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A study of the influence of solar radiation on the thermal performance of evaporators of heat pump systems
Applied Thermal Engineering 23 (2003) 1551–1557 www.elsevier.com/locate/apthermeng
A study of the influence of solar radiation on the thermal performance of evaporators of heat pump systems Xiande Fang a
a,*
, Baoyi Chen b, Zhuyi Zheng
c
Institute of Air Conditioning and Refrigeration, Nanjing University of Aeronautics and Astronautics, Nanjing 210016, PR China b Division of Thermal Engineering, Wuhan Institute of Naval Engineering, Wuhan, PR China c Department of Thermal Engineering, Tsinghua University, Beijing, PR China Received 1 July 1998; accepted 6 March 2003
Abstract Solar radiation has remarkable influence on the thermal performance of evaporators of heat pump systems. This paper proposes a method for considering the influence with correcting the heat transfer coefficient of evaporators. A test setup of a heat pump system was built, and long-term experimental data were obtained. The heat transfer coefficient with solar radiation is 36.2% greater than that without it in the experimental conditions. The correction method presented can be used not only to improve analysis and simulation of the thermal performance and energy consumption of heat pump systems, but also to assist in the design and installation of evaporators and the heat pump systems. Ó 2003 Elsevier Science Ltd. All rights reserved. Keywords: Evaporator; Heat pump; Solar radiation; Heat transfer
1. Introduction A heat pump collects energy from outdoor environment through its evaporator. A refrigerant flows through and evaporates in the tube of the evaporator, and absorbs heat from outdoor heat sources, including convective heat from outdoor air, thermal radiation heat from surrounding surfaces, and solar energy from solar radiation.
*
Corresponding author. Address: P.O. Box 297, 4300 Stanley Avenue, Niagara Falls, ON, Canada L2E 6T7. Tel.: +1-905-354-7424; fax: +1-905-354-5706. E-mail address: [email protected] (X.D. Fang). 1359-4311/03/$ - see front matter Ó 2003 Elsevier Science Ltd. All rights reserved. doi:10.1016/S1359-4311(03)00078-4
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X.D. Fang et al. / Applied Thermal Engineering 23 (2003) 1551–1557
Nomenclature Cp F GH2 O GR22 h K K0 Qd Qf Qs qs Rf Rn Rw t Dtm 0 Dtm Greek an aw b gs s
specific heat of water exterior area of the evaporator mass flow rater of water mass flow rate of refrigerant enthalpy of refrigerant overall heat transfer coefficient without the influence of solar radiation overall heat transfer coefficient with the influence of solar radiation transmission heat gain solar gain solar radiation absorbed by the exterior surface of evaporator Qs =F tube-wall thermal resistance of evaporator thermal resistance of tube-wall and refrigerant-side of evaporator air-side thermal resistance of evaporator temperature logarithmic mean temperature difference without influence of solar radiation logarithmic mean temperature difference with influence of solar radiation symbols refrigerant-side convection heat transfer coefficient air-side convection heat transfer coefficient solar gain factor fin efficiency fin effective factor
In the thermal calculation of the evaporator of a heat pump system, the influence of solar radiation on the heat transfer coefficient is usually not considered [2–4,6]. In fact, the thermal performances of evaporators of heat pump systems are affected not only by convection and thermal radiation heat transfer, but also by solar radiation. The experimental study in this paper indicates that the influence of solar radiation on the thermal performance of the evaporator of the heat pump system is considerable, which may cause great inaccuracy in the energy analysis and heating performance simulation if it is neglected. To consider the effect of solar radiation, a correction method for heat transfer coefficients of evaporators of heat pump systems is put forward. A test setup of a heat pump system was built. From the test data, the correction result has been derived.
2. Theoretical analysis of correction methods Let Q stand for the heat absorbed by the refrigerant inside evaporator tubes (see Fig. 1a). It can be calculated by [1,5]
X.D. Fang et al. / Applied Thermal Engineering 23 (2003) 1551–1557 Air flow
1553
Sun Qs Rw
Rn
Qs – Qf
Qf
Refrigerant Refrigerant side
Air side
Q (a)
(b)
Fig. 1. Heat transfer principle and thermal network of the evaporator.
