Therefore, efficient heating and cooling units for the cabin should be developed to increase the driving range of electric vehicles. In this study, heat pumps are considered as alternative heating units to the conventional electric heaters to increase the heating efficiency. The heat pump can be converted to a cooling system with the adoption of 4-way valve by changing the refrigerant flow direction. The conventional cooling system such as a vapor compression cycle has been successfully adopted in electric vehicles with the effort increasing the system efficiency.
⇑ Corresponding author. Tel.: +82 2 3290 3366; fax: +82 2 921 5439. E-mail address: [email protected] (Y. Kim). 0306-2619/$ - see front matter Ó 2014 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.apenergy.2013.12.065
Therefore, it is required to investigate the heating performance of the heat pump using all possible heat sources in electric vehicles. Heat pumps have been widely used as heating devices for commercial and residential buildings because of their high energy efficiency [7–12]. In addition, the application of an air-source heat pump to an internal combustion engine vehicle has been investigated. Antonijevic and Heckt [13] reported that the heat pump was superior to other heating solutions with respect to heating performance and fuel consumption as an automotive heating unit. Hosoz and Direk [14] investigated the operating characteristics of an R134a heat pump using the air-source. The tested heat pump provided sufficient heating performance in mild weather conditions, but its heating capacity dropped rapidly with the decrease of the outdoor temperature, necessitating the use of a supplemental heating device for proper heating operations. The use of waste heat from batteries and electric devices to increase the heating performance of a heat pump has been studied. Kim et al. [15,16] investigated a CO2 heat pump that used partial heat from the stack coolant in fuel cell vehicles. Cho et al. [17] investigated the performance of an R134a heat pump that used waste heat from the electric devices in an electric bus. They
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J.H. Ahn et al. / Applied Energy 119 (2014) 1–9
Nomenclature COP cp CR DB EEV h IDHX _ m OD-EEV ODHX P q_ Q_
coefficient of performance specific heat (kJ kg1 k1) compression ratio dry bulb electronic expansion valve enthalpy (kJ kg1) indoor heat exchanger mass flow rate (kg s1) EEV at the outdoor heat exchanger inlet outdoor heat exchanger pressure (kPa) heat transfer rate (kW) volumetric flow rate (m3 s1) R ratio T temperature (°C) v specific volume (m3 kg1) V velocity (m s1) _ W work (kW) WB wet bulb WH-EEV EEV at the waste heat exchanger inlet
suggested that the heat pump was applicable as a cabin heating device. Lee et al. [18] investigated the performance of a CO2 heat pump that used the stack coolant as the heat source for fuel cell electric vehicles under cold weather conditions. However, the study on the performance of a dual source heat pump that combined air and waste heat for electric vehicles is very limited in the open literature. Previous studies on heat pumps for internal combustion engines and electric vehicles focused on heat pumps that used a single heat source, either air or waste heat. However, all possible heat sources in an electric vehicle must be used to enhance the heating performance of a heat pump at low outdoor temperatures. In this study, an R134a heat pump using dual heat sources of both air and waste heat was investigated to replace the conventional electric heaters used in electric vehicles. The heating performance of the heat pump was measured at various heat source operation modes: air source-only, waste heat-only, and dual heat source. The operating characteristics of the heat pump were analyzed at various heat source conditions, such as air velocity, air temperature, waste heat amount, and coolant flow rate. In addition, system operation methods were suggested to increase the COP and heating capacity of the heat pump.
