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R32 blends in a water-to-water heat pump system
Applied Thermal Engineering 162 (2019) 114303
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Experimental study on CO2/R32 blends in a water-to-water heat pump system
T
⁎
Zhili Sun , Qi Cui, Qifan Wang, Jinghong Ning, Jianghe Guo, Baomin Dai, Yongqiang Liu, Yibo Xu Tianjin Key Laboratory of Refrigeration Technology, Tianjin University of Commerce, Tianjin 300134, China
H I GH L IG H T S
experiment of CO /R32 blends in a water-to-water heat pump system is conducted. • The and 65.2% maximum efficiency improvement in heating and cooling are obtained. • 23.3% optimal mixing ratio of CO /R32 (0.6/0.4) is found. • The pressure, cooling capacity, and heating capacity are reduced. • System • The internal heat exchