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A solar-air hybrid source heat pump for space heating and domestic hot water
Solar Energy 199 (2020) 347–359
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A solar-air hybrid source heat pump for space heating and domestic hot water
T
Siyuan Rana,b, Xianting Lia, , Wei Xub, Baolong Wanga ⁎
a b
Department of Building Science, School of Architecture, Tsinghua University, Beijing 100084, People’s Republic of China China Academy of Building Research, Beijing 100013, People’s Republic of China
ARTICLE INFO
ABSTRACT
Keywords: Solar energy Air source heat pump Defrost Hybrid source heat pump Space heating
Traditional solar collectors (SCs) cannot effectively use solar energy of varying intensity, and air source heat pumps cannot supply heat steadily during defrosting conditions. Accordingly, a solar-air hybrid source heat pump (HSHP) system is proposed to solve these problems. A SC was indirectly connected to a heat pump evaporator and water tank for a user at the same time, to allow direct heating when the solar energy is abundant and to supply heat to the heat pump. Multiple air heat exchangers were also connected indirectly to the heat pump. When one of these is in defrosting mode, the others can still work to guarantee a continuous heat output. The structure and operation strategies are introduced, and a numerical model is established to evaluate the performance of the solar-air HSHP. The performance was also compared with a traditional SC plus an air source heat pump (SC + ASHP). The results show that: (1) the heat provided by the solar-air source heat pump mode is 15%, 11%, and 12% of the total heat production in Chengdu, Beijing, and Shenyang, respectively, which indicates that weak solar radiation can also be used effectively; (2) the energy for defrosting using a solar-air HSHP is only between 30% and 36% of that when using a SC + ASHP system, and between 16% and 35% of that in the ASHP system, which indicates a large energy savings potential of the system in defrosting conditions; and (3) the seasonal performance factor (SPF) during a heating season is 3.61, 3.27, and 2.45 in Chengdu, Beijing, and Shenyang, respectively, and it can reach up to 6.92 in Lhasa due to sufficient solar energy resources.
1. Introduction The energy use for space heating takes up 23% of total energy use for buildings in northern China, and hot water uses almost a quarter of the energy in residential buildings (Jiang, 2011). Moreover, energy use is growing rapidly due to the increased intensity of energy use and increased building floor space. In contrast, China is limiting total primary energy use to less than 4.8 billion tons of standard coal equivalent (tce), and the ceiling for building energy use is 1.1 billion tce (Peng et al., 2015). Thus, renewable and clean energy should be introduced to resolve the conflict between energy use demands and limitations. Solar thermal energy is a potential solution to these problems, and is widely used across the world (Tian et al., 2019). However, there are some disadvantages with conventional solar thermal systems. First, solar collectors (SCs) can only supply hot water at a high capacity when solar radiation is sufficient, meaning they cannot supply hot water during the night. Therefore, an auxiliary heat source (driven by electricity or gas) is typically attached to the system. Second, the temperature of the solar absorber must be higher than the temperature of
⁎
the hot water provided to the users, which is in turn higher than that of ambient air, and heat loss occurs by convection from the SC to the ambient air. If solar radiation is weak (e.g. 0–300 W/m2), the heat from solar is mostly (or even completely) lost; therefore, weak solar radiation cannot be used efficiently in traditional SCs. To overcome these shortcomings, a solar assisted heat pump system has been suggested by many researches that combines the SC and an air source heat pump (ASHP). Sterling and Collins (2012) found that electrical use and operating costs of solar assisted heat pump were the lowest with the proposed system when compared with traditional electric and solar hot water systems. Panaras et al. (2014) introduced a method to predict the thermal energy provided to the user and the electrical energy consumed by the heat pump, which is based on dynamic system testing results. Thermal storage, an important part in solar heat pump systems, was calculated and optimized by Badescu (2003). The results showed that smaller storage units provide greater heat flux to the heat pump, while the larger units provide heat for longer time periods. And phase change material can also be combined with solar powered heat pump systems
Corresponding author. E-mail address: [email protected] (X. Li).
https://doi.org/10.1016/j.solener.2020.02.038 Received 27 November 2019; Received in revised form 15 January 2020; Accepted 9 February 2020 0038-092X/ © 2020 International Solar Energy Society. Published by Elsevier Ltd. All rights reserved.
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density, kg/m3
Nomenclature
A cp, a Dh h H Isolar k L Le Nu P Pr q Q R Re T w y
Abbreviations
area, m2 specific heat capacity, J/(kg·K) hydraulic diameter of the tube, mm heat transfer coefficient, W/(m2·K) mass transfer coefficient, kg/(m2·s) total solar irradiation on tilted surface, W/m2 thermal conductivity, W/(m·K) length of the cooling plate, mm Lewis number Nusselt number electric power, W Prandtl number latent heat of sublimation, J/kg heat transfer capacity, W radius of the tube of the AHX, mm Reynolds number temperature, °C absolute humidity, kg/kg dry air thickness, mm
AHX air heat exchanger ASHP air source heat pump COP coefficient of performance Mode AHP air source heat pump mode Mode SAHP solar air source heat pump mode Mode SD solar direct heating mode HSHP hybrid source heat pump SC solar collector SPF seasonal performance factor Subscripts a coll f fin i sen sys tube
Greek letters efficiency (Esen, 2000). There are also solar heat pump systems coupled with geothermal pipes. Chen and Yang (2012) optimized the design of a solar assisted ground coupled heat pump, showing that the annual total heat extraction (plus 75% of the hot water requirement) can be provided by solar energy. Simulation models were built for solar-assisted ground source heat pump systems, and the performances of the systems with different types of ground heat exchangers are compared (Esen et al., 2017). In addition to the simulation mentioned above, there are also laboratory tests for solar heat pump systems. Dikici and Akbulut (2008) tested a solar assisted heat pump with flat plate collectors for space heating for a 60 m2 room in Turkey. The coefficient of performance (COP) reached 3.08 while the exergy loss of the SC was 1.92 kW. A twophase thermosyphon solar collector using refrigerants R-134a, R407C, and R410A was tested and compared (Esen and Esen, 2005). Another experimental set-up was constructed with a liquid-to-liquid vapor compression heat pump and twelve flat-plate SCs (Bakirci and Yuksel, 2011). Caglar and Yamali (2012) designed a solar heat pump system and evaluated its performance both theoretically and experimentally. The evaporating temperature varied between 5.2 and 20.7 °C, while storage tank temperature varied between 9 and 35 °C and the COP was 6.38. When the solar heat pump system was integrated into a passive house, it consumed electricity as its convectional energy source, which only accounted for about one-third of the total energy supplied for heating (Shan et al., 2016). Esen and Yuksel (2013) experimentally investigated greenhouse heating system including solar, biogas and ground energy. In the above research, solar assisted heat pumps were mainly divided into two categories according to the connection of the SC and heat pump: series connected and parallel connected. In series connected solar heat pumps, the SC is the heat source of the heat pump unit. This type of system cannot make full use of solar energy because it still needs to consume electricity, even when solar energy is abundant. Moreover, it cannot supply hot water during the night; therefore, it is mainly used for domestic hot water, which means having a water tank is essential. In parallel connected systems, the SCs supply hot water directly when solar energy is sufficient, but cannot take advantage of weak solar
ambient air solar collector frost/frost surface fin of the heat exchanger inlet sensible convection heat transfer system tube of the heat exchanger
energy. Based on the above systems, some researchers proposed a solar air heat pump system that is able to use solar energy effectively. In effect, this means directly heating when solar energy is sufficient, and providing heat to the heat pump evaporator unit when solar intensity is weak. Liu et al. developed a solar air composite heat source heat pump system with a three medium composite heat exchanger as the evaporator of the system. The heat exchanger was made by inserting another heat exchanger tube into the tube of a fin tube heat exchanger, which allows the system to use either (or both) the air source and solar heat as the heat source (Liu et al., 2012). The system contains three operation modes: single air source, solar-air dual-source, and single solar. Further, the optimal control strategy in different working conditions was studied (Zhang et al., 2011). To investigate system performance further, an experiment table was established, and the results showed that when the ambient temperature was −15 °C, compared to the single air heat source mode, the dual heat source mode heat capacity increased by 62% and the COP increased by 59% (Liu et al., 2016). When the ambient temperature is −7 °C, the heating capacity and COP of the solar air source mode increase by approximately 24% and 25%, respectively (Zhou et al., 2010). Furthermore, a new three medium composite heat exchanger has been designed to make this process easier and to make batch production more convenient (Zhang et al., 2012). Although this type of system can efficiently utilize solar energy and heat from ambient air, the structure of the evaporator is too complex for commercial viability. Moreover, the control strategy did not include the direct use of solar energy. Ni et al. (2016b) proposed a solar assisted air source heat pump with phase change material (PCM), consisting of an air source heat pump, a triple-sleeve energy storage exchanger, and a solar thermal collector. The system enhances performance at low ambient temperatures and eliminates the discrepancy between the demand and supply sides. To test the system performance under various conditions, a prototype was set-up and implemented, and the results showed that the heating COP of the proposed system increased by 65% compared to the ASHP system when the outdoor air temperature was below 10 °C (Qv et al., 2015). The performance in severe external conditions was tested further (Ni et al., 2016a), and the results indicated that there was a great 348
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improvement in both the system efficiency and heating and cooling capacity in the severe external conditions. As the triple-sleeve energy storage exchanger is the core component of the system, a mathematical model was established to analyze the impact parameters (Niu et al., 2011). In addition, the integrated system cool storage performance was studied (Niu et al., 2012). However, there were disadvantages similar to the previous system: (1) the structure of the triple-sleeve energy storage exchanger is complex; and (2) the operation modes did not include a direct solar heating mode, which decreases the efficiency when solar energy is sufficient. As air source heat exchanger is essential in the solar assisted heat pumps introduced above, and there exists a frosting problem when it works in low temperature and high humidity condition. Although the defrosting methods have been proposed for many years and are widely used nowadays, there are still some drawbacks. There are two kinds of defrosting method for ASHP: reverse-cycle defrosting and hot-gas bypass defrosting (Huang et al., 2009). When the ASHP is in reverse-cycle defrosting mode, it extracts heat from indoor air (or circulating hot water), which has a severer impact on indoor thermal comfort. As for the hot-gas bypass defrosting method, the ASHP can provide little heat in defrosting mode. Moreover, the heat used for melting the ice is from electricity, which is not energy efficient. In the above researches on solar assisted heat pumps, there is no newly proposed defrosting method. Therefore, there is a need for a new defrosting method to improve thermal comfort and reduce energy use. In this study, a solar-air hybrid source heat pump (HSHP) system is proposed to efficiently use solar energy, and a novel defrosting method is applied to the system, allowing it to provide steady heat during defrosting conditions and decreasing the energy used in defrosting.