Q ¼ Qd þ Qf
ð1Þ
where transmission heat gain Qd and solar gain Qf can be determined by Eqs. (2) and (3), respectively. Qd ¼ KF Dtm
ð2Þ
Rw Qs Rw þ Rn
ð3Þ
Qf ¼ where
K¼ Rw ¼
1 s þ Rf þ aw gs an
1 aw gs
Rn ¼ Rf þ
1
ðair-side resistanceÞ s an
ðtube-wall and refrigerant-side resistanceÞ
ð4Þ ð5Þ ð6Þ
where the fin efficiency gs and fin effective factor s are defined as gs ¼ s¼
actual heat transferred from the fin attached to the tube wall heat to be transferred if the entire fin is at wall temperature
area of finned surface wall area inside the tube
Defining solar gain factor b as b¼
Rw Rw þ Rn
ð7Þ
it follows that Qf ¼ bQs ¼ bFqs
ð8Þ
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X.D. Fang et al. / Applied Thermal Engineering 23 (2003) 1551–1557
Let 0 Q ¼ Qd þ Qf ¼ K 0 F Dtm
ð9Þ
The solution of Eqs. (2), (8) and (9) gives qs Dtm 0 K ¼ K þb 0 Dtm Dtm
ð10Þ
0 and Dtm can be measured by experiments, K can be determined with Eq. (4) or its where Dtm experimental equation, b can be calculated with Eq. (7) or measured by experiments with the following steps:
1. Measure Qd and Dtm without solar radiation, and obtain K based on Eq. (4). 0 considering solar radiation, and obtain K 0 based on Eq. (9). 2. Measure Q and Dtm 3. Calculate solar gain factor b based on the following equation derived from Eq. (10): 0 Dtm Dtm 0 b¼ K K 0 Dtm qs
ð11Þ
3. Experiments and results 3.1. Test facility A heat pump system for testing the evaporator is illustrated in Fig. 2, where items A to J are the general components of the system. The evaporator is composed of six identical air-refrigerant heat exchangers in parallel connection. Its total outside surface area is 1600 m2 , and the total irradiated area is 70 m2 . The water flow meter consists of a standard orifice flow meter and a double-pipe differential pressure meter. Points 1–5 measure refrigerant pressures and temperatures. Points 6 and 7 measure air temperatures. Point 6 is located 1 m from the suction side. Points 8 and 9 measure water temperaA A
3 8
2
Refrigerant 1
B 9
Air 7
4 C
H D
5
K J
I
Wa ter E F
G
6 Air
Fig. 2. Schematic of the test setup. (A) Compressor, (B) condenser (water-cooled), (C) accumulator, (D) expansion valve, (E) water pump, (F) thermostatic water tank, (G) cooling tower, (H) evaporator, (I) outdoor fan, (J) electric heater, (K) water flow meter. (1) Refrigerant outlet of the evaporator, (2) refrigerant inlet of the compressor, (3) refrigerant inlet of the condenser, (4) refrigerant outlet of the condenser, (5) refrigerant inlet of evaporator, (6) air inlet of the evaporator, (7) air outlet of the evaporator, (8) water outlet of the condenser, (9) water inlet of the condenser.
