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Analysis of a direct expansion solar assisted heat pump using different refrigerants ARTICLE in ENERGY CONVERSION AND MANAGEMENT · SEPTEMBER 2005 Impact Factor: 4.38 · DOI: 10.1016/j.enconman.2004.12.001
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Energy Conversion and Management 46 (2005) 2614–2624 www.elsevier.com/locate/enconman
Analysis of a direct expansion solar assisted heat pump using different refrigerants F.B. Gorozabel Chata, S.K. Chaturvedi *, A. Almogbel Department of Mechanical Engineering, Old Dominion University, Norfolk, Virginia 23529, USA Received 4 March 2004; received in revised form 23 July 2004; accepted 29 December 2004 Available online 25 February 2005
Abstract The thermal performance of a direct expansion solar assisted heat pump (DX-SAHP) is analyzed for several refrigerants using two collector configurations, namely a bare collector and a one cover collector. The REFPROP computer program, developed by the National Institute of Science and Technology, is employed to predict the refrigerant properties involved in the energy balance across the collector. The thermal performance, as characterized by the coefficient of performance (COP), is determined for a variety of pure refrigerants as well as refrigerant mixtures. The performance degradation due to switching from R-12 to pure hydrofluorocarbon (HFC) refrigerants as well as refrigerant blends is investigated. A graphical procedure is developed and illustrated for several refrigerants for sizing the solar collector area and the heat pump compressor displacement capacity for the two collector configurations considered in this study. Ó 2005 Elsevier Ltd. All rights reserved. Keywords: Solar heat pump; New refrigerants; Direct expansion; Graphical method for component sizing
1. Introduction A conventional solar assisted heat pump (SAHP) system employs a solar collector and a heat pump as separate units that are integrated through an intermediate heat exchanger loop for *
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0196-8904/$ - see front matter Ó 2005 Elsevier Ltd. All rights reserved. doi:10.1016/j.enconman.2004.12.001
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transferring the solar energy to the refrigerant loop [1]. In another SAHP configuration, known as the direct expansion solar assisted heat pump (DX-SAHP), the solar collector and the heat pump evaporator are integrated into a single unit [2–16]. The refrigerant is directly expanded in the integrated solar collector–evaporator panel where it undergoes a phase transition from liquid to vapor due to the solar energy input. The present study is concerned primarily with the thermal performance analysis of a direct expansion solar assisted heat pump for a variety of refrigerants and ambient conditions. The performance of a direct expansion (DX) solar assisted heat pump is favorably affected by solar energy because of the integration of the solar collector and the heat pump evaporator into a single unit, henceforth to be referred to as the collector–evaporator panel. Circulation of a refrigerant in the collector reduces the collector temperature and brings it down to levels slightly above ambient temperature. This improves the collector performance since the heat loss from the collector is significantly reduced due to the reduced collector to ambient temperature difference. The heat pump performance is also improved, since it is well known that for a given condensing temperature, a heat pump is more efficient when the evaporator temperature is raised. It should be noted that without the solar coupling, the evaporator temperature would be typically 10–15 °C lower than the ambient temperature. This would translate into a lower coefficient of performance. The critical issue in this system pertains to sizing the solar collector so that its thermal energy collecting capacity is properly matched to the heat pumping (evaporative) capacity of the compressor. The goal generally is to match the collector and the heat pump properly so that the collector–evaporator panel is maintained at 5–10 °C above ambient temperature at the design condition. This allows the advantages pertaining to low collector heat loss and improved heat pump performance to be realized. At off design conditions, the collector–evaporator temperature deviates from that range, but it can be maintained in the range by using compressor capacity modulation [17]. There are several advantages in using the direct solar assisted heat pump instead of the conventional solar assisted heat pump. These include minimal corrosion problems due to the use of refrigerant R-134A or other hydrofluorocarbon refrigerants. This would increase the life of the collector significantly over that of a water based solar