When there is no solar radiation, all the valves that related to the SC are closed, and the AHXs valves are opened (Fig. 2c). The AHXs are used to extract heat from the air and provide the low-grade heat to the heat pump unit to produce hot water for users. (4) Defrosting mode In high humidity and low temperatures, the defrosting mode is essential, and the AHXs run the defrosting process individually. Only one AHX is in operation during the defrosting process at one time, and the other AHXs still operate in the heating mode and extract heat from the ambient air. Via the water-water heat exchanger, part of the heat from the water tank is used for the AHX in the defrosting process. After the defrosting process of one AHX is finished, another AHX begins the defrosting process, and the other AHXs continue to operate normally. According to the structure of the solar-air HSHP introduced above, there are two points of originality: (1) The AHX and SC are both indirectly connected to the evaporator of the heat pump, in the meanwhile the SC is connected to the hot water tank via a water-water heat exchanger; (2) A novel defrosting method is proposed suitable for the system. Thus two benefits are brought by the originalities: stable heating supply during defrosting condition, and efficient utilization of weak solar energy, which are explained as follows: Since only one AHX is undergoing the defrosting process and the other AHXs are operating normally, the heating capacity can be maintained for users at to nearly the same extent as that without frost. Thus, the comfort and reliability of heating can be guaranteed. Moreover, less energy is consumed by the defrosting process because the hot water produced by the solar-air HSHP itself is used for defrosting. Therefore, the drawbacks of using a traditional ASHP in winter for heating can be significantly mitigated. The solar-air HSHP system can perform all the following functions: (1) extracting heat from the ambient air efficiently and without solar radiation, (2) utilizing both solar energy and the ambient air when there is weak solar radiation, and (3) direct solar heating when there is sufficient solar radiation. Therefore, compared with the traditional solar heating system, the proposed solar-air HSHP system can fully use solar radiation with different intensities. Compared with the traditional ASHP system, the inlet water temperature of the heat pump evaporator increases due to the utilization of weak solar radiation, which is beneficial to the performance of the heat pump unit. Compared with the solar assisted air source heat pump in previous research, the components of the solar-air HSHP (i.e. SC, AHX, heat pump unit and water tank) are widely used products, which significantly reduces the cost and will contribute to the popularity. In general, this type of solar-air HSHP system has great application potential in areas with abundant solar energy resources.
2. Concept of the solar-air HSHP A solar-air hybrid source heat pump (HSHP) system was developed to use solar energy efficiently and stably heat a building in defrosting conditions. As indicated in Fig. 1, the system mainly consists of several air heat exchangers (AHXs), a solar collector (SC), a water-water heat exchanger, and a heat pump unit. The evaporator of the heat pump unit and the SC are connected by secondary fluid. The SC is connected with both the evaporator of the heat pump unit and the water tank through the water-water heat exchanger, which enables the SC to supply heat either to the heat pump or directly to the user. The AHXs are connected to the evaporator and act as the heat source of the heat pump. Meanwhile, heat can be transferred to the AHXs via the water-water heat exchanger for defrosting purposes. There are four operation modes of the solar-air HSHP, based on the weather conditions. (1) Solar direct heating mode (Mode SD) When solar energy is sufficient and the temperature of the SC is higher than that of the water tank, the plate heat exchanger transfers the heat harvested by the SC directly to users via the water-water heat exchanger (Fig. 2a). The greatest advantage of the Mode SD is that it consumes no electricity because solar energy is used directly. (2) Solar air source heat pump mode (Mode SAHP) When solar radiation is weak, the valves for Mode SD are closed, and the valves of the AHXs and SC (used to extract heat from air) are opened (Fig. 2b). Thus, low-grade heat can be extracted from both the air and solar energy using the AHXs and SC, respectively. Then, highgrade heat is supplied to users after the heat pump unit process is completed. (3) Air source heat pump mode (Mode AHP)
Fig. 1. Schematic diagram of the solar-air HSHP system. 349
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(a) Mode SD.
(b) Mode SAHP.
(c) Mode AHP.
(d) Defrosting mode.
Fig. 2. Operation modes of the solar-air HSHP system.
Fig. 3. Diagram of the TRNSYS model. 350
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3. Modeling the system
water tank is 80 °C. If T_load is lower than 80 °C and T_solar is higher than T_load, then the temperature of the SC is high enough for direct heating, and the system is in Mode SD. In another case, the model examines the on/off functioning of the heat pump unit for the previous time step and then compares the T_load with minimum and maximum thresholds of 40 and 45 °C, respectively. The aim of these two steps is to make sure the water tank temperature is between 40 and 45 °C, and to decide whether the heat pump unit should be turned on. If the heat pump should turn on and there is solar radiation, the system is in Mode SAHP; otherwise, it is in Mode AHP. The output of this process determines which mode the system should use.
A mathematical model was built to evaluate the system performance under different conditions, including the defrosting condition. Besides, the models for the benchmark systems are introduced. The system electricity use consists of heat pump and circulating water pump for SC and AHX, and excludes the electricity use of fan coil unit and the circulating water pump for it. Both the models of benchmark system and the solar-air HSHP follow this rule for fair comparison. 3.1. TRNSYS model and control strategy
3.2. Modeling of frosting and defrosting process of AHX
The first step is to build a model that can calculate the performance without regard to the defrosting condition. The performance of the solar-air HSHP is calculated by TRNSYS (Klein, 2006), as shown in Fig. 3. The weather data are from the China Standard Weather Data (China Meteorological Bureau and Tsinghua University, 2005), and the building heating load is calculated by DeST (Yan et al., 2008). In order to enhance the convergence and robustness of the model, there are some differences between the TRNSYS model and real system. The real system only need one set of SCs which is used in both Mode SD and Mode SAHP, while the model contains two SC modules with the same parameters, which are used for the two operation modes separately. The two AHX modules are handled in the same way. The type numbers of the modules are listed in Table 1. The efficiency of the SC is evaluated using the following equation from a test report of a flat plate SC product (National Center for Quality Supervision and Testing of Solar Heating Systems (Beijing), 2017). coll
= 0.775
5.559
(Ti Ta ) Isolar
In the TRNSYS model, the frosting and defrosting processes is not included. Therefore, the models of these two processes are developed in this study. 3.2.1. Frosting process The frosting process is considered a quasi-steady state. A fin-tube heat exchanger is used as the AHX, each row of which is assumed to be homogeneous. The frosting-defrosting period is set to 60 min, including a frosting period of 54 min and a defrosting period of 6 min. The calculation flow diagrams for both a single time step and the entire heat exchanger frosting period are shown in Figs. 6 and 7, respectively. The input parameters are the structure of the heat exchanger, ambient air parameters (including dry bulb temperature, relative humidity, and air velocity), and temperature of the fluid inside the tube. According to Yang et al. (2006), both the latent and sensible convection heat transfers are calculated for each time step and heat exchanger row. The frosting condition for each time step and each heat exchanger row is calculated several times, after which the heat exchange rate, frost density, and frost thickness are obtained. The equations and correlations are referred to in Yang et al. (2006). The sensible convection heat transfer between the surface of the frost and the air is determined as:
(1)
where coll is the SC efficiency; Ti is the inlet water temperature of the SC, °C; Ta is the ambient air temperature, °C; and Isolar is the total solar irradiation on a tilted surface, W/m2. The performance of the water-water heat pump is set based on a real product (Yantai Land Air-conditioning Industry Co., Ltd., 2015), Because the system is designed for space heating by fan coil unit, whose inlet water temperature is 45 °C in design condition, the water tank is set at the same temperature. The water entering the condenser of the heat pump is from water tank, so the inlet water temperature of the condenser is equal to the tank, which is 45 °C. As shown in Fig. 4, the COP increases with inlet temperature of evaporator from 2.5 to 4.3, and heating capacity has similar characteristic. It should be noted that the heating capacity is only adopted in the performance calculation in Section 4, and it changes based on the city in Section 5 to meet the different heating demands for various meteorological conditions. The performance of the solar-air HSHP during a period is evaluated by seasonal performance factor (SPF) (Malenković, 2013)
SPF =
Qsen = Qsen, fin + Qsen, tube =
Tfin, f ) + htube Atube (Ta
Ttube, f ) (3)
where Qsen, fin is the sensible convection heat transfer to the heat exchanger fin, W; Qsen, tube is the heat transfer to the tube, W; fin is the fin efficiency; hfin and htube are the heat transfer coefficients from the ambient air and frost surface to the fin and tube, respectively, W/(m2·K); Afin and Atube are total area of the fin and tube, respectively, m2; Ta is the inlet air temperature, K; and Tfin, f and Ttube, f are the temperatures of the frost surface on the fin and tube, respectively, K. The heat transfer coefficients are computed using Eqs. (4) and (5):
Nufin = hfin L /ka = 0.204ReL0.657 Pr 1.334
Qsys· dt Psys· dt
fin h fin Afin (Ta
(4)
and
(2)
(5)
Nutube = htube Dh / ka = 0.146ReL0.917 Pr 2.844
where Qsys is the heat production of the system, W; Psys is the total electric power of the system, including heat pump unit and water pumps, W. To consider the electricity used for water distribution in solar system, the empirical values are adopted in the model to reflect the power use for water distribution in typical systems. The water temperature difference between inlet and outlet of the solar collector is 5 °C, based on which the flow rate is calculated. The pump head is 20 m, and pump efficiency is 0.7. The control strategy is set following the principle in Section 2, as shown in Fig. 5. The input parameters are as follows: T_load (temperature to the building load in °C); T_solar (outlet temperature of the collector in °C); and I_solar (solar intensity on a tilted surface in W/m2). To prevent the water tank from overheating, the upper limit of the
where Nufin and Nutube are Nusselt numbers of the fin and tube, respectively; L is the length of the cooling plate, mm; Dh is the hydraulic Table 1 TRNSYS type number of the model.