X.D. Fang et al. / Applied Thermal Engineering 23 (2003) 1551–1557
1555
tures. There are 10 platinum resistance thermometers per square meter distributed uniformly in the air outlet of the evaporator. The average reading of the thermometers is regarded as the outlet air temperature. The refrigerant is R22 whose evaporating process in the evaporator is from Point 5 to Point 1, while super heat process in the pipe is from 1 to 2. 3.2. Experimental methods and results Parameters are measured hourly. A number of data are taken in long-term experiments. A night and a day whose average ambient temperatures are equal to each other were chosen to compare the effect of solar radiation on the evaporator thermal performance. The chosen night was used to simulate a day with the same ambient temperature as the chosen day but without solar radiation. Some of arithmetic mean parameters of the chosen day and the chosen night are shown in Table 1. The velocity of the evaporator inlet air is 2 m/s (6.6 ft/s). 3.3. Analysis According to the experimental results in Table 1 and the thermal properties of refrigerant R22, the enthalpies at points 2–5, h2 h5 , could be calculated. R22 at point 1 (see Fig. 2) is saturated vapor, and its enthalpy h1 can be obtained assuming t1 ¼ t5 . When passing through the evaporator, the air gave off heat, and its temperature felt from t6 to t7 , while refrigerant evaporated and absorbed the heat with temperature remaining constant. The heat lost by the air is equal to the heat absorbed by the refrigerant. That is ½GR22 ðh1 h5 Þ day ¼ Q
ð12aÞ
½GR22 ðh1 h5 Þ night ¼ Qd
ð12bÞ
and
where GR22 is given by the thermal balance of the water-cooled condenser, GR22 ¼
GH2 O Cp ðt8 t9 Þ h3 h4
ð13Þ
Logarithmic mean temperature difference is Dtm ¼
ðt7 t5 Þ ðt6 t1 Þ 5 ln tt76 t t1
ð14Þ
For t1 to t7 , see Fig. 2. Some calculated results are listed in Table 2. Table 1 Measured data of temperature t (°C), pressure p (105 Pa), and water flow rate G (kg/s) Time
t2
t3
t4
t5
t6
t7
t8
t9
p2
p3
p4
p5
G
Day Night
3 1
68 65
42.5 42.1
)8 )10
2 2
)3.3 )3.5
40.1 40.2
36 36
3.5 3.1
16.4 16.1
16.2 16.1
3.8 3.5
24.41 22.33
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X.D. Fang et al. / Applied Thermal Engineering 23 (2003) 1551–1557
Table 2 Some calculated results Time Day Night
Heat absorbed by refrigerant, Q
Log mean temperature, Dtm
Heat transfer coefficient, K
kW
Btu/h
°C
°F
W/(m2 °C)
Btu/(h ft2 °F)
333.5 314.9
113.8 104 107.4 104
7.02 8.97
44.64 48.15
29.78 21.87
5.245 3.852
It can be derived from Table 2 that in the given conditions, Q=Qd ¼ ð333:5=314:9Þ ¼ 1:059 and K 0 =K ¼ ð29:78=21:87Þ ¼ 1:362. That is to say the heat transfer coefficient with solar radiation is 36.2% greater than that without it. Under the experimental conditions, solar radiation intensity Ir ¼ 350 W/m2 . Letting n denote the proportion of the irradiated area of the evaporator to its total exterior area, then n ¼ 0:04375. The solar energy absorptivity of the exterior surface is 0.8, thus qs is equal to 12.25 W/m2 . Calculated with Eq. (11), solar gain factor b is equal to 0.949. 4. Discussion In the given experimental conditions, overall heat transfer coefficient of the evaporator increases by 36%, and the heat output of the heat pump increases by 6%, considering the influence of solar radiation. The higher the solar radiant intensity becomes, the more they increase. With solar gain factor b and Eqs. (9) and (10), the thermal performance of the evaporator of a heat pump system can be corrected. Because Rn is far smaller than Rw , its variation can be neglected in calculating b, and b can be considered as constant in a given evaporator. b can be derived from experimental data as discussed before, and should be able to be determined with calculating Rn and Rw . The later may be inaccurate since remarkable error may exist in the calculation of heat transfer coefficients. In the design of evaporators, solar radiation should be considered in order to predict the potential capacity of the system. Increasing the irradiated area proportion and the exterior-surface absorptivity of evaporators can receive more benefit from solar radiant energy. This is true both in summer and in winter. The fan at the air inlet of the evaporator might be designed as air flow rate adjustable so that it can decrease air flow to reduce energy consumption when ambient air temperature and/or solar radiant intensity is rather high. Moreover, a heat pump system can be equipped with an appropriate auxiliary heating device, which is used only during cloudy days (without solar radiation). A heat pump system can be designed with the rule of meeting the need during sunny days in order to reduce the initial investment and operation cost. 5. Conclusions The influence of solar radiation on the thermal performance of evaporators of heat pump systems is considerable. The heat transfer coefficient with solar radiation is 36.2% larger than that without it under the experimental conditions.