2. Experimental setup and test procedure 2.1. Experimental setup Fig. 1 shows a schematic diagram of the test setup. The experimental setup was composed of two sections: a vapor compression heating cycle and a coolant loop. The vapor compression heating cycle using R134a was composed of a twin rotary compressor, an outdoor heat exchanger (ODHX), an indoor heat exchanger (IDHX), a waste heat exchanger (WHX), two electric expansion valves (EEVs), and an accumulator. The designed heating capacity of the test setup was 2.5 kW. The hermetic twin rotary compressor had a displacement volume of 24.0 cm3 rev1. The IDHX was a crossflow type, plate-fin brazed-aluminum heat exchanger of 100 mm in depth, 198 mm in height, and 147 mm in width. The ODHX was a parallel-flow type, louvered-fin brazed-aluminum heat
WHO WHX
waste heat-only waste heat exchanger
Subscripts a air c coolant d discharge eva evaporation h heating i inlet ID indoor o outlet OD outdoor ODHX outdoor heat exchanger s suction T total WH waste heat WHX waste heat exchanger
exchanger of 17 mm in depth, 350 mm in height, and 670 mm in width. The maximum capacity and efficiency of the ODHX were 2.3 kW and 0.28, respectively, at the outdoor air temperature of 7 °C and air velocity of 4.5 m s1. The WHX was a plate heat exchanger of 60 mm in depth, 370 mm in height, and 140 mm in width. The maximum capacity and efficiency of the WHX were 2.5 kW and 0.20, respectively, at the waste heat amount of 2.5 kW and coolant flow rate of 0.009 m3 min1. The two EEVs, which were used as expansion devices, had an orifice diameter of 1.6 mm. The coolant loop consisted of a pump, an electric heater, and a water bath, and was designed to simulate the thermal load from the batteries or electric devices in an electric vehicle. The coolant was 50% ethylene–glycol water solution. The thermal load in the coolant loop was adjusted by controlling the electric power to the heater. The heat pump for an electric vehicle has three heat source operation modes: air source-only, waste heat-only, and dual heat source. In the air source-only mode (Fig. 1(a)), the discharged refrigerant from the compressor flows into the IDHX used as a condenser. After heat rejection in the IDHX, the condensed liquid refrigerant enters the ODHX through the EEV at the ODHX inlet (OD-EEV). Consequently, the refrigerant evaporated in the ODHX flows into the compressor. In the waste heat-only mode (Fig. 1(b)), the condensed liquid refrigerant in the IDHX flows into the WHX (coolant-to-refrigerant heat exchanger) through the EEV at the WHX inlet (WH-EEV). The refrigerant evaporated in the WHX using the waste heat enters the compressor and then rejects heat in the IDHX, which converts the waste heat into heating capacity. In the dual heat source mode (Fig. 1(c)), the condensed refrigerant in the IDHX splits into the WHX and ODHX. The heats gained by the refrigerant from the coolant-to-refrigerant heat transfer in the WHX and the air-to-refrigerant heat transfer in the ODHX convert to heating capacity through the IDHX. The heat pump was experimented in the air-enthalpy calorimeter having two chambers: one for the IDHX and the other for the ODHX. The indoor chamber was designed to control the air temperature from 0 °C to 50 °C and relative humidity (RH) from 20% to 90%. The outdoor chamber was designed to control the air temperature from 20 °C and 50 °C and relative humidity (RH) from 20% to 90%. A resistance temperature sensor with an accuracy of
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J.H. Ahn et al. / Applied Energy 119 (2014) 1–9
OD-EEV
M
Nozzle P
P T P
Nozzle
ODHX P T
T H
IDHX
T H
H
Distributor T H
M
Distributor
Pressure sensor Temperature sensor Mass flow meter ` Humidify sensor
T H
P T
P T P T
Fan
P T
Fan Accumulator Compressor
M
WH-EEV P T
P T
P
Nozzle
Heater T
Distributor T H T H
IDHX
M V H
Pressure sensor Temperature sensor Mass flow meter ` Flow meter Humidify sensor
WHX
Pump
P T
P
V
P T P T P T
P T
Fan Accumulator Compressor
WH-EEV M
P T
P T
M
Nozzle
P
P
Nozzle
OD-EEV
P T
T H
WHX Pump P T
T H
T H
P T
P T
P T
Fan
ODHX
`
T H
Distributor
Distributor
IDHX
Heater
Fan
V
P T
P T
Accumulator Compressor
Fig. 1. Schematic diagram of the experimental setup.
±0.2 °C was used to measure the refrigerant temperature, and a digital pressure transducer with an accuracy of ±0.5% was utilized to measure the refrigerant pressure. A mass flow meter with an accuracy of ±0.2% was used to measure the mass flow rate of the refrigerant. A digital power meter with an accuracy of ±0.5% was applied to measure the powers to the compressor and the electric
heater. A magnetic flow meter with an accuracy of ±0.35% was utilized to measure the coolant flow rate. A digital power meter with an accuracy of ±0.2% was used to measure the power to the pump. A digital pressure transducer with an accuracy of ±0.3% was applied to measure the pressure difference across the nozzle. The nozzle method [19] was utilized to measure the air flow rate.