351
Object
Type number
Heat pump unit SC for Mode SD, SC for Mode SAHP AHX for Mode AHP, AHX for Mode SAHP Water tank Control on/off Pumps Weather Data
668 1 5 4 2 3 15
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Fig. 4. Heating capacity and COP of the heat pump unit (Yantai Land Airconditioning Industry Co., Ltd., 2015).
diameter of the tube, mm; ka is the thermal conductivity, W/(m·K); and Re and Pr are the Reynolds number and Prandtl number, respectively. The latent heat transfer between the frost surface and moist air is calculated as follows:
Qlat = Qlat , fin + Qlat , tube =
fin Hfin Afin (wa
wfin, f ) q + Htube Atube (wa
wtube, f ) q
(6)
Fig. 6. Flow diagram of calculation in a single time step.
where Qlat , fin is the latent heat transfer to the heat exchanger fin, W; Qlat , tube is the latent heat transfer to the tube, W; Hfin and Htube are the mass transfer coefficients of the fin and tube, respectively, kg/(m2·s); q is the latent heat of sublimation, J/kg; wa is the absolute humidity of the air, kg/kg dry air; and wfin, f and wtube, f are the absolute humidity of the surface of frost on the fin and tube, respectively, kg/kg dry air. The mass transfer coefficients are calculated using the analogy between heat and mass transfer as Eq. (7):
H=
h cp, a Le 2/3
Qfin = Qsen, fin + Qlat , fin
(8)
and
Qtube = Qsen, tube + Qlat , tube
(9)
where Qfin is the total heat transfer of the fin, W; and Qtube is the total heat transfer of the tube, W. In addition to the heat and mass transfers between the frost and moist air, the heat conduction inside the frost layer is considered using the following equations. Moreover, the heat transfer between the frost and moist air should be equal to that inside the frost layer.
(7)
where Le is the Lewis number, and cp, a is the specific heat capacity, J/ (kg·K). The total sensible and latent heat transfer of the fin and tube should be balanced and can be expressed as Eqs. (8) and (9):
Qfin = kf Afin
Tfin, f
Fig. 5. Flow chart of the control strategy. 352
yfin, f
Tfin (10)
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The model of the AHX defrosting process is similar to that of ASHP, except that the total heat consumed is produced by the heat pump. 3.3. Modeling of benchmark systems In Section 5, there will be a comparison between the solar-air HSHP and benchmark systems, i.e. air source heat pump (ASHP) and solar collector plus air source heat pump (SC + ASHP). Therefore, the modeling of the ASHP and SC + ASHP systems is introduced. The performance of the ASHP is set based on a real product. When the outlet water temperature of the condenser is fixed at 45 °C, the COP of the air source heat pump is fitted from the heat pump catalogs provided by the manufacturer
COP = 0.00006Ta3 + 0.0012Ta 2 + 0.0632Ta + 2.941
(13)
where COP is the coefficient of performance of the ASHP, and Ta is the ambient temperature, °C. Based on the COP and heating load, the input power of the ASHP is calculated hourly, which is accumulated to get the total energy use of ASHP system during a heating season. In the SC + ASHP system, the performances of SC and ASHP are the same as those in solar-air HSHP system and ASHP system respectively, to guarantee fair comparisons. The SC is preferred as long as the outlet temperature of the SC is higher than the inlet temperature. The ASHP is used as an auxiliary heat source in case the SC cannot provide heat during the night. 4. Performance analysis of typical working conditions There are two main advantages of the solar-air HSHP: effective utilization of solar energy, and stable heat supply under defrosting conditions. To evaluate these advantages quantitatively, the system performances of the solar-air HSHP under various solar intensity and defrosting conditions were calculated.
Fig. 7. Flow diagram of calculation in a whole frosting period.
and
Qtube = 2 kf W
Ttube, f
4.1. Performance under various solar intensity
Ttube
ln[(R + ytube, f )/ R]
The area of the SC is set to 300 m2, and the heat transfer capacity of the AHX is 12 kW/K. The solar intensity on the SC surface ranges from 0 W/m2 to 1000 W/m2, and the ambient temperature is between −5 °C and 0 °C. The inlet and outlet temperatures of the SC, AHX, and heat pump are shown in Fig. 8. When the ambient temperature is −5 °C and the solar radiation is above 400 W/m2, the system is in Mode SD. The SC provides heat directly, and the heat pump is off; thus, there are only inlet and outlet water temperatures of SC in the figure, and the outlet temperature is higher than that of the inlet. If the solar irradiation falls below 400 W/m2, which disables the SC direct heating, the system is in Mode SAHP. The outlet temperatures of the SC and AHX are higher than that of the heat pump evaporator outlet temperature, which means they can absorb heat from both solar energy and ambient air. The heat pump evaporator inlet temperature increases with an increase in solar intensity, indicating that solar energy is helpful to improve the performance of the heat pump. When the ambient temperature is 0 °C, the solar intensity switching point of Mode SAHP and Mode SD is lower than that when the ambient air is −5 °C. This is because a higher ambient temperature leads to less SC heat loss, and the SC can directly heat in a lower solar intensity. The temperature differences result in heat transfer, as presented in Fig. 9, and the ambient temperature is fixed at −5 °C and 0 °C. In Mode SD, the heat is entirely from solar energy, which ranges from 20 kW to 140 kW depending on the solar intensity. In Mode SAHP, the heat is from both solar energy and ambient air, which ranges from 0 kW to 90 kW and 10 kW to 70 kW, respectively. As the solar radiation grows stronger, the heat pump extracts more heat from solar energy than from the ambient air, indicating that the solar energy and ambient air are alternative heat sources. The evaporating temperature of the heat pump
(11)
where kf is the thermal conductivity of frost, W/(m·K); and ytube, f and yfin, f are the thickness of the frost layer on the tube and the fin, respectively, mm. Tfin and Ttube are the temperature of the fin and tube, respectively, K; and R is the radius of the tube, mm. The thermal conductivity of frost is calculated using Eq. (12):
kf = 1.202 × 10
3
f
0.963
(12)
where f is the density of the frost, kg/m3. The above equations calculate the heat and mass transfer amounts. Based on the above equations, the frost mass increases, and the heat transfer during the frosting process is calculated. 3.2.2. Defrosting process The defrosting process of the ASHP unit is analyzed with the hot-gas bypass defrosting method. The heat consumption for defrosting is divided into four types: the heat to warm the fin and tube, the heat to warm and melt the frost, the heat to evaporate the residual water after the frost is melted, and the heat loss caused by the temperature difference between the heat exchanger and ambient air. As the defrosting process is too complex to be described by a mathematical model, experimental results are used to estimate the heat consumed by defrosting. According to the experiment results of Han (2007), the average heat used to melt frost makes up 44% of the total heat consumption for defrosting. Thus, the total heat used is calculated according to the frost mass and latent heat of melting (334 kJ/kg for ice). Then the electricity consumed by the heat pump can be obtained by the hot-gas bypass defrosting method. 353
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(a) Ambient temperature of -5 ºC.
Fig. 10. Solar energy utilization of solar-air HSHP and traditional SCs.
the solar energy utilization of the two systems is proposed in Fig. 10. The curves on the right side of the figure (solar intensity higher than 400 W/m2) is the heating capacity of the traditional SC, which is the same as Mode SD of the solar-air HSHP. When solar intensity is lower than 400 W/m2, the traditional SCs cannot use solar energy in this condition due to the ambient heat loss. While the solar-air HSHP can still take advantage of solar energy (the shaded part on the left side), because the collector temperature is lower than ambient in mode SAHP, and the convection heat loss is avoided. The collector efficiency and COP of the heat pump unit are illustrated in Fig. 11, and the efficiency of the SC is defined as the useful heat gain divided by total solar energy on a tilted surface. In Mode SD, the collector efficiency is influenced by solar intensity and ambient temperature, changing from 0.15 to 0.52. In Mode SAHP, due to the absence of heat loss between the SC and ambient air, the SC efficiency increased considerably, which reaches 0.775. The COP of the heat pump unit is approximately 2.5–3.3 depending on the ambient temperature and solar energy.
(b) Ambient temperature of 0 ºC. Fig. 8. Temperatures of the system in various working conditions.
4.2. Performance under the defrosting condition For the conventional ASHP, the heat transfer rate of the ASHP evaporator and condenser under a typical frosting condition (e.g., dry bulb temperature 0 °C, relative humidity 85%) for 2 h was calculated based on the hot-gas bypass defrosting method. As indicated in Fig. 12, between 0 min and 54 min, the heat pump operates in heating mode, and the heating capacity decreases as frosting becomes more serious. During the defrosting period (between 54 min and 60 min), the heat pump cannot provide any heat for the indoor space. If the reverse cycle defrosting method is adopted, the ASHP would even absorb heat from the indoor environment. The inherent drawbacks of defrosting the ASHP would cause an uncomfortable indoor thermal environment. Moreover, to meet the building heating load, a high capacity for the ASHP should be designed to offset the decline in heating capacity during the defrosting process. For the solar-air HSHP system, the indirectly connected heat pump
(a) Ambient temperature of -5 ºC.
(b) Ambient temperature of 0 ºC. Fig. 9. Energy income and outcome of the solar-air HSHP.
increases with an increase in solar radiation and therefore raises the total heat supply by the heat pump condenser. Compared with traditional SCs, the advantage of solar-air HSHP is to use solar energy more efficiently, and the quantitative comparison of
Fig. 11. Collector efficiency and COP of the heat pump unit. 354
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160
Heating capacity (kW)
5. Applicability and design comparison
Heat extracted from ambient Heat production
140
To investigate the performance of the solar-air HSHP system comprehensively, a case study for space heating in different cities in China was conducted. The performances of two traditional systems (SC + ASHP and ASHP) are also calculated as benchmarks. SC + ASHP includes a SC and ASHP connected in parallel, and ASHP is the traditional air source heat pump system.