X.D. Fang et al. / Applied Thermal Engineering 23 (2003) 1551–1557
1557
This paper provides a method for considering the influence of solar radiation on evaporators of heat pump systems by using solar gain factor b. An evaporator is studied and some available results are obtained. b can be considered as constant in a given heat pump system, and derived from experimental data. Besides, b should be able to be determined by Eq. (7) with calculating Rn and Rw though some inaccuracy may yield because of the remarkable error in the calculation of heat transfer coefficients. The correction method may help not only to improve the accuracy of the heating performance simulation and the energy analysis of heat pump systems, but also to assist in the system design and installations.
References [1] ASHRAE Handbook of Fundamentals, 1997 (Chapter 29). [2] M.J. Chimack, Development of a heat pump simulation model for the BLAST energy analysis computer program. Master Thesis, University of Illinois at Urbana-Champaign, 1987, pp. 50–51. [3] Q.T. Guo, J.H. Wu, Design Handbook of Refrigeration, Press of the Architecture Industries of China, Beijing, 1994, p. 25. [4] I. Haider, R. Radermacher, Simulation results for an air-conditioning system with secondary refrigerant loops, Heat Pump and Refrigeration Systems Design, Analysis and Applications AES-34 (1995) 59–64. [5] R.Z. Shou, X.D. Fang, H.S. He, Air conditioning, Press of Beijing University of Aeronautics and Astronautics, Beijing, 1992, pp. 47–48. [6] Q.S. Yan, Refrigeration for air conditioning, Press of the Architecture Industries of China, Beijing, 1985, p. 73, 95.
A study of the influence of solar radiation on the thermal performance of evaporators of heat pump systems Xiande Fang a
a,*
, Baoyi Chen b, Zhuyi Zheng
c
Institute of Air Conditioning and Refrigeration, Nanjing University of Aeronautics and Astronautics, Nanjing 210016, PR China b Division of Thermal Engineering, Wuhan Institute of Naval Engineering, Wuhan, PR China c Department of Thermal Engineering, Tsinghua University, Beijing, PR China Received 1 July 1998; accepted 6 March 2003
Abstract Solar radiation has remarkable influence on the thermal performance of evaporators of heat pump systems. This paper proposes a method for considering the influence with correcting the heat transfer coefficient of evaporators. A test setup of a heat pump system was built, and long-term experimental data were obtained. The heat transfer coefficient with solar radiation is 36.2% greater than that without it in the experimental conditions. The correction method presented can be used not only to improve analysis and simulation of the thermal performance and energy consumption of heat pump systems, but also to assist in the design and installation of evaporators and the heat pump systems. Ó 2003 Elsevier Science Ltd. All rights reserved. Keywords: Evaporator; Heat pump; Solar radiation; Heat transfer
1. Introduction A heat pump collects energy from outdoor environment through its evaporator. A refrigerant flows through and evaporates in the tube of the evaporator, and absorbs heat from outdoor heat sources, including convective heat from outdoor air, thermal radiation heat from surrounding surfaces, and solar energy from solar radiation.