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2.2. Test procedure
COP ¼
The optimum refrigerant charge amount was determined as 1700 g based on preliminary tests, yielding the maximum COP at the standard heating test condition [20] and compressor frequency of 60 Hz. Table 1 shows the test conditions for the heat pump. The indoor air conditions were fixed: dry bulb temperature of 20 °C, wet bulb temperature of 15 °C, and air flow rate of 7.0 m3 min1. In the air source-only mode, the operating characteristics of the heat pump were measured as the outdoor air temperature was varied from 10 °C to 7 °C, and the air velocity from 0.5 to 4.5 m s1. The maximum air velocity was determined as 4.5 m s1 because the effects of the air velocity on the heating capacity became negligible as the air velocity increased beyond 4.5 m s1. In the waste heat-only mode, the operating characteristics of the heat pump were measured as the waste heat amount was varied from 1.0 to 2.5 kW and the coolant flow rate from 0.0015 to 0.012 m3 min1. The waste heat amount is dependent on the power and driving conditions of an electric vehicle. In this study, the range of the waste heat amount was determined to cover wide driving conditions. The waste heat amount of 1.0 kW corresponded to the minimum value at low battery load condition, while the value of 2.5 kW corresponded to the target heating capacity. In the dual heat source mode, the operating characteristics of the heat pump were measured as the outdoor temperature was varied from -10 °C to 7 °C and the waste heat amount from 1.0 to 2.5 kW. The superheats at the outlets of the ODHX and WHX were controlled at 5 °C by modulating the EEV openings. As given in Eq. (1), the heating capacity of the heat pump was evaluated by the air flow rate and air temperature difference across the IDHX [21] where only sensible heat transfer occurred. As given in Eq. (2), the air-side heat transfer rate in the ODHX was evaluated by using the air flow rate and air enthalpy difference across the ODHX where both sensible and latent heat transfer occurred. The air enthalpies were evaluated by measuring the temperatures (DB, WB) and pressure of the humid air at the inlet and outlet of the ODHX. As given in Eq. (3), the heat transfer rate in the WHX was evaluated by using the coolant flow rate and temperature difference across the WHX. The COP of the heat pump was evaluated by Eq. (4). In addition, the uncertainties of the COP and heating capacity were estimated as ±3.7% and ±3.6%, respectively [22]. The estimated bias and precision errors for the COP were 2.35 and 2.87, respectively.
q_ h ¼
Q_ a
v
q_ ev a ¼
q_ w ¼
cp ðT a;o T a;i Þ
Q_ a
ðha;o ha;i Þ
ð2Þ
cp ðT c;o T c;i Þ
ð3Þ
v
Q_ c
vc
ð1Þ
q_ h _ W
ð4Þ
3. Results and discussion 3.1. Performance comparison in single heat source operation modes The operating characteristics of the heat pump were compared between single heat source modes: air source-only and waste heat-only modes. Fig. 2 represents the suction pressure, discharge pressure, and refrigerant flow rate of the heat pump in the air source-only mode. The suction and discharge pressures in the air source-only mode increased with increasing outdoor air velocity because of the increased overall heat transfer coefficient of the ODHX. It should be noted that the increasing rates of these pressures at air velocities below 2.0 m s1 were higher than those at higher air velocities because the increase of the convective heat transfer coefficient according to the air velocity was higher at lower air velocities [23]. In addition, the suction and discharge pressures in the air source-only mode increased with increasing outdoor air temperature due to the increased temperature difference between the refrigerant and air in the ODHX. The refrigerant mass flow rate through the compressor increased with increasing outdoor air velocity