collector, whose life is often limited to 10 years due to corrosion problems. Another advantage of using the refrigerant as the collector fluid is that the freeze problem is eliminated due to the very low freezing temperature of refrigerants. In contrast, water collectors are susceptible to the freeze problem unless an additional thermal loop that permits use of an anti-freeze such as an ethylene glycol–water mixture as the collector fluid. Elimination of the intermediate heat exchanger, which is required in the convectional SAHP system, also improves the thermal performance of the DX-SAHP system since the inefficiency inherent in the heat exchange loop is entirely eliminated. Several studies concerning direct expansion solar assisted heat pumps have been performed in the past 25 years [2–17]. However, most of them have used refrigerant R-12 as the working fluid. As a result of the ban on R-12 due to the Montreal agreement [18,19], the results from previous studies are of a very limited value, and studies using newly proposed refrigerants need to be conducted. One such study is by Hawlader et al. [20] who have experimentally and theoretically analyzed a direct expansion solar assisted heat pump designed for the ambient conditions of Singapore using R-134A as the refrigerant. A variable speed compressor was used to ensure proper match between the collector–evaporator load and the compressor capacity. Their results show
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that the performance of the system is influenced by the area of the collector, the speed of the compressor and the solar irradiation. The phase out of refrigerant R-12 makes a majority of studies dealing with the direct expansion solar assisted heat pump system obsolete and not very helpful for future purposes. The goal of the present work is to analyze the thermal performance of a direct expansion solar assisted heat pump with new refrigerants such as R-134a, R-404A, R-410A and R-407C. Specifically, the objectives of this study are: (a) to analyze and characterize the extent of degradation of the thermal performance due to the use of new refrigerants and to identify the refrigerants that are best suited for solar heat pump applications and (b) to devise a simple graphical procedure to determine a collector size (collector area) for the direct expansion solar assisted heat pump system that is properly matched to the chosen compressor heat pumping capacity The above two aspects of direct expansion solar assisted heat pumps are the novel features of this study, and to the best of the authorsÕ knowledge, they have not been described in the existing literature. The graphical procedure, using thermal performance charts, presented in this work is expected to facilitate design of direct expansion solar assisted systems that use environmentally friendly refrigerants. The direct expansion heat pump system, shown in Fig. 1, works on the mechanical vapor compression cycle illustrated in Fig. 2. In this system, the solar collector also performs the function of the heat pump evaporator. The refrigerant evaporates in the collector as it flows through it, resulting in the thermodynamic process 4 ! 1 in the mechanical vapor compression cycle. The evaporated refrigerant is compressed (process 1 ! 2), and the high pressure refrigerant is condensed (2 ! 3), releasing thermal energy to heat the water. This process is followed by the throttling process (3 ! 4) in which high pressure liquid flashes through the thermostatic expansion valve.
Fig. 1. Direct expansion of solar-assisted heat pump.
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Fig. 2. T–S diagram for a single-stage vapor-compression cycle.
2. Theoretical prediction of thermal performance In order to analyze the DX-SAHP cycle, it is important to predict the collector–evaporator temperature (Tf) for a given operating condition (condensing temperature) and ambient conditions (solar insolation and ambient temperature). This can be done by performing an energy balance on the collector–evaporator panel that reflects the fact that the net energy absorbed (solar input minus losses) is manifested as the enthalpy increase to the refrigerant flowing in the collector panel tubes. The steady state energy balance on the collector–evaporator panel can be stated as _ 1 h4 Þ ¼ F 0 Ac ½SðsaÞ U L ðT f T a Þ mðh
ð1Þ
_ 1 h4 Þ is the energy gained by the refrigerant during evaporation, and where mðh 0 F Ac[S(sa)UL(TfTa)] is the net solar energy absorbed by collector–evaporator panel. The symbol Ac is the collector area, S is the solar radiation in the collector plane, F 0 is the collector efficiency factor, UL is the loss coefficient, Ta is the ambient temperature, Tf is the temperature of the fin, m_ is the mass flow rate and h is the enthalpy. Neglecting any frictional pressure drop in the collector, the refrigerant temperature (Tf) inside the collector tube remains constant at the saturation value corresponding to the saturation pressure in the collector. For constant com_ is given by the pressor speed operation, the mass pumped and circulated by the compressor, m, following expression: m_ ¼