120 100 80 60 40
5.1. Energy use and applicability for space heating
20 0
0
20
40
60 Time (min)
80
100
The parameters of the main components of solar-air HSHP and SC + ASHP are listed in Table 2. The SC areas are 200 and 300 m2, and the capacity of the heat pump unit is chosen according to the heating load of the different locations. The results of Case 6 from January 15th to 17th are used to analyze the operation mode and performance of the solar-air HSHP system. As shown in Fig. 14, the solar intensity on the tilted surface on January 17th reached 1000 W/m2, representing a typical clear day during which the system is in Mode SD. On January 15th and 16th, the solar intensity was lower than 400 W/m2 most of the time, making it challenging to use solar energy directly from the SC. Therefore, Mode SAHP is frequently in operation during these two days to utilize the weak solar radiation. During the night when solar energy is absent, the system is in Mode AHP to guarantee continuous heat production. Corresponding with Fig. 14, the temperatures in the system and ambient air are shown in Fig. 15. The temperature of the water tank is controlled between 40 °C and 45 °C at night and on cloudy days (January 15th and 16th), and is higher than 45 °C when solar energy is abundant (January 17th), allowing the water tank to store more heat from solar energy. If the heat pump unit is on, the inlet and outlet temperatures of the evaporator are always lower than those of the ambient temperature, indicating that the air heat exchanger is able to absorb heat from ambient air, and there is no heat loss from the SC to ambient heat transfer by convection. The solar-air HSHP system is able to take advantage of both solar
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Fig. 12. Heat production of ASHP in a complete frosting and defrosting period.
system can defrost efficiently using the hot water produced by the system itself. The defrosting period is set to 10 min, and it is assumed that the heat consumption for defrosting the AHX of the solar-air HSHP system is the same as that of ASHP system. Six AHXs are included in this solar-air HSHP system (AHX1, AHX2, …, AHX6), one of which is a standby to deal with the defrosting process. Fig. 13 shows the heat transfer rates of the six AHXs and the heat production of the solar-air HSHP in the same frosting condition (dry bulb temperature 0 °C, relative humidity 85%) as that of the conventional ASHP. It is assumed that the standby AHX is always defrosting during the frosting condition. For example, during 0–10 min, AHX2 operates in defrosting mode and the heating capacity is 0, while the other AHXs are still working. Thus, the solar-air HSHP can always provide continuous heat. After 10 min, the defrosting process has already finished, and AHX2 can start to extract heat from the ambient air. Although there is always one AHX in defrosting mode, the other five AHXs are still able to efficiently absorb heat from the ambient air, ensuring the stability of the solar-air HSHP system, and a smaller heat pump unit capacity would meet the heating demand compared with that of the conventional ASHP system.
Fig. 13. Heat production of solar-air HSHP in a complete frosting and defrosting period.
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Table 2 Parameters of the main system components. Case number
1
2
3
Location Heating season
Lhasa Nov. 15 - Mar. 31
Chengdu Dec.1 - Feb. 28
Solar-air HSHP
Area of SC m2 Capacity of heat pump kW Water tank volume m3 Capacity of AHX kW/°C
200 120 10 30
300 120 10 30
200 90 10 23
SC + ASHP
Area of SC m2 Capacity of heat pump kW Water tank volume m3
200 120 10
300 120 10
200 90 10
4
5
6
7
8
Beijing Nov. 15 - Mar. 15
Shenyang Nov. 1 - Mar. 31
300 90 10 23
200 180 10 45
300 180 10 45
200 270 10 68
300 270 10 68
300 90 10
200 180 10
300 180 10
200 270 10
300 270 10
Fig. 14. Operation modes and solar intensity (Jan.15 – Jan.17).
energy and ambient air, as demonstrated in Fig. 16. Ambient air makes up the greater part of incoming energy, which is mainly used during the night when solar energy is not available. During the day, there are two modes that are able to use solar energy (Mode SD and Mode SAHP). Most of the solar energy is used by Mode SD when the solar energy intensity is sufficient, and remaining solar energy is used by Mode SAHP along with energy from the ambient air. For the other cases, the solar-air HSHP has a clear advantage over the traditional system. Fig. 17 is the total energy used for defrosting during the heating season. The energy for defrosting in the solar-air HSHP system is only 30%–36% of that in the SC + ASHP system and 16%–35% of that in ASHP system, which indicates that there is a large
energy saving potential with the solar-air HSHP in the defrosting condition. The SPF of the solar-air HSHP in heating season is 3.61, 3.27, and 2.45 in Chengdu, Beijing, and Shenyang, respectively, and it can reach 6.92 in Lhasa due to sufficient solar energy resources (Fig. 18). The heat provided by Mode SAHP makes up a considerable part of the total heat production, which is 15%, 11%, and 12% in Chengdu, Beijing, and Shenyang, respectively. Compared with the ASHP system, the energy saving rates of the solar-air HSHP are 9% to 14% greater in cities with medium solar resources (e.g., Beijing, Cases 5 and 6), and it is greater than 50% in locations with abundant solar resources (e.g., Lhasa, Cases 1 and 2). Compared with the traditional SC + ASHP, the energy saving
Fig. 15. Temperatures of the system (Jan.15 – Jan.17). 356
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Fig. 16. Energy income of the system (January 15th–January 17th).
5.2. Comparison between different designs In solar-air HSHP, the area of SC and AHX varies in a wide range, which may be ambiguous in real applications. To explore how it could be used in real projects, the energy use and cost of different selections of SC and AHX area are discussed. ASHP is set as the benchmark system, and Lhasa (abundant solar energy resources) and Beijing (medium solar energy resources) are chosen as the typical location. To be consistent with the results in Section 5.1, the systems are used for space heating for the same heating load, and the capacity of the heat pump unit and water tank volume are also the same as Table 2. As a benchmark, the total electricity use of ASHP in a heating season is 35,196 kWh in Lhasa, and 72,437 kWh in Beijing separately. The energy saving rates of solar-air HSHP are shown in Tables 3 and 4, in which each row represents a certain AHX size (e.g. 7 kW/K, 17 kW/K …), and similarly, each column represents a certain SC area (e.g. 50 m2, 100 m2 …). The table cell is the energy saving rate of the corresponding AHX size and SC area. The system does not contain a SC when the SC area is 0 (the first column in the table). It’s obvious that the energy saving rate increases with the area of SC and AHX, which can reach 55% in Lhasa and 21% in Beijing. However, it also means additional investment, and may be not suitable on the aspect of economical efficiency. Therefore the annual cost is further calculated. The annual cost of ASHP is 32637 CNY in Lhasa, and 25868 CNY in Beijing. Tables 5 and 6 presents the increase rate of annual cost of each SC and AHX area. The results are mostly below 0, which means the annual cost of the solar-air HSHP is lower than ASHP. With the increase of SC area, the annual cost declines in the beginning because of the retrenchment of energy use, and then rises, because the additional initial cost is more influential than the energy expense. Similar phenomena are also got when area of AHX changes. In conclusion, the annual cost represents both initial investment and operating cost, and the best solution is a balance between the two aspects. The annual cost reaches a minimum value when SC area is 200 m2 and UA value of AHX is 17 kW/K in Lhasa, which is considered as an optimal point. And the annual cost is minimum when SC area is 100 m2 and UA value of AHX is 45 kW/K in Beijing.
Fig. 17. Total energy use for defrosting in the heating season.
Fig. 18. SPF in heating season.
rates of the proposed system in Chengdu and Beijing are remarkable. A comparison is made with other studies on solar assisted heat pumps for heating. Bakirci and Yuksel (2011) reported the system COP of a solar source heat-pump system for residential heating in cold climate region. The average solar radiation in the literature is 15.6 MJ/m2.day, which is higher than that in Beijing (9.0 MJ/m2.day). The average outdoor temperature in the literature is 4.7 °C, and that in Beijing is 0.16 °C. The average COP of the system in the literature is 2.5–2.9, and the SPF of the solar-air HSHP is 3.27. There are some other studies on the solar assisted heat pump for heating, and provided the COP of the system, which ranges from 2.4 to 2.9 (Bi et al., 2004; Zhou et al., 2010). It can be seen that there is an advantage of the proposed system over previous research.
6. Conclusion To overcome the disadvantage of inadequate solar energy utilization in traditional solar heat pump systems, a solar-air HSHP was proposed that is able to efficiently use solar energy of various intensities and steadily supply heat during defrosting. A numerical model was established to evaluate the performance of the system under typical 357
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Table 3 Energy saving rate compared with ASHP in Lhasa (%).
Table 4 Energy saving rate compared with ASHP in Beijing (%).
Table 5 Increase rate of annual cost compared with ASHP in Lhasa (%).
Table 6 Increase rate of annual cost compared with ASHP in Beijing (%).
working conditions and during a heating season, and the following can be concluded from the analysis:
(2) In the defrosting condition, only one of the air heat exchangers is in defrosting mode, and the others can still extract heat from ambient air. The air heat exchangers run in defrosting mode one at a time, allowing the solar-air HSHP to provide heat continuously. Moreover, the energy for defrosting using the solar-air HSHP is only 30%–36% of that in the SC + ASHP system and 16%–35% of that in the ASHP system, which demonstrates the large energy saving potential of the solar-air HSHP system in the defrosting condition. (3) When used for space heating, the SPF of the proposed system can reach 6.92, 3.61, 3.27, and 2.45 in Lhasa, Chengdu, Beijing, and Shenyang, respectively, and the energy saving rate is 9%–52% greater compared with traditional air source heat pump system.