and air temperature because of the increased refrigerant density at the compressor inlet from the increased suction pressure. The increase of the refrigerant mass flow rate can provide higher COP and heating capacity in the air source-only mode due to the increased heat transfer rate in the ODHX. In the waste heat-only mode, the coolant was used to recover waste heat. The coolant was circulated by the pump in the closed loop, and this circulation consumed additional pump power. Fig. 3 represents the suction pressure, discharge pressure, and refrigerant flow rate of the heat pump in the waste heat-only mode. The discharge and suction pressures remained constant according to the coolant flow rate at a given waste heat amount because the refrigerant mass flow rate was constant regardless of the coolant flow rate through the WHX. Although the overall heat transfer coefficient in the WHX increased with increasing coolant flow rate, the heat transfer rate from the coolant remained constant because the temperature difference between the coolant and refrigerant decreased at a given waste heat amount. However, the discharge and suction pressures increased with increasing waste heat amount because the refrigerant mass flow rate increased linearly. Therefore, in the waste heat-only mode, the refrigerant mass flow rate and heat transfer rate in the WHX were strongly dependent on the amount of waste heat. Fig. 4 represents the work, heating capacity, and COP of the heat pump in the air source-only mode. The work and heating capacity in the air source-only mode increased with increasing outdoor air velocity and outdoor air temperature because of the increased refrigerant mass flow rate. The heating capacity satisfied the design value of 2.5 kW at the outdoor air temperature of 7 °C and outdoor air velocity of 4.5 m s1. The COP also increased with
Table 1 Test conditions. Parameters
Air source-only mode
Waste heat-only mode
Dual heat source mode
Indoor temperature (°C) Indoor air flow rate (m3 min1) Outdoor temperature (°C) Outdoor air velocity (m s1) Superheat at the ODHX outlet (°C) Waste heat transfer rate (kW) Coolant flow rate (m3 min1) Superheat at the WHX outlet (°C)
20 DB/15 WB 7.0 10, 0, 7 DB 0.5, 1.0, 2.0, 3.0, 4.5 5 – – –
20 DB/15 WB 7.0 7 – – 1.0, 1.5, 2.0, 2.5 0.0015, 0.003, 0.006, 0.009, 0.012 5
20 DB/15 WB 7.0 10, 0, 7 DB 4.5 5 1.0, 1.5, 2.0, 2.5 0.009 5
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350
1400
(a)
TOD=-10oC o
o
TOD= 0 C
TOD= 7oC
TOD= 0 C
TOD= 7oC
1200
(c)
TOD=-10oC
o
TOD= 0 C
300
80
(b)
TOD=-10oC
TOD= 7oC
60
200
-1
m (kg h )
Pd (kPa)
Ps (kPa)
250
1000
40
•
150 800
20
600
0
100 Air source-only mode
50 0
1
2
3
4
5
0
1
2
3
4
5
0
1
2
3
4
5
-1
Vair (ms ) Fig. 2. Variations of (a) suction pressure, (b) discharge pressure, and (c) refrigerant mass flow rate with air velocity and outdoor air temperature in the air source-only mode.
500
1600
(a)
•
qWH= 1.0 kW
•
qWH= 1.5 kW
•
qWH= 2.0 kW
qWH= 1.5 kW
•
1400
qWH= 2.0 kW
1300
qWH= 2.5 kW
•
80
qWH= 2.0 kW
70
qWH= 2.5 kW
•
Pd (kPa)
300
200
100
•
m (kg h-1)
•
qWH= 2.5 kW
Ps (kPa)
qWH= 1.0 kW
90
•
qWH= 1.5 kW
400
(c)
•
qWH= 1.0 kW
1500
•
100
(b)
•
1200 1100
60 50
•
1000
40
900
30
800
20
700
10
Waste heat-only mode
0 0
3
6
9
12
0
3
•
6
9
12
3
3
-1
0
3
6
9
12
Qcoolant x 10 (m min ) Fig. 3. Variations of (a) suction pressure, (b) discharge pressure, and (c) refrigerant mass flow rate with coolant flow rate and waste heat amount in the waste heat-only mode.
1.0 TOD=-10oC
(a)
3.0
TOD=-10oC
o
0.9
4.0
(b)
TOD=-10oC
o
TOD= 0 C
TOD= 0 C
TOD= 7oC
TOD= 7oC
2.5
(c)
o
TOD= 0 C
3.5
TOD= 7oC
0.7
•
0.6
(a)
0.5 0
1
2
2.0
COP
qh (kW)
•
W (kW)
3.0 0.8
1.5
2.5
2.0
1.0
1.5
0.5
1.0
Air source-only mode
3
4
5
0
1
2
3
4
5
0
1
2
3
4
5
Vair (ms -1) Fig. 4. Variations of (a) work, (b) heating capacity, and (c) COP with air velocity and outdoor air temperature in the air source-only mode.