gv VD v1
ð2Þ
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The symbol VD represents the volume displacement rate of the compressor, v1 is the specific volume of the refrigerant at the inlet of the compressor and gv is the volumetric efficiency. The energy equation can be expressed as gv VD ðhg ðT f Þ h4 Þ ¼ F 0 Ac ½SðsaÞ U L ðT f T a Þ vg
ð3Þ
gv VD hfg ðT f Þð1 xÞ ¼ Kgcoll Ac S c vs
ð4Þ
or
where x is the quality of the refrigerant after the expansion valve and K is a non-dimensional ratio similar to the clearness index, a characteristic of a given location. It is defined as the ratio of the solar energy S in the collector plane to the solar constant, Sc. The collector efficiency gcoll is given by the expression U L F 0 ðT f T a Þ 0 gcoll ¼ F ðsaÞ S Eq. (4) governs the evaporator/collector temperature (Tf) for a given compressor displacement (VD), collector area (Ac), location (clearness index K and Ta) and refrigerant (hfg and x). Although Eq. (4) can be solved analytically, a graphical procedure is proposed for sizing and matching the direct solar assisted heat pump components, namely VD and Ac. Properties vg, hfg and x are calculated from REFPROP [21], a computer code for predicting the thermodynamic properties of refrigerants. Once Tf is determined for a given S, Ta and refrigerant, thermodynamic analysis of the heat pump is done to determine the compressor work (Wc), thermal energy delivery (QH) and the coefficient of performance (COPH = QH/Wc). A compressor efficiency of 0.70 and a volumetric efficiency of 0.75 was used in the calculations to determine the performance charts.
3. Results and discussion The first part of this study analyzed and characterized the degradation of the thermal performance in a direct expansion solar assisted heat pump system due to the use of new refrigerants in place of refrigerant R-12. In order to accomplish this, a computer program written in C++ language was developed, and several input files obtained from the REFPROP data base program were used. The new refrigerants used in this study are R-134a and the zeotropic refrigerants R-404A, R407-C and R-410A (Table 1). Also, classical refrigerants such as R-12 and R-22 were also analyzed so as to present a comparative analysis involving the new and phased out refrigerants. Fig. 3 shows the variation of the coefficient of thermal performance for collector temperatures ranging from 0 to 20 °C for the several refrigerants mentioned earlier. The condensing temperature is maintained at 60 °C. This figure shows that R-12 yields the highest value of COPH, followed by R-22 and R-134A. The system performance degradation is about 2–4% in the 0–20 °C temperature range, when R-12 is replaced by R-134A. For the mixture refrigerants, R-410A is
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Table 1 Refrigerants and chemical composition ASHRAE designation
Components
Composition
R-12 R-22 R-134a R-404A R-407C R-410A
Dichlorodifluoromethane Monochlorodifluoromethane 1,1,1,2-tetrafluoroethane R125/143A/134A R32/125/134A R32/125
Pure Pure Pure 44/52/4 23/25/52 50/50
Fig. 3. Comparison of COPH for different refrigerants at 60 °C.
shown to be more efficient than either R-407C or R-404A, but it has a performance level about 15– 20% lower than that obtained with R-134A. This study also addresses the sizing or design problem by proposing a simple graphical procedure to determine the required collector area of the direct expansion solar assisted heat pump system for proper matching to the compressor heat pumping capacity. To accomplish this, the collector heat collection capacity, the right hand side of Eq. (4), Kgcoll, and the heat pump evaporative capacity, the left hand side of Eq. (4), are plotted as functions of solar radiation, DT, VD/Ac and Tf. Using REFPROP, the refrigerant property data base program, performance charts have been developed for six refrigerants namely, R-12, R-22, R-134A, R-407C, R-410A and R-407C, two different collector designs (namely a bare and a one cover collector), six VD/Ac ratios, eight solar radiation values and five condensing temperatures. A comprehensive listing of the charts is given in Ref. [22]. Fig. 4 shows the variation of (gcollS)/Sc as a function of collector to ambient temperature difference (DT = Tf Ta). It also shows the variation of COPH and the heat pump evaporative
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Fig. 4. A chart for sizing of collector and compressor for one glass cover collector and refrigerant R-134A.