(1) When solar radiation is abundant, the system provides heat to the user directly, without using heat pump electricity. When solar energy is insufficient (e.g., solar intensity is lower than 400 W/m2 or ambient temperature is less than −5 °C), the heat pump extracts heat from both solar energy and the ambient air, which ranges from 0 to 90 kW and 10 to 70 kW, respectively. When used for space heating, the heat provided by Mode SAHP makes up a considerable part of the total heat production, which is 15%, 11%, and 12% in Chengdu, Beijing, and Shenyang, respectively. Thus, weak solar radiation can also be used effectively. 358
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Declaration of Competing Interest
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The authors declare that they have no known competing financial interests or personal relationships that could have appeared to influence the work reported in this paper. Acknowledgement This study is supported by the National Natural Science Foundation of China (Grant No. 51638010) and the Innovative Research Groups of the National Natural Science Foundation of China (Grant No. 51521005) References Badescu, V., 2003. Model of a thermal energy storage device integrated into a solar assisted heat pump system for space heating. Energy Convers. Manage. 44 (10), 1589–1604. Bakirci, K., Yuksel, B., 2011. Experimental thermal performance of a solar source heatpump system for residential heating in cold climate region. Appl. Therm. Eng. 31 (8–9), 1508–1518. Bi, Y., Guo, T., Zhang, L., Chen, L., 2004. Solar and ground source heat-pump system. Appl. Energy 78 (2), 231–245. Caglar, A., Yamali, C., 2012. Performance analysis of a solar-assisted heat pump with an evacuated tubular collector for domestic heating. Energy Build. 54, 22–28. Chen, X., Yang, H., 2012. Performance analysis of a proposed solar assisted ground coupled heat pump system. Appl. Energy 97, 888–896. China Meteorological Bureau, C.I.C., Climate Data Office, Tsinghua University, D.o.B.S.a. T., 2005. China Standard Weather Data for Analyzing Building Thermal Conditions. China Building Industry Publishing House, Beijing. Dikici, A., Akbulut, A., 2008. Performance characteristics and energy-exergy analysis of solar-assisted heat pump system. Build. Environ. 43 (11), 1961–1972. Esen, H., Esen, M., Ozsolak, O., 2017. Modelling and experimental performance analysis of solar-assisted ground source heat pump system. J. Exp. Theor. Artif. Intell. 29 (1), 1–17. Esen, M., 2000. Thermal performance of a solar-aided latent heat store used for space heating by heat pump. Sol. Energy 69 (1), 15–25. Esen, M., Esen, H., 2005. Experimental investigation of a two-phase closed thermosyphon solar water heater. Sol. Energy 79 (5), 459–468. Esen, M., Yuksel, T., 2013. Experimental evaluation of using various renewable energy sources for heating a greenhouse. Energy Build. 65, 340–351. Han, Z., 2007. Experimental study on characteristics of conventional defrosting and energy storage defrosting of air source heat pump. Harbin Institute of Technology. Huang, D., Li, Q.X., Yuan, M.L., 2009. Comparison between hot-gas bypass defrosting and reverse-cycle defrosting methods on an air-to-water heat pump. Appl. Energy 86 (9), 1697–1703. Jiang, Y., 2011. 2011 Annual Report on China Building Energy Efficiency. China
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A solar-air hybrid source heat pump for space heating and domestic hot water
T
Siyuan Rana,b, Xianting Lia, , Wei Xub, Baolong Wanga ⁎
a b
Department of Building Science, School of Architecture, Tsinghua University, Beijing 100084, People’s Republic of China China Academy of Building Research, Beijing 100013, People’s Republic of China
ARTICLE INFO
ABSTRACT
Keywords: Solar energy Air source heat pump Defrost Hybrid source heat pump Space heating
Traditional solar collectors (SCs) cannot effectively use solar energy of varying intensity, and air source heat pumps cannot supply heat steadily during defrosting conditions. Accordingly, a solar-air hybrid source heat pump (HSHP) system is proposed to solve these problems. A SC was indirectly connected to a heat pump evaporator and water tank for a user at the same time, to allow direct heating when the solar energy is abundant and to supply heat to the heat pump. Multiple air heat exchangers were also connected indirectly to the heat pump. When one of these is in defrosting mode, the others can still work to guarantee a continuous heat output. The structure and operation strategies are introduced, and a numerical model is established to evaluate the performance of the solar-air HSHP. The performance was also compared with a traditional SC plus an air source heat pump (SC + ASHP). The results show that: (1) the heat provided by the solar-air source heat pump mode is 15%, 11%, and 12% of the total heat production in Chengdu, Beijing, and Shenyang, respectively, which indicates that weak solar radiation can also be used effectively; (2) the energy for defrosting using a solar-air HSHP is only between 30% and 36% of that when using a SC + ASHP system, and between 16% and 35% of that in the ASHP system, which indicates a large energy savings potential of the system in defrosting conditions; and (3) the seasonal performance factor (SPF) during a heating season is 3.61, 3.27, and 2.45 in Chengdu, Beijing, and Shenyang, respectively, and it can reach up to 6.92 in Lhasa due to sufficient solar energy resources.
1. Introduction The energy use for space heating takes up 23% of total energy use for buildings in northern China, and hot water uses almost a quarter of the energy in residential buildings (Jiang, 2011). Moreover, energy use is growing rapidly due to the increased intensity of energy use and increased building floor space. In contrast, China is limiting total primary energy use to less than 4.8 billion tons of standard coal equivalent (tce), and the ceiling for building energy use is 1.1 billion tce (Peng et al., 2015). Thus, renewable and clean energy should be introduced to resolve the conflict between energy use demands and limitations. Solar thermal energy is a potential solution to these problems, and is widely used across the world (Tian et al., 2019). However, there are some disadvantages with conventional solar thermal systems. First, solar collectors (SCs) can only supply hot water at a high capacity when solar radiation is sufficient, meaning they cannot supply hot water during the night. Therefore, an auxiliary heat source (driven by electricity or gas) is typically attached to the system. Second, the temperature of the solar absorber must be higher than the temperature of
⁎
the hot water provided to the users, which is in turn higher than that of ambient air, and heat loss occurs by convection from the SC to the ambient air. If solar radiation is weak (e.g. 0–300 W/m2), the heat from solar is mostly (or even completely) lost; therefore, weak solar radiation cannot be used efficiently in traditional SCs. To overcome these shortcomings, a solar assisted heat pump system has been suggested by many researches that combines the SC and an air source heat pump (ASHP). Sterling and Collins (2012) found that electrical use and operating costs of solar assisted heat pump were the lowest with the proposed system when compared with traditional electric and solar hot water systems. Panaras et al. (2014) introduced a method to predict the thermal energy provided to the user and the electrical energy consumed by the heat pump, which is based on dynamic system testing results. Thermal storage, an important part in solar heat pump systems, was calculated and optimized by Badescu (2003). The results showed that smaller storage units provide greater heat flux to the heat pump, while the larger units provide heat for longer time periods. And phase change material can also be combined with solar powered heat pump systems
Corresponding author. E-mail address: [email protected] (X. Li).
https://doi.org/10.1016/j.solener.2020.02.038 Received 27 November 2019; Received in revised form 15 January 2020; Accepted 9 February 2020 0038-092X/ © 2020 International Solar Energy Society. Published by Elsevier Ltd. All rights reserved.
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density, kg/m3
Nomenclature
A cp, a Dh h H Isolar k L Le Nu P Pr q Q R Re T w y
Abbreviations
area, m2 specific heat capacity, J/(kg·K) hydraulic diameter of the tube, mm heat transfer coefficient, W/(m2·K) mass transfer coefficient, kg/(m2·s) total solar irradiation on tilted surface, W/m2 thermal conductivity, W/(m·K) length of the cooling plate, mm Lewis number Nusselt number electric power, W Prandtl number latent heat of sublimation, J/kg heat transfer capacity, W radius of the tube of the AHX, mm Reynolds number temperature, °C absolute humidity, kg/kg dry air thickness, mm
AHX air heat exchanger ASHP air source heat pump COP coefficient of performance Mode AHP air source heat pump mode Mode SAHP solar air source heat pump mode Mode SD solar direct heating mode HSHP hybrid source heat pump SC solar collector SPF seasonal performance factor Subscripts a coll f fin i sen sys tube
Greek letters efficiency (Esen, 2000). There are also solar heat pump systems coupled with geothermal pipes. Chen and Yang (2012) optimized the design of a solar assisted ground coupled heat pump, showing that the annual total heat extraction (plus 75% of the hot water requirement) can be provided by solar energy. Simulation models were built for solar-assisted ground source heat pump systems, and the performances of the systems with different types of ground heat exchangers are compared (Esen et al., 2017). In addition to the simulation mentioned above, there are also laboratory tests for solar heat pump systems. Dikici and Akbulut (2008) tested a solar assisted heat pump with flat plate collectors for space heating for a 60 m2 room in Turkey. The coefficient of performance (COP) reached 3.08 while the exergy loss of the SC was 1.92 kW. A twophase thermosyphon solar collector using refrigerants R-134a, R407C, and R410A was tested and compared (Esen and Esen, 2005). Another experimental set-up was constructed with a liquid-to-liquid vapor compression heat pump and twelve flat-plate SCs (Bakirci and Yuksel, 2011). Caglar and Yamali (2012) designed a solar heat pump system and evaluated its performance both theoretically and experimentally. The evaporating temperature varied between 5.2 and 20.7 °C, while storage tank temperature varied between 9 and 35 °C and the COP was 6.38. When the solar heat pump system was integrated into a passive house, it consumed electricity as its convectional energy source, which only accounted for about one-third of the total energy supplied for heating (Shan et al., 2016). Esen and Yuksel (2013) experimentally investigated greenhouse heating system including solar, biogas and ground energy. In the above research, solar assisted heat pumps were mainly divided into two categories according to the connection of the SC and heat pump: series connected and parallel connected. In series connected solar heat pumps, the SC is the heat source of the heat pump unit. This type of system cannot make full use of solar energy because it still needs to consume electricity, even when solar energy is abundant. Moreover, it cannot supply hot water during the night; therefore, it is mainly used for domestic hot water, which means having a water tank is essential. In parallel connected systems, the SCs supply hot water directly when solar energy is sufficient, but cannot take advantage of weak solar
ambient air solar collector frost/frost surface fin of the heat exchanger inlet sensible convection heat transfer system tube of the heat exchanger
energy. Based on the above systems, some researchers proposed a solar air heat pump system that is able to use solar energy effectively. In effect, this means directly heating when solar energy is sufficient, and providing heat to the heat pump evaporator unit when solar intensity is weak. Liu et al. developed a solar air composite heat source heat pump system with a three medium composite heat exchanger as the evaporator of the system. The heat exchanger was made by inserting another heat exchanger tube into the tube of a fin tube heat exchanger, which allows the system to use either (or both) the air source and solar heat as the heat source (Liu et al., 2012). The system contains three operation modes: single air source, solar-air dual-source, and single solar. Further, the optimal control strategy in different working conditions was studied (Zhang et al., 2011). To investigate system performance further, an experiment table was established, and the results showed that when the ambient temperature was −15 °C, compared to the single air heat source mode, the dual heat source mode heat capacity increased by 62% and the COP increased by 59% (Liu et al., 2016). When the ambient temperature is −7 °C, the heating capacity and COP of the solar air source mode increase by approximately 24% and 25%, respectively (Zhou et al., 2010). Furthermore, a new three medium composite heat exchanger has been designed to make this process easier and to make batch production more convenient (Zhang et al., 2012). Although this type of system can efficiently utilize solar energy and heat from ambient air, the structure of the evaporator is too complex for commercial viability. Moreover, the control strategy did not include the direct use of solar energy. Ni et al. (2016b) proposed a solar assisted air source heat pump with phase change material (PCM), consisting of an air source heat pump, a triple-sleeve energy storage exchanger, and a solar thermal collector. The system enhances performance at low ambient temperatures and eliminates the discrepancy between the demand and supply sides. To test the system performance under various conditions, a prototype was set-up and implemented, and the results showed that the heating COP of the proposed system increased by 65% compared to the ASHP system when the outdoor air temperature was below 10 °C (Qv et al., 2015). The performance in severe external conditions was tested further (Ni et al., 2016a), and the results indicated that there was a great 348
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improvement in both the system efficiency and heating and cooling capacity in the severe external conditions. As the triple-sleeve energy storage exchanger is the core component of the system, a mathematical model was established to analyze the impact parameters (Niu et al., 2011). In addition, the integrated system cool storage performance was studied (Niu et al., 2012). However, there were disadvantages similar to the previous system: (1) the structure of the triple-sleeve energy storage exchanger is complex; and (2) the operation modes did not include a direct solar heating mode, which decreases the efficiency when solar energy is sufficient. As air source heat exchanger is essential in the solar assisted heat pumps introduced above, and there exists a frosting problem when it works in low temperature and high humidity condition. Although the defrosting methods have been proposed for many years and are widely used nowadays, there are still some drawbacks. There are two kinds of defrosting method for ASHP: reverse-cycle defrosting and hot-gas bypass defrosting (Huang et al., 2009). When the ASHP is in reverse-cycle defrosting mode, it extracts heat from indoor air (or circulating hot water), which has a severer impact on indoor thermal comfort. As for the hot-gas bypass defrosting method, the ASHP can provide little heat in defrosting mode. Moreover, the heat used for melting the ice is from electricity, which is not energy efficient. In the above researches on solar assisted heat pumps, there is no newly proposed defrosting method. Therefore, there is a need for a new defrosting method to improve thermal comfort and reduce energy use. In this study, a solar-air hybrid source heat pump (HSHP) system is proposed to efficiently use solar energy, and a novel defrosting method is applied to the system, allowing it to provide steady heat during defrosting conditions and decreasing the energy used in defrosting.