*
Corresponding author. Address: P.O. Box 297, 4300 Stanley Avenue, Niagara Falls, ON, Canada L2E 6T7. Tel.: +1-905-354-7424; fax: +1-905-354-5706. E-mail address: [email protected] (X.D. Fang). 1359-4311/03/$ - see front matter Ó 2003 Elsevier Science Ltd. All rights reserved. doi:10.1016/S1359-4311(03)00078-4
1552
X.D. Fang et al. / Applied Thermal Engineering 23 (2003) 1551–1557
Nomenclature Cp F GH2 O GR22 h K K0 Qd Qf Qs qs Rf Rn Rw t Dtm 0 Dtm Greek an aw b gs s
specific heat of water exterior area of the evaporator mass flow rater of water mass flow rate of refrigerant enthalpy of refrigerant overall heat transfer coefficient without the influence of solar radiation overall heat transfer coefficient with the influence of solar radiation transmission heat gain solar gain solar radiation absorbed by the exterior surface of evaporator Qs =F tube-wall thermal resistance of evaporator thermal resistance of tube-wall and refrigerant-side of evaporator air-side thermal resistance of evaporator temperature logarithmic mean temperature difference without influence of solar radiation logarithmic mean temperature difference with influence of solar radiation symbols refrigerant-side convection heat transfer coefficient air-side convection heat transfer coefficient solar gain factor fin efficiency fin effective factor
In the thermal calculation of the evaporator of a heat pump system, the influence of solar radiation on the heat transfer coefficient is usually not considered [2–4,6]. In fact, the thermal performances of evaporators of heat pump systems are affected not only by convection and thermal radiation heat transfer, but also by solar radiation. The experimental study in this paper indicates that the influence of solar radiation on the thermal performance of the evaporator of the heat pump system is considerable, which may cause great inaccuracy in the energy analysis and heating performance simulation if it is neglected. To consider the effect of solar radiation, a correction method for heat transfer coefficients of evaporators of heat pump systems is put forward. A test setup of a heat pump system was built. From the test data, the correction result has been derived.
2. Theoretical analysis of correction methods Let Q stand for the heat absorbed by the refrigerant inside evaporator tubes (see Fig. 1a). It can be calculated by [1,5]
X.D. Fang et al. / Applied Thermal Engineering 23 (2003) 1551–1557 Air flow
1553
Sun Qs Rw
Rn
Qs – Qf
Qf
Refrigerant Refrigerant side
Air side
Q (a)
(b)
Fig. 1. Heat transfer principle and thermal network of the evaporator.
Q ¼ Qd þ Qf
ð1Þ
where transmission heat gain Qd and solar gain Qf can be determined by Eqs. (2) and (3), respectively. Qd ¼ KF Dtm
ð2Þ
Rw Qs Rw þ Rn
ð3Þ
Qf ¼ where
K¼ Rw ¼
1 s þ Rf þ aw gs an
1 aw gs
Rn ¼ Rf þ
1
ðair-side resistanceÞ s an
ðtube-wall and refrigerant-side resistanceÞ
ð4Þ ð5Þ ð6Þ
where the fin efficiency gs and fin effective factor s are defined as gs ¼ s¼
actual heat transferred from the fin attached to the tube wall heat to be transferred if the entire fin is at wall temperature
area of finned surface wall area inside the tube
Defining solar gain factor b as b¼
Rw Rw þ Rn
ð7Þ
it follows that Qf ¼ bQs ¼ bFqs
ð8Þ
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X.D. Fang et al. / Applied Thermal Engineering 23 (2003) 1551–1557
Let 0 Q ¼ Qd þ Qf ¼ K 0 F Dtm
ð9Þ
The solution of Eqs. (2), (8) and (9) gives qs Dtm 0 K ¼ K þb 0 Dtm Dtm
ð10Þ
0 and Dtm can be measured by experiments, K can be determined with Eq. (4) or its where Dtm experimental equation, b can be calculated with Eq. (7) or measured by experiments with the following steps:
1. Measure Qd and Dtm without solar radiation, and obtain K based on Eq. (4). 0 considering solar radiation, and obtain K 0 based on Eq. (9). 2. Measure Q and Dtm 3. Calculate solar gain factor b based on the following equation derived from Eq. (10): 0 Dtm Dtm 0 b¼ K K 0 Dtm qs
ð11Þ