increasing outdoor air velocity and outdoor air temperature because the increasing rate of the heating capacity was higher than that of the work. The COP increased by 19.3% with increasing
outdoor air velocity from 0.5 to 4.5 m s1 at the outdoor temperature of 0 °C. In addition, the COP increased by 49.6% with increasing outdoor air temperature from 10 °C to 0 °C at the outdoor air
J.H. Ahn et al. / Applied Energy 119 (2014) 1–9
velocity of 4.5 m s1. In the air source-only mode, the maximum COP was 3.1 at the outdoor air temperature of 7 °C and outdoor air velocity of 4.5 m s1. Fig. 5 shows the work, heating capacity, and COP of the heat pump in the waste heat-only mode. The work in the waste heatonly mode increased with increasing coolant flow rate at a given waste heat amount due to the increased pumping power, while the heating capacity remained constant. Therefore, the COP decreased slightly with increasing coolant flow rate at a given waste heat amount. In addition, both the work and heating capacity in the waste heat-only mode increased with increasing waste heat amount because of the increased refrigerant mass flow rate. The COP increased with increasing waste heat amount because the increasing rate of the work was lower than that of the heating capacity. At the coolant flow rate of 0.0015 m3 min1, the COP increased by 73% with increasing waste heat amount from 1.0 to 2.5 kW. In the waste heat-only mode, the maximum COP was 3.2 at the waste heat amount of 2.5 kW and coolant flow rate of 0.0015 m3 min1. Fig. 6(a) and (b) represents the heating capacities of the heat pump in the air source-only and waste heat-only modes according to the outdoor air temperature and waste heat amount, respectively. In the air source-only mode, the heating capacity was strongly dependent on the outdoor air temperature. As the outdoor air temperature decreased from 7 °C to 10 °C at the outdoor air velocity of 4.5 m s1, the heating capacity in the air source-only mode decreased from 2.5 to 0.9 kW, which was 36% of the design value. Therefore, in the air source-only mode, the heat pump may not operate properly at low outdoor air temperatures. In the waste heat-only mode, the heating capacity decreased linearly with decreasing waste heat amount. As the waste heat amount decreased from 2.5 to 1.0 kW, the heating capacity in the waste heat-only mode decreased from 2.8 to 1.3 kW, which was 52.4% of the design value. Therefore, the heating capacity should be increased by applying dual heat sources of both air and waste heat at low ambient temperature or low waste heat conditions.
3.5 3.0
•
1.5
1.0
1.0
-8
-4
0
1.5
5
qh (kW)
(b)
•
qWH= 1.5 kW •
qWH= 2.0 kW
4
•
qWH= 2.5 kW
qWH= 2.5 kW
COP
3.0
(c)
•
qWH= 1.0 kW
2.5
3
•
0.8
2.0
0.7
1.5
0.6
1.0
2
Waste heat-only mode
0
3
6
9
2.0
2.5
•
qWH (kW)
than those in the air source-only and waste heat-only modes. The increased suction pressure in the dual heat source mode led to increase the refrigerant mass flow rate. Fig. 8 represents the suction pressure, discharge pressure, and compression ratio of the dual source heat pump according to the waste heat amount and
•
qh (kW)
1.0
8
Fig. 6. Variations of heating capacities (a) with outdoor air temperature in the air source-only mode and (b) with waste heat amount in the waste heat-only mode.
•
0.9
4
TOD (o C)
•
•
W (kW)
0.5
qWH= 2.0 kW
•
Waste heat-only mode
Air source-only mode
0.5 -12
qWH= 1.5 kW
3.5
•
qWH= 2.5 kW
2.0
1.5
qWH= 1.0 kW
qWH= 2.0 kW
1.0
2.5
•
•
qWH= 1.5 kW
3
Fig. 7. Pressure-enthalpy diagrams of the heat pump cycles.
qWH= 1.0 kW
1.1
-1 3.0 Qcoolant= 0.009 m min
Vair= 4.5 m s
2.0
4.0
(a)
•
•
The performance of the heat pump in the dual heat source mode was analyzed as a function of the outdoor air temperature and waste heat amount. Fig. 7 shows the pressure-enthalpy diagram for the heat pump cycle in each operation mode at the outdoor temperature of 0 °C and waste heat amount of 1.5 kW. The suction and discharge pressures in the dual heat source mode were higher
•
(b)
TID= DB 20oC DB
-1
2.5
3.2. Performance improvement in the dual heat source mode
1.2
3.5
(a)
TID= 20oC DB
qh (kW)
6
12
1 0
•
3
6 3
9 3
12
0
3
6
9
12
-1
Qcoolantx 10 (m min ) Fig. 5. Variations of (a) work, (b) heating capacity, and (c) COP with coolant flow rate and waste heat amount in the waste heat-only mode.