When there is no solar radiation, all the valves that related to the SC are closed, and the AHXs valves are opened (Fig. 2c). The AHXs are used to extract heat from the air and provide the low-grade heat to the heat pump unit to produce hot water for users. (4) Defrosting mode In high humidity and low temperatures, the defrosting mode is essential, and the AHXs run the defrosting process individually. Only one AHX is in operation during the defrosting process at one time, and the other AHXs still operate in the heating mode and extract heat from the ambient air. Via the water-water heat exchanger, part of the heat from the water tank is used for the AHX in the defrosting process. After the defrosting process of one AHX is finished, another AHX begins the defrosting process, and the other AHXs continue to operate normally. According to the structure of the solar-air HSHP introduced above, there are two points of originality: (1) The AHX and SC are both indirectly connected to the evaporator of the heat pump, in the meanwhile the SC is connected to the hot water tank via a water-water heat exchanger; (2) A novel defrosting method is proposed suitable for the system. Thus two benefits are brought by the originalities: stable heating supply during defrosting condition, and efficient utilization of weak solar energy, which are explained as follows: Since only one AHX is undergoing the defrosting process and the other AHXs are operating normally, the heating capacity can be maintained for users at to nearly the same extent as that without frost. Thus, the comfort and reliability of heating can be guaranteed. Moreover, less energy is consumed by the defrosting process because the hot water produced by the solar-air HSHP itself is used for defrosting. Therefore, the drawbacks of using a traditional ASHP in winter for heating can be significantly mitigated. The solar-air HSHP system can perform all the following functions: (1) extracting heat from the ambient air efficiently and without solar radiation, (2) utilizing both solar energy and the ambient air when there is weak solar radiation, and (3) direct solar heating when there is sufficient solar radiation. Therefore, compared with the traditional solar heating system, the proposed solar-air HSHP system can fully use solar radiation with different intensities. Compared with the traditional ASHP system, the inlet water temperature of the heat pump evaporator increases due to the utilization of weak solar radiation, which is beneficial to the performance of the heat pump unit. Compared with the solar assisted air source heat pump in previous research, the components of the solar-air HSHP (i.e. SC, AHX, heat pump unit and water tank) are widely used products, which significantly reduces the cost and will contribute to the popularity. In general, this type of solar-air HSHP system has great application potential in areas with abundant solar energy resources.
2. Concept of the solar-air HSHP A solar-air hybrid source heat pump (HSHP) system was developed to use solar energy efficiently and stably heat a building in defrosting conditions. As indicated in Fig. 1, the system mainly consists of several air heat exchangers (AHXs), a solar collector (SC), a water-water heat exchanger, and a heat pump unit. The evaporator of the heat pump unit and the SC are connected by secondary fluid. The SC is connected with both the evaporator of the heat pump unit and the water tank through the water-water heat exchanger, which enables the SC to supply heat either to the heat pump or directly to the user. The AHXs are connected to the evaporator and act as the heat source of the heat pump. Meanwhile, heat can be transferred to the AHXs via the water-water heat exchanger for defrosting purposes. There are four operation modes of the solar-air HSHP, based on the weather conditions. (1) Solar direct heating mode (Mode SD) When solar energy is sufficient and the temperature of the SC is higher than that of the water tank, the plate heat exchanger transfers the heat harvested by the SC directly to users via the water-water heat exchanger (Fig. 2a). The greatest advantage of the Mode SD is that it consumes no electricity because solar energy is used directly. (2) Solar air source heat pump mode (Mode SAHP) When solar radiation is weak, the valves for Mode SD are closed, and the valves of the AHXs and SC (used to extract heat from air) are opened (Fig. 2b). Thus, low-grade heat can be extracted from both the air and solar energy using the AHXs and SC, respectively. Then, highgrade heat is supplied to users after the heat pump unit process is completed. (3) Air source heat pump mode (Mode AHP)
Fig. 1. Schematic diagram of the solar-air HSHP system. 349
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(a) Mode SD.
(b) Mode SAHP.
(c) Mode AHP.
(d) Defrosting mode.
Fig. 2. Operation modes of the solar-air HSHP system.
Fig. 3. Diagram of the TRNSYS model. 350
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3. Modeling the system
water tank is 80 °C. If T_load is lower than 80 °C and T_solar is higher than T_load, then the temperature of the SC is high enough for direct heating, and the system is in Mode SD. In another case, the model examines the on/off functioning of the heat pump unit for the previous time step and then compares the T_load with minimum and maximum thresholds of 40 and 45 °C, respectively. The aim of these two steps is to make sure the water tank temperature is between 40 and 45 °C, and to decide whether the heat pump unit should be turned on. If the heat pump should turn on and there is solar radiation, the system is in Mode SAHP; otherwise, it is in Mode AHP. The output of this process determines which mode the system should use.
A mathematical model was built to evaluate the system performance under different conditions, including the defrosting condition. Besides, the models for the benchmark systems are introduced. The system electricity use consists of heat pump and circulating water pump for SC and AHX, and excludes the electricity use of fan coil unit and the circulating water pump for it. Both the models of benchmark system and the solar-air HSHP follow this rule for fair comparison. 3.1. TRNSYS model and control strategy
3.2. Modeling of frosting and defrosting process of AHX
The first step is to build a model that can calculate the performance without regard to the defrosting condition. The performance of the solar-air HSHP is calculated by TRNSYS (Klein, 2006), as shown in Fig. 3. The weather data are from the China Standard Weather Data (China Meteorological Bureau and Tsinghua University, 2005), and the building heating load is calculated by DeST (Yan et al., 2008). In order to enhance the convergence and robustness of the model, there are some differences between the TRNSYS model and real system. The real system only need one set of SCs which is used in both Mode SD and Mode SAHP, while the model contains two SC modules with the same parameters, which are used for the two operation modes separately. The two AHX modules are handled in the same way. The type numbers of the modules are listed in Table 1. The efficiency of the SC is evaluated using the following equation from a test report of a flat plate SC product (National Center for Quality Supervision and Testing of Solar Heating Systems (Beijing), 2017). coll
= 0.775
5.559
(Ti Ta ) Isolar
In the TRNSYS model, the frosting and defrosting processes is not included. Therefore, the models of these two processes are developed in this study. 3.2.1. Frosting process The frosting process is considered a quasi-steady state. A fin-tube heat exchanger is used as the AHX, each row of which is assumed to be homogeneous. The frosting-defrosting period is set to 60 min, including a frosting period of 54 min and a defrosting period of 6 min. The calculation flow diagrams for both a single time step and the entire heat exchanger frosting period are shown in Figs. 6 and 7, respectively. The input parameters are the structure of the heat exchanger, ambient air parameters (including dry bulb temperature, relative humidity, and air velocity), and temperature of the fluid inside the tube. According to Yang et al. (2006), both the latent and sensible convection heat transfers are calculated for each time step and heat exchanger row. The frosting condition for each time step and each heat exchanger row is calculated several times, after which the heat exchange rate, frost density, and frost thickness are obtained. The equations and correlations are referred to in Yang et al. (2006). The sensible convection heat transfer between the surface of the frost and the air is determined as:
(1)
where coll is the SC efficiency; Ti is the inlet water temperature of the SC, °C; Ta is the ambient air temperature, °C; and Isolar is the total solar irradiation on a tilted surface, W/m2. The performance of the water-water heat pump is set based on a real product (Yantai Land Air-conditioning Industry Co., Ltd., 2015), Because the system is designed for space heating by fan coil unit, whose inlet water temperature is 45 °C in design condition, the water tank is set at the same temperature. The water entering the condenser of the heat pump is from water tank, so the inlet water temperature of the condenser is equal to the tank, which is 45 °C. As shown in Fig. 4, the COP increases with inlet temperature of evaporator from 2.5 to 4.3, and heating capacity has similar characteristic. It should be noted that the heating capacity is only adopted in the performance calculation in Section 4, and it changes based on the city in Section 5 to meet the different heating demands for various meteorological conditions. The performance of the solar-air HSHP during a period is evaluated by seasonal performance factor (SPF) (Malenković, 2013)
SPF =
Qsen = Qsen, fin + Qsen, tube =
Tfin, f ) + htube Atube (Ta
Ttube, f ) (3)
where Qsen, fin is the sensible convection heat transfer to the heat exchanger fin, W; Qsen, tube is the heat transfer to the tube, W; fin is the fin efficiency; hfin and htube are the heat transfer coefficients from the ambient air and frost surface to the fin and tube, respectively, W/(m2·K); Afin and Atube are total area of the fin and tube, respectively, m2; Ta is the inlet air temperature, K; and Tfin, f and Ttube, f are the temperatures of the frost surface on the fin and tube, respectively, K. The heat transfer coefficients are computed using Eqs. (4) and (5):
Nufin = hfin L /ka = 0.204ReL0.657 Pr 1.334
Qsys· dt Psys· dt
fin h fin Afin (Ta
(4)
and
(2)
(5)
Nutube = htube Dh / ka = 0.146ReL0.917 Pr 2.844
where Qsys is the heat production of the system, W; Psys is the total electric power of the system, including heat pump unit and water pumps, W. To consider the electricity used for water distribution in solar system, the empirical values are adopted in the model to reflect the power use for water distribution in typical systems. The water temperature difference between inlet and outlet of the solar collector is 5 °C, based on which the flow rate is calculated. The pump head is 20 m, and pump efficiency is 0.7. The control strategy is set following the principle in Section 2, as shown in Fig. 5. The input parameters are as follows: T_load (temperature to the building load in °C); T_solar (outlet temperature of the collector in °C); and I_solar (solar intensity on a tilted surface in W/m2). To prevent the water tank from overheating, the upper limit of the
where Nufin and Nutube are Nusselt numbers of the fin and tube, respectively; L is the length of the cooling plate, mm; Dh is the hydraulic Table 1 TRNSYS type number of the model.