3. Experiments and results 3.1. Test facility A heat pump system for testing the evaporator is illustrated in Fig. 2, where items A to J are the general components of the system. The evaporator is composed of six identical air-refrigerant heat exchangers in parallel connection. Its total outside surface area is 1600 m2 , and the total irradiated area is 70 m2 . The water flow meter consists of a standard orifice flow meter and a double-pipe differential pressure meter. Points 1–5 measure refrigerant pressures and temperatures. Points 6 and 7 measure air temperatures. Point 6 is located 1 m from the suction side. Points 8 and 9 measure water temperaA A
3 8
2
Refrigerant 1
B 9
Air 7
4 C
H D
5
K J
I
Wa ter E F
G
6 Air
Fig. 2. Schematic of the test setup. (A) Compressor, (B) condenser (water-cooled), (C) accumulator, (D) expansion valve, (E) water pump, (F) thermostatic water tank, (G) cooling tower, (H) evaporator, (I) outdoor fan, (J) electric heater, (K) water flow meter. (1) Refrigerant outlet of the evaporator, (2) refrigerant inlet of the compressor, (3) refrigerant inlet of the condenser, (4) refrigerant outlet of the condenser, (5) refrigerant inlet of evaporator, (6) air inlet of the evaporator, (7) air outlet of the evaporator, (8) water outlet of the condenser, (9) water inlet of the condenser.
X.D. Fang et al. / Applied Thermal Engineering 23 (2003) 1551–1557
1555
tures. There are 10 platinum resistance thermometers per square meter distributed uniformly in the air outlet of the evaporator. The average reading of the thermometers is regarded as the outlet air temperature. The refrigerant is R22 whose evaporating process in the evaporator is from Point 5 to Point 1, while super heat process in the pipe is from 1 to 2. 3.2. Experimental methods and results Parameters are measured hourly. A number of data are taken in long-term experiments. A night and a day whose average ambient temperatures are equal to each other were chosen to compare the effect of solar radiation on the evaporator thermal performance. The chosen night was used to simulate a day with the same ambient temperature as the chosen day but without solar radiation. Some of arithmetic mean parameters of the chosen day and the chosen night are shown in Table 1. The velocity of the evaporator inlet air is 2 m/s (6.6 ft/s). 3.3. Analysis According to the experimental results in Table 1 and the thermal properties of refrigerant R22, the enthalpies at points 2–5, h2 h5 , could be calculated. R22 at point 1 (see Fig. 2) is saturated vapor, and its enthalpy h1 can be obtained assuming t1 ¼ t5 . When passing through the evaporator, the air gave off heat, and its temperature felt from t6 to t7 , while refrigerant evaporated and absorbed the heat with temperature remaining constant. The heat lost by the air is equal to the heat absorbed by the refrigerant. That is ½GR22 ðh1 h5 Þ day ¼ Q
ð12aÞ
½GR22 ðh1 h5 Þ night ¼ Qd
ð12bÞ
and
where GR22 is given by the thermal balance of the water-cooled condenser, GR22 ¼
GH2 O Cp ðt8 t9 Þ h3 h4
ð13Þ
Logarithmic mean temperature difference is Dtm ¼
ðt7 t5 Þ ðt6 t1 Þ 5 ln tt76 t t1
ð14Þ
For t1 to t7 , see Fig. 2. Some calculated results are listed in Table 2. Table 1 Measured data of temperature t (°C), pressure p (105 Pa), and water flow rate G (kg/s) Time
t2
t3
t4
t5
t6
t7
t8
t9
p2
p3
p4
p5
G
Day Night
3 1
68 65
42.5 42.1
)8 )10
2 2
)3.3 )3.5
40.1 40.2
36 36
3.5 3.1
16.4 16.1
16.2 16.1
3.8 3.5
24.41 22.33
1556
X.D. Fang et al. / Applied Thermal Engineering 23 (2003) 1551–1557
Table 2 Some calculated results Time Day Night
Heat absorbed by refrigerant, Q
Log mean temperature, Dtm
Heat transfer coefficient, K
kW
Btu/h
°C
°F
W/(m2 °C)
Btu/(h ft2 °F)
333.5 314.9
113.8 104 107.4 104
7.02 8.97
44.64 48.15
29.78 21.87
5.245 3.852