7
J.H. Ahn et al. / Applied Energy 119 (2014) 1–9
500
Ps (kPa)
350
TOD= 0oC
1400
TOD=-10 C
TOD=-10oC
TOD= 0oC
TOD= 0oC
TOD= 7 C
TOD= 7oC
WHO
WHO
o
300 250
(c)
TOD= 7oC
8
WHO
1200
CR (-)
400
TOD=-10 C
10
(b)
o
Pd (kPa)
450
1600
(a) o
1000
6
200 4
150
800
100 Dual heat source mode
600 0.0 0.5 1.0 1.5 2.0 2.5
50 0.0 0.5 1.0 1.5 2.0 2.5
2 0.0 0.5 1.0 1.5 2.0 2.5
•
qWH (kW) Fig. 8. Variations of (a) suction pressure, (b) discharge pressure, and (c) compression ratio with waste heat amount and outdoor air temperature in the dual heat source mode.
outdoor air temperature at the outdoor air velocity of 4.5 m s1 and coolant flow rate of 0.009 m3 min1. The suction and discharge pressures in the dual heat source mode increased with increasing waste heat amount at a given outdoor air temperature. At outdoor air temperatures of 0 °C and 7 °C, the suction and discharge pressures in the dual heat source mode were significantly higher than those in the waste heat-only mode because of the use of air source in the ODHX. However, at the outdoor air temperature of 10 °C, the suction and discharge pressures in the dual heat source mode were very close to those in the waste heat-only mode because the heat absorption from the air source was negligible at this lower outdoor air temperature. The evaporating temperature in the dual heat source mode increased from 16.3 °C to 8.9 °C with increasing waste heat amount from 1.0 to 2.5 kW at the outdoor air temperature of 10 °C. Therefore, the heat transfer rate in the ODHX (air source) decreased with increasing waste heat amount because of the decreased temperature difference between the outdoor air and refrigerant. In addition, the compression ratio of the compressor in the dual heat source mode decreased with increasing waste heat amount at a given outdoor air temperature because the increasing rate of the suction pressure was higher than that of the discharge pressure. Fig. 9 shows the refrigerant mass flow rates of the dual source heat pump according to the outdoor air temperature and waste
100 TOD=-10oC
100
(a)
TOD=-10oC
o
100
(b)
TOD= 0 C
TOD= 7oC
80
80
TOD= 7oC
60
-1
40
mT (kg h )
-1
mODHX (kg h )
-1
mWHX (kg h )
•
(c)
TOD= 7oC WHO
WHO
60
TOD=-10oC TOD= 0oC
o
TOD= 0 C
80
heat amount. The refrigerant mass flow rate through the WHX increased significantly with increasing waste heat amount at all outdoor air temperatures. The mass flow rate in the dual heat source mode was higher than that in the waste heat-only mode at outdoor air temperatures of 0 °C and 7 °C due to the increased suction pressure, although it was close to that in the waste heatonly mode at the outdoor air temperature of 10 °C. However, the mass flow rate through the ODHX decreased with the increase in the waste heat amount because the superheat had to be controlled at 5 °C with the increased evaporating temperature in the ODHX. Especially, at the outdoor air temperature of 10 °C, the refrigerant mass flow rate became zero when the waste heat amount increased by over 1.5 kW. The total refrigerant mass flow rate increased with increasing waste heat amount at all outdoor air temperatures due to the simultaneous use of air and water heat sources. At the outdoor air temperature of 10 °C, the total refrigerant mass flow rate with increasing waste heat amount became very close to that in the waste heat-only mode because of the negligible heat absorption from the air source at this low outdoor air temperature. Fig. 10 represents the evaporating capacity in the ODHX and the total evaporating capacity of the dual source heat pump according to the waste heat amount and outdoor air temperature. The trends of the evaporating capacities according to the waste heat amount
60
40
•
40
•
20
20
20 Dual heat source mode
0 0.0 0.5 1.0 1.5 2.0 2.5
0 0.0 0.5 1.0 1.5 2.0 2.5
0 0.0 0.5 1.0 1.5 2.0 2.5
•
qWH (kW) Fig. 9. Variations of (a) refrigerant flow rate through the WHX, (b) refrigerant flow rate through the ODHX, and (c) total refrigerant flow rate with waste heat amount and outdoor air temperature in the dual heat source mode.