351
Object
Type number
Heat pump unit SC for Mode SD, SC for Mode SAHP AHX for Mode AHP, AHX for Mode SAHP Water tank Control on/off Pumps Weather Data
668 1 5 4 2 3 15
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Fig. 4. Heating capacity and COP of the heat pump unit (Yantai Land Airconditioning Industry Co., Ltd., 2015).
diameter of the tube, mm; ka is the thermal conductivity, W/(m·K); and Re and Pr are the Reynolds number and Prandtl number, respectively. The latent heat transfer between the frost surface and moist air is calculated as follows:
Qlat = Qlat , fin + Qlat , tube =
fin Hfin Afin (wa
wfin, f ) q + Htube Atube (wa
wtube, f ) q
(6)
Fig. 6. Flow diagram of calculation in a single time step.
where Qlat , fin is the latent heat transfer to the heat exchanger fin, W; Qlat , tube is the latent heat transfer to the tube, W; Hfin and Htube are the mass transfer coefficients of the fin and tube, respectively, kg/(m2·s); q is the latent heat of sublimation, J/kg; wa is the absolute humidity of the air, kg/kg dry air; and wfin, f and wtube, f are the absolute humidity of the surface of frost on the fin and tube, respectively, kg/kg dry air. The mass transfer coefficients are calculated using the analogy between heat and mass transfer as Eq. (7):
H=
h cp, a Le 2/3
Qfin = Qsen, fin + Qlat , fin
(8)
and
Qtube = Qsen, tube + Qlat , tube
(9)
where Qfin is the total heat transfer of the fin, W; and Qtube is the total heat transfer of the tube, W. In addition to the heat and mass transfers between the frost and moist air, the heat conduction inside the frost layer is considered using the following equations. Moreover, the heat transfer between the frost and moist air should be equal to that inside the frost layer.
(7)
where Le is the Lewis number, and cp, a is the specific heat capacity, J/ (kg·K). The total sensible and latent heat transfer of the fin and tube should be balanced and can be expressed as Eqs. (8) and (9):
Qfin = kf Afin
Tfin, f
Fig. 5. Flow chart of the control strategy. 352
yfin, f
Tfin (10)
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The model of the AHX defrosting process is similar to that of ASHP, except that the total heat consumed is produced by the heat pump. 3.3. Modeling of benchmark systems In Section 5, there will be a comparison between the solar-air HSHP and benchmark systems, i.e. air source heat pump (ASHP) and solar collector plus air source heat pump (SC + ASHP). Therefore, the modeling of the ASHP and SC + ASHP systems is introduced. The performance of the ASHP is set based on a real product. When the outlet water temperature of the condenser is fixed at 45 °C, the COP of the air source heat pump is fitted from the heat pump catalogs provided by the manufacturer
COP = 0.00006Ta3 + 0.0012Ta 2 + 0.0632Ta + 2.941
(13)
where COP is the coefficient of performance of the ASHP, and Ta is the ambient temperature, °C. Based on the COP and heating load, the input power of the ASHP is calculated hourly, which is accumulated to get the total energy use of ASHP system during a heating season. In the SC + ASHP system, the performances of SC and ASHP are the same as those in solar-air HSHP system and ASHP system respectively, to guarantee fair comparisons. The SC is preferred as long as the outlet temperature of the SC is higher than the inlet temperature. The ASHP is used as an auxiliary heat source in case the SC cannot provide heat during the night. 4. Performance analysis of typical working conditions There are two main advantages of the solar-air HSHP: effective utilization of solar energy, and stable heat supply under defrosting conditions. To evaluate these advantages quantitatively, the system performances of the solar-air HSHP under various solar intensity and defrosting conditions were calculated.
Fig. 7. Flow diagram of calculation in a whole frosting period.
and
Qtube = 2 kf W
Ttube, f
4.1. Performance under various solar intensity
Ttube
ln[(R + ytube, f )/ R]
The area of the SC is set to 300 m2, and the heat transfer capacity of the AHX is 12 kW/K. The solar intensity on the SC surface ranges from 0 W/m2 to 1000 W/m2, and the ambient temperature is between −5 °C and 0 °C. The inlet and outlet temperatures of the SC, AHX, and heat pump are shown in Fig. 8. When the ambient temperature is −5 °C and the solar radiation is above 400 W/m2, the system is in Mode SD. The SC provides heat directly, and the heat pump is off; thus, there are only inlet and outlet water temperatures of SC in the figure, and the outlet temperature is higher than that of the inlet. If the solar irradiation falls below 400 W/m2, which disables the SC direct heating, the system is in Mode SAHP. The outlet temperatures of the SC and AHX are higher than that of the heat pump evaporator outlet temperature, which means they can absorb heat from both solar energy and ambient air. The heat pump evaporator inlet temperature increases with an increase in solar intensity, indicating that solar energy is helpful to improve the performance of the heat pump. When the ambient temperature is 0 °C, the solar intensity switching point of Mode SAHP and Mode SD is lower than that when the ambient air is −5 °C. This is because a higher ambient temperature leads to less SC heat loss, and the SC can directly heat in a lower solar intensity. The temperature differences result in heat transfer, as presented in Fig. 9, and the ambient temperature is fixed at −5 °C and 0 °C. In Mode SD, the heat is entirely from solar energy, which ranges from 20 kW to 140 kW depending on the solar intensity. In Mode SAHP, the heat is from both solar energy and ambient air, which ranges from 0 kW to 90 kW and 10 kW to 70 kW, respectively. As the solar radiation grows stronger, the heat pump extracts more heat from solar energy than from the ambient air, indicating that the solar energy and ambient air are alternative heat sources. The evaporating temperature of the heat pump
(11)
where kf is the thermal conductivity of frost, W/(m·K); and ytube, f and yfin, f are the thickness of the frost layer on the tube and the fin, respectively, mm. Tfin and Ttube are the temperature of the fin and tube, respectively, K; and R is the radius of the tube, mm. The thermal conductivity of frost is calculated using Eq. (12):
kf = 1.202 × 10
3
f
0.963
(12)
where f is the density of the frost, kg/m3. The above equations calculate the heat and mass transfer amounts. Based on the above equations, the frost mass increases, and the heat transfer during the frosting process is calculated. 3.2.2. Defrosting process The defrosting process of the ASHP unit is analyzed with the hot-gas bypass defrosting method. The heat consumption for defrosting is divided into four types: the heat to warm the fin and tube, the heat to warm and melt the frost, the heat to evaporate the residual water after the frost is melted, and the heat loss caused by the temperature difference between the heat exchanger and ambient air. As the defrosting process is too complex to be described by a mathematical model, experimental results are used to estimate the heat consumed by defrosting. According to the experiment results of Han (2007), the average heat used to melt frost makes up 44% of the total heat consumption for defrosting. Thus, the total heat used is calculated according to the frost mass and latent heat of melting (334 kJ/kg for ice). Then the electricity consumed by the heat pump can be obtained by the hot-gas bypass defrosting method. 353
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(a) Ambient temperature of -5 ºC.
Fig. 10. Solar energy utilization of solar-air HSHP and traditional SCs.
the solar energy utilization of the two systems is proposed in Fig. 10. The curves on the right side of the figure (solar intensity higher than 400 W/m2) is the heating capacity of the traditional SC, which is the same as Mode SD of the solar-air HSHP. When solar intensity is lower than 400 W/m2, the traditional SCs cannot use solar energy in this condition due to the ambient heat loss. While the solar-air HSHP can still take advantage of solar energy (the shaded part on the left side), because the collector temperature is lower than ambient in mode SAHP, and the convection heat loss is avoided. The collector efficiency and COP of the heat pump unit are illustrated in Fig. 11, and the efficiency of the SC is defined as the useful heat gain divided by total solar energy on a tilted surface. In Mode SD, the collector efficiency is influenced by solar intensity and ambient temperature, changing from 0.15 to 0.52. In Mode SAHP, due to the absence of heat loss between the SC and ambient air, the SC efficiency increased considerably, which reaches 0.775. The COP of the heat pump unit is approximately 2.5–3.3 depending on the ambient temperature and solar energy.
(b) Ambient temperature of 0 ºC. Fig. 8. Temperatures of the system in various working conditions.
4.2. Performance under the defrosting condition For the conventional ASHP, the heat transfer rate of the ASHP evaporator and condenser under a typical frosting condition (e.g., dry bulb temperature 0 °C, relative humidity 85%) for 2 h was calculated based on the hot-gas bypass defrosting method. As indicated in Fig. 12, between 0 min and 54 min, the heat pump operates in heating mode, and the heating capacity decreases as frosting becomes more serious. During the defrosting period (between 54 min and 60 min), the heat pump cannot provide any heat for the indoor space. If the reverse cycle defrosting method is adopted, the ASHP would even absorb heat from the indoor environment. The inherent drawbacks of defrosting the ASHP would cause an uncomfortable indoor thermal environment. Moreover, to meet the building heating load, a high capacity for the ASHP should be designed to offset the decline in heating capacity during the defrosting process. For the solar-air HSHP system, the indirectly connected heat pump
(a) Ambient temperature of -5 ºC.
(b) Ambient temperature of 0 ºC. Fig. 9. Energy income and outcome of the solar-air HSHP.
increases with an increase in solar radiation and therefore raises the total heat supply by the heat pump condenser. Compared with traditional SCs, the advantage of solar-air HSHP is to use solar energy more efficiently, and the quantitative comparison of
Fig. 11. Collector efficiency and COP of the heat pump unit. 354
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160
Heating capacity (kW)
5. Applicability and design comparison
Heat extracted from ambient Heat production
140
To investigate the performance of the solar-air HSHP system comprehensively, a case study for space heating in different cities in China was conducted. The performances of two traditional systems (SC + ASHP and ASHP) are also calculated as benchmarks. SC + ASHP includes a SC and ASHP connected in parallel, and ASHP is the traditional air source heat pump system.