capacity as a function of the collector temperature Tf. The reason for using two variables along the x-axis is that the collector efficiency is characterized by the collector heat loss, which scales with (Tf Ta), whereas the COPH and the heat pump evaporative capacity are a function of Tf. The chart in Fig. 4 can be used in the so-called design problem, when the system is initially being designed for a design point operation. The location dependent parameters S/Sc and Ta, as well as the operational parameters, namely load temperature and thermal energy delivery rate (QH) are specified, and it is required to determine the collector area (Ac) and the
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Table 2 Collector parameters Collector type
F0
(sa)
UL Watts/m2 °C
One cover collector
0.85
0.85
6.0
compressor size (compressor displacement volume rate, VD). In addition, it is also specified that the system operates with a given DT at the design point condition. This design procedure is illustrated by choosing an example in which refrigerant R-134a is employed for a one glass cover collector. Similar procedures can be adopted for other refrigerants and/or collector configurations and are described in Ref. [22] in more details. The collector parameters are given in Table 2. For illustration of the procedure for the design problem, Ta = 5 °C, S = 500 W/m2, Tcond. = 60 °C and DT = 5 °C are chosen. This translates into Tf being equal to 10 °C. With this value, using Fig. 4, one can draw the line DF. Using DT (5 °C in the present example), one draws the vertical line that intersects the S = 500 W/m2 line at point A. The horizontal line BA is drawn parallel to the x-axis to intersect the line DF at point C that, in this example, lies on the VD/ Ac = 0.685 m/h line for the compressor evaporative capacity. This fixes the VD/Ac ratio for the specified design parameters. The collector area Ac is determined from the thermal energy delivery capacity. It is noted that for Tf = 10 °C, the COPH is 3.8 (point G on the 60 °C condensing temperature line). The solar energy collected (QL) for a given thermal load (QH = 7000 W in this example) can be expressed by the following equation: COPH 1 7000 2:8 ¼ 5158 W QL ¼ QH ¼ COPH 3:8 With (gcollS)/Sc = 0.245 (point A), one can determine gcollS = 331.5 W/m2 (since Sc = 1353 W/m2). Finally, Ac can be calculated from the equation Ac ¼ QL =ðgcoll SÞ ¼ 15:6 m2 From VD/Ac = 0.685 m/h, one can determine VD ¼ 0:685 15:6 m3 =h ¼ 10:66 m3 =h Using a compressor motor RPM of 1750 as an example, one can determine the piston displacement volume (PD) of the compressor from the relationship, PD ¼
VD 10:66 ¼ ¼ 0:0000711 m3 =rev: 60N 1750 60
Calculations similar to those performed for a one cover flat plate collector were also conducted for a bare collector for the same design point conditions described earlier. The procedure illustrated in Fig. 5 is similar to the one followed in Fig. 4 for a one cover collector and resulted in a value of VD equal to 10.5 m3/h and a value of 17.2 m2 for the collector area.
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Fig. 5. A chart for sizing and compressor for bare collector and refrigerant R-134 A.
4. Conclusions Thermodynamic analysis of a direct expansion solar assisted heat pump is performed for a variety of refrigerants. The performance degradation due to switching from R-12, a chloro-fluorocarbon refrigerant, to hydrofluorocarbon refrigerants is investigated. The results show that R-12 produces the highest value of COPH, followed by R-22 and R-134A. The system performance degradation is about 2–4% in the 0–20 °C collector temperature range when R-12 is replaced with
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R-134A. For the mixture refrigerants, R-410A is shown to be more efficient than either R-407C or R-404A but not as good as R-134A. The refrigerant R-410A produces COPH values that are 15–20% lower than those obtained with R-134A. A simple graphical procedure is proposed to design a collector area for the direct expansion solar assisted heat pump system that is properly matched to the compressor heat pumping capacity. This graphical procedure is summarized using performance charts for a one glass cover collector using R-134A as the refrigerant. These charts are very convenient for designing the DX-SAHP system and provide, as illustrated in this study, a quick estimate of the collector area and compressor displacement for specified solar radiation, ambient temperature, load temperature and thermal load.
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