8
J.H. Ahn et al. / Applied Energy 119 (2014) 1–9
corresponded to those of the refrigerant mass flow rates (Fig. 8). The evaporating capacity in the ODHX in the dual heat source mode decreased linearly with increasing waste heat amount at all outdoor air temperatures because the temperature difference between the refrigerant and air in the ODHX decreased from the increased evaporating temperature with increasing waste heat amount. At the outdoor air temperature of 10 °C, the evaporating capacity in the ODHX became zero when the waste heat amount increased by over 1.5 kW because the evaporating temperature in the ODHX was higher than the outdoor air temperature. In addition, the total evaporating capacity in the ODHX and in the WHX increased with increasing waste heat amount and outdoor air temperature in the dual heat source mode. The total evaporating capacity in the dual heat source mode at the outdoor air temperature of 0 °C increased by 10.7% with increasing waste heat amount from 1.0 to 2.5 kW. At the outdoor air temperature of 10 °C, the total evaporating capacity in the dual heat source mode became very close to that in the waste heat-only mode with increasing waste heat amount, indicating no heat absorption from the air source. Fig. 11 shows the work, heating capacity, and COP of the dual source heat pump according to the waste heat amount and outdoor air temperature. The work in the dual heat source mode increased 4 4
TOD=-10oC
(a)
TOD=-10oC
(b)
TOD= 0oC
TOD= 0oC
TOD= 7oC
TOD= 7oC
3
qeva,T (kW)
3
qeva,ODHX (kW)
with increasing waste heat amount and outdoor air temperature due to the increased total mass flow rate. The work in the dual heat source mode was higher than those in the air source-only mode (qWH = 0) and waste heat-only mode (WHO) because of the higher total mass flow rate. However, at the outdoor air temperature of 10 °C, the work in the dual heat source mode approached that in the waste heat-only mode with increasing waste heat amount. The COP and heating capacity in the dual heat source mode also increased with increasing waste heat amount and outdoor air temperature due to the increased total mass flow rate. The COP and heating capacity in the dual heat source mode at the outdoor air temperature of 0 °C increased by 9.3% and 31.5%, respectively, with increasing waste heat amount from 0 to 2.5 kW. In addition, the COP and heating capacity in the dual heat source mode at the waste heat amount of 1.0 kW increased by 35.1% and 55.5%, respectively, with increasing outdoor air temperature from 10 °C to 0 °C. It should be noted that the COP and heating capacity in the dual heat source mode were higher than those in the air source-only mode (qWH = 0) and waste heat-only mode (WHO) because of the use of both air and waste heat as heat sources. However, at the outdoor air temperature of 10 °C, the COP and heating capacity in the dual heat source mode approached those in the waste heat-only mode with increasing waste heat amount because the evaporating capacity in the ODHX approached zero. Therefore, at low outdoor air temperatures, the heating performance of the dual source heat pump was strongly dependent on the waste heat amount. An alternating single mode operation of air source-only and waste heat-only modes was proposed to use the air source more effectively at low outdoor air temperatures because the heat absorption from the ODHX in the dual heat source mode was negligible. During the operation in the air source-only mode, the waste heat is stored in the coolant loop without any heat transfer to the heat pump. Consequently, in the waste heat-only mode, the heat pump receives more heat from the coolant loop at an elevated coolant temperature. Fig. 12(a) shows the air temperature at the IDHX outlet and coolant temperature at the WHX inlet according to time. In the air source-only mode, the coolant temperature increased with time, while the air temperature remained constant. Generally, an auxiliary heating such as a PTC heater is required to provide thermal comfort for the passengers at low outdoor temperatures because the heating capacity in the air source-only mode is not sufficient. As the heat pump started to operate in the waste heat-only mode, the air temperature increased due to the higher coolant temperature, even though the coolant temperature decreased with time. The average air temperature at the IDHX out-
2
WHO
2
•
•
1
1
0 0.0
0 0.0
Dual heat source mode
0.5
1.0
1.5
2.0
2.5
0.5
1.0
1.5
2.0
2.5
•
qWH (kW) Fig. 10. Variations of (a) evaporating capacity in the ODHX and (b) total evaporating capacity with waste heat amount and outdoor air temperature in the dual heat source mode.