120 100 80 60 40
5.1. Energy use and applicability for space heating
20 0
0
20
40
60 Time (min)
80
100
The parameters of the main components of solar-air HSHP and SC + ASHP are listed in Table 2. The SC areas are 200 and 300 m2, and the capacity of the heat pump unit is chosen according to the heating load of the different locations. The results of Case 6 from January 15th to 17th are used to analyze the operation mode and performance of the solar-air HSHP system. As shown in Fig. 14, the solar intensity on the tilted surface on January 17th reached 1000 W/m2, representing a typical clear day during which the system is in Mode SD. On January 15th and 16th, the solar intensity was lower than 400 W/m2 most of the time, making it challenging to use solar energy directly from the SC. Therefore, Mode SAHP is frequently in operation during these two days to utilize the weak solar radiation. During the night when solar energy is absent, the system is in Mode AHP to guarantee continuous heat production. Corresponding with Fig. 14, the temperatures in the system and ambient air are shown in Fig. 15. The temperature of the water tank is controlled between 40 °C and 45 °C at night and on cloudy days (January 15th and 16th), and is higher than 45 °C when solar energy is abundant (January 17th), allowing the water tank to store more heat from solar energy. If the heat pump unit is on, the inlet and outlet temperatures of the evaporator are always lower than those of the ambient temperature, indicating that the air heat exchanger is able to absorb heat from ambient air, and there is no heat loss from the SC to ambient heat transfer by convection. The solar-air HSHP system is able to take advantage of both solar
120
Fig. 12. Heat production of ASHP in a complete frosting and defrosting period.
system can defrost efficiently using the hot water produced by the system itself. The defrosting period is set to 10 min, and it is assumed that the heat consumption for defrosting the AHX of the solar-air HSHP system is the same as that of ASHP system. Six AHXs are included in this solar-air HSHP system (AHX1, AHX2, …, AHX6), one of which is a standby to deal with the defrosting process. Fig. 13 shows the heat transfer rates of the six AHXs and the heat production of the solar-air HSHP in the same frosting condition (dry bulb temperature 0 °C, relative humidity 85%) as that of the conventional ASHP. It is assumed that the standby AHX is always defrosting during the frosting condition. For example, during 0–10 min, AHX2 operates in defrosting mode and the heating capacity is 0, while the other AHXs are still working. Thus, the solar-air HSHP can always provide continuous heat. After 10 min, the defrosting process has already finished, and AHX2 can start to extract heat from the ambient air. Although there is always one AHX in defrosting mode, the other five AHXs are still able to efficiently absorb heat from the ambient air, ensuring the stability of the solar-air HSHP system, and a smaller heat pump unit capacity would meet the heating demand compared with that of the conventional ASHP system.
Fig. 13. Heat production of solar-air HSHP in a complete frosting and defrosting period.
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Table 2 Parameters of the main system components. Case number
1
2
3
Location Heating season
Lhasa Nov. 15 - Mar. 31
Chengdu Dec.1 - Feb. 28
Solar-air HSHP
Area of SC m2 Capacity of heat pump kW Water tank volume m3 Capacity of AHX kW/°C
200 120 10 30
300 120 10 30
200 90 10 23
SC + ASHP
Area of SC m2 Capacity of heat pump kW Water tank volume m3
200 120 10
300 120 10
200 90 10
4
5
6
7
8
Beijing Nov. 15 - Mar. 15
Shenyang Nov. 1 - Mar. 31
300 90 10 23
200 180 10 45
300 180 10 45
200 270 10 68
300 270 10 68
300 90 10
200 180 10
300 180 10
200 270 10
300 270 10
Fig. 14. Operation modes and solar intensity (Jan.15 – Jan.17).
energy and ambient air, as demonstrated in Fig. 16. Ambient air makes up the greater part of incoming energy, which is mainly used during the night when solar energy is not available. During the day, there are two modes that are able to use solar energy (Mode SD and Mode SAHP). Most of the solar energy is used by Mode SD when the solar energy intensity is sufficient, and remaining solar energy is used by Mode SAHP along with energy from the ambient air. For the other cases, the solar-air HSHP has a clear advantage over the traditional system. Fig. 17 is the total energy used for defrosting during the heating season. The energy for defrosting in the solar-air HSHP system is only 30%–36% of that in the SC + ASHP system and 16%–35% of that in ASHP system, which indicates that there is a large
energy saving potential with the solar-air HSHP in the defrosting condition. The SPF of the solar-air HSHP in heating season is 3.61, 3.27, and 2.45 in Chengdu, Beijing, and Shenyang, respectively, and it can reach 6.92 in Lhasa due to sufficient solar energy resources (Fig. 18). The heat provided by Mode SAHP makes up a considerable part of the total heat production, which is 15%, 11%, and 12% in Chengdu, Beijing, and Shenyang, respectively. Compared with the ASHP system, the energy saving rates of the solar-air HSHP are 9% to 14% greater in cities with medium solar resources (e.g., Beijing, Cases 5 and 6), and it is greater than 50% in locations with abundant solar resources (e.g., Lhasa, Cases 1 and 2). Compared with the traditional SC + ASHP, the energy saving
Fig. 15. Temperatures of the system (Jan.15 – Jan.17). 356
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Fig. 16. Energy income of the system (January 15th–January 17th).
5.2. Comparison between different designs In solar-air HSHP, the area of SC and AHX varies in a wide range, which may be ambiguous in real applications. To explore how it could be used in real projects, the energy use and cost of different selections of SC and AHX area are discussed. ASHP is set as the benchmark system, and Lhasa (abundant solar energy resources) and Beijing (medium solar energy resources) are chosen as the typical location. To be consistent with the results in Section 5.1, the systems are used for space heating for the same heating load, and the capacity of the heat pump unit and water tank volume are also the same as Table 2. As a benchmark, the total electricity use of ASHP in a heating season is 35,196 kWh in Lhasa, and 72,437 kWh in Beijing separately. The energy saving rates of solar-air HSHP are shown in Tables 3 and 4, in which each row represents a certain AHX size (e.g. 7 kW/K, 17 kW/K …), and similarly, each column represents a certain SC area (e.g. 50 m2, 100 m2 …). The table cell is the energy saving rate of the corresponding AHX size and SC area. The system does not contain a SC when the SC area is 0 (the first column in the table). It’s obvious that the energy saving rate increases with the area of SC and AHX, which can reach 55% in Lhasa and 21% in Beijing. However, it also means additional investment, and may be not suitable on the aspect of economical efficiency. Therefore the annual cost is further calculated. The annual cost of ASHP is 32637 CNY in Lhasa, and 25868 CNY in Beijing. Tables 5 and 6 presents the increase rate of annual cost of each SC and AHX area. The results are mostly below 0, which means the annual cost of the solar-air HSHP is lower than ASHP. With the increase of SC area, the annual cost declines in the beginning because of the retrenchment of energy use, and then rises, because the additional initial cost is more influential than the energy expense. Similar phenomena are also got when area of AHX changes. In conclusion, the annual cost represents both initial investment and operating cost, and the best solution is a balance between the two aspects. The annual cost reaches a minimum value when SC area is 200 m2 and UA value of AHX is 17 kW/K in Lhasa, which is considered as an optimal point. And the annual cost is minimum when SC area is 100 m2 and UA value of AHX is 45 kW/K in Beijing.
Fig. 17. Total energy use for defrosting in the heating season.
Fig. 18. SPF in heating season.
rates of the proposed system in Chengdu and Beijing are remarkable. A comparison is made with other studies on solar assisted heat pumps for heating. Bakirci and Yuksel (2011) reported the system COP of a solar source heat-pump system for residential heating in cold climate region. The average solar radiation in the literature is 15.6 MJ/m2.day, which is higher than that in Beijing (9.0 MJ/m2.day). The average outdoor temperature in the literature is 4.7 °C, and that in Beijing is 0.16 °C. The average COP of the system in the literature is 2.5–2.9, and the SPF of the solar-air HSHP is 3.27. There are some other studies on the solar assisted heat pump for heating, and provided the COP of the system, which ranges from 2.4 to 2.9 (Bi et al., 2004; Zhou et al., 2010). It can be seen that there is an advantage of the proposed system over previous research.
6. Conclusion To overcome the disadvantage of inadequate solar energy utilization in traditional solar heat pump systems, a solar-air HSHP was proposed that is able to efficiently use solar energy of various intensities and steadily supply heat during defrosting. A numerical model was established to evaluate the performance of the system under typical 357
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Table 3 Energy saving rate compared with ASHP in Lhasa (%).
Table 4 Energy saving rate compared with ASHP in Beijing (%).
Table 5 Increase rate of annual cost compared with ASHP in Lhasa (%).
Table 6 Increase rate of annual cost compared with ASHP in Beijing (%).
working conditions and during a heating season, and the following can be concluded from the analysis:
(2) In the defrosting condition, only one of the air heat exchangers is in defrosting mode, and the others can still extract heat from ambient air. The air heat exchangers run in defrosting mode one at a time, allowing the solar-air HSHP to provide heat continuously. Moreover, the energy for defrosting using the solar-air HSHP is only 30%–36% of that in the SC + ASHP system and 16%–35% of that in the ASHP system, which demonstrates the large energy saving potential of the solar-air HSHP system in the defrosting condition. (3) When used for space heating, the SPF of the proposed system can reach 6.92, 3.61, 3.27, and 2.45 in Lhasa, Chengdu, Beijing, and Shenyang, respectively, and the energy saving rate is 9%–52% greater compared with traditional air source heat pump system.
(1) When solar radiation is abundant, the system provides heat to the user directly, without using heat pump electricity. When solar energy is insufficient (e.g., solar intensity is lower than 400 W/m2 or ambient temperature is less than −5 °C), the heat pump extracts heat from both solar energy and the ambient air, which ranges from 0 to 90 kW and 10 to 70 kW, respectively. When used for space heating, the heat provided by Mode SAHP makes up a considerable part of the total heat production, which is 15%, 11%, and 12% in Chengdu, Beijing, and Shenyang, respectively. Thus, weak solar radiation can also be used effectively. 358
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Declaration of Competing Interest
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