It can be derived from Table 2 that in the given conditions, Q=Qd ¼ ð333:5=314:9Þ ¼ 1:059 and K 0 =K ¼ ð29:78=21:87Þ ¼ 1:362. That is to say the heat transfer coefficient with solar radiation is 36.2% greater than that without it. Under the experimental conditions, solar radiation intensity Ir ¼ 350 W/m2 . Letting n denote the proportion of the irradiated area of the evaporator to its total exterior area, then n ¼ 0:04375. The solar energy absorptivity of the exterior surface is 0.8, thus qs is equal to 12.25 W/m2 . Calculated with Eq. (11), solar gain factor b is equal to 0.949. 4. Discussion In the given experimental conditions, overall heat transfer coefficient of the evaporator increases by 36%, and the heat output of the heat pump increases by 6%, considering the influence of solar radiation. The higher the solar radiant intensity becomes, the more they increase. With solar gain factor b and Eqs. (9) and (10), the thermal performance of the evaporator of a heat pump system can be corrected. Because Rn is far smaller than Rw , its variation can be neglected in calculating b, and b can be considered as constant in a given evaporator. b can be derived from experimental data as discussed before, and should be able to be determined with calculating Rn and Rw . The later may be inaccurate since remarkable error may exist in the calculation of heat transfer coefficients. In the design of evaporators, solar radiation should be considered in order to predict the potential capacity of the system. Increasing the irradiated area proportion and the exterior-surface absorptivity of evaporators can receive more benefit from solar radiant energy. This is true both in summer and in winter. The fan at the air inlet of the evaporator might be designed as air flow rate adjustable so that it can decrease air flow to reduce energy consumption when ambient air temperature and/or solar radiant intensity is rather high. Moreover, a heat pump system can be equipped with an appropriate auxiliary heating device, which is used only during cloudy days (without solar radiation). A heat pump system can be designed with the rule of meeting the need during sunny days in order to reduce the initial investment and operation cost. 5. Conclusions The influence of solar radiation on the thermal performance of evaporators of heat pump systems is considerable. The heat transfer coefficient with solar radiation is 36.2% larger than that without it under the experimental conditions.
X.D. Fang et al. / Applied Thermal Engineering 23 (2003) 1551–1557
1557
This paper provides a method for considering the influence of solar radiation on evaporators of heat pump systems by using solar gain factor b. An evaporator is studied and some available results are obtained. b can be considered as constant in a given heat pump system, and derived from experimental data. Besides, b should be able to be determined by Eq. (7) with calculating Rn and Rw though some inaccuracy may yield because of the remarkable error in the calculation of heat transfer coefficients. The correction method may help not only to improve the accuracy of the heating performance simulation and the energy analysis of heat pump systems, but also to assist in the system design and installations.
References [1] ASHRAE Handbook of Fundamentals, 1997 (Chapter 29). [2] M.J. Chimack, Development of a heat pump simulation model for the BLAST energy analysis computer program. Master Thesis, University of Illinois at Urbana-Champaign, 1987, pp. 50–51. [3] Q.T. Guo, J.H. Wu, Design Handbook of Refrigeration, Press of the Architecture Industries of China, Beijing, 1994, p. 25. [4] I. Haider, R. Radermacher, Simulation results for an air-conditioning system with secondary refrigerant loops, Heat Pump and Refrigeration Systems Design, Analysis and Applications AES-34 (1995) 59–64. [5] R.Z. Shou, X.D. Fang, H.S. He, Air conditioning, Press of Beijing University of Aeronautics and Astronautics, Beijing, 1992, pp. 47–48. [6] Q.S. Yan, Refrigeration for air conditioning, Press of the Architecture Industries of China, Beijing, 1985, p. 73, 95.