TOD=-10oC
5
(a)
TOD=-10oC
TOD= 0 C
4
TOD= 7oC
•
0.8
q h (kW)
W (kW)
1.0
TOD= 0 C
TOD= 0 C
TOD= 7oC
3
2
(c)
o
4
WHO
WHO
•
TOD=-10oC
o
o
1.2
5
(b)
TOD= 7oC WHO
COP (-)
1.4
3
2 0.6
1
0.4
0 0.0 0.5 1.0 1.5 2.0 2.5
Dual heat source mode
0.0 0.5 1.0 1.5 2.0 2.5
1 0.0 0.5 1.0 1.5 2.0 2.5
•
qWH (kW) Fig. 11. Variations of (a) work, (b) heating capacity, and (c) COP with waste heat amount and outdoor air temperature in the dual heat source mode.
J.H. Ahn et al. / Applied Energy 119 (2014) 1–9
50
TOD= -10 C • qWH= 1.5 kW
30
T (oC)
130
(a)
(b)
o
20
Ta,o Tc,i
10 0
120
Performance ratio (%)
Air Waste source heat 40 -only -only
Acknowledgements
100
This research was supported by the Korean Evaluation Institute of Industrial Technology (No. 10035530) and the Human Resources Program in Energy Technology of the Korea Institute of Energy Technology Evaluation and Planning (KETEP) grant financial resource from the Ministry of Trade, Industry & Energy, Republic of Korea (No. 20124010203250).
90
80
70 10
20
30
40
50
peratures. At the outdoor air temperature of 10 °C and waste heat amount of 1.5 kW, the COP and heating capacity in the alternating single mode operation were 4.3% and 10.5%, respectively, higher than those in the dual heat source mode.
110
-10 0
9
•
qh
•
W
COP
Time (minute) Fig. 12. (a) Variations of air temperature at the IDHX outlet and coolant temperature at the WHX inlet with time and (b) ratio of heat pump performance in the alternating single mode operation to that in the dual heat source mode.
let was 1.2 °C higher than that in the dual heat source mode at a given waste heat amount. Fig. 12(b) shows the ratio of the heat pump performance in the alternating single mode operation to that in the dual heat source mode at the outdoor air temperature of 10 °C and waste heat amount of 1.5 kW. The heating capacity in the alternating single mode operation was 10.5% higher than that in the dual heat source mode. Even though the operation in the air-source only mode did not provide any benefit in the viewpoint of the total work in the heat pump and heater, the increased heating capacity in the waste heat-only mode led to reduce the total work. In addition, the COP was 4.3% higher than that in the dual heat source mode because the increase in the heating capacity was higher than the increase in the work. 4. Conclusions In this study, an R134a heat pump with dual heat sources of both air and waste heat was investigated experimentally to replace the conventional electric heaters used in electric vehicles. In the air source-only mode, the COP and heating capacity increased with increasing outdoor air velocity and outdoor air temperature. The heat pump in the air source-only mode yielded the design heating capacity at the outdoor temperature of 7 °C, but it provided approximately 36% of the design value at the outdoor air temperature of 10 °C. In the waste heat-only mode, the COP and heating capacity increased with increasing waste heat amount. The heating capacity in the waste heat-only mode satisfied the design value at the waste heat amount of 2.5 kW, but it was 52.4% of the design value at the waste heat amount of 1.0 kW. In the dual heat source mode, the COP and heating capacity increased with increasing waste heat amount and outdoor air temperature because of the increased total mass flow rate. The COP and heating capacity at the outdoor air temperature of 0 °C increased by 9.3% and 31.5%, respectively, with increasing waste heat amount from 0 to 2.5 kW. The heating performance in the dual heat source mode was higher than those in the air source-only and waste heat-only modes because of the use of both air and waste heat as heat sources. However, at the outdoor air temperature of 10 °C, the heating performance in the dual heat source mode was strongly dependent on the waste heat amount because the heat absorption from the air source was negligible. An alternating single mode operation of air source-only and waste heat-only modes was proposed to use the air source more effectively at low outdoor air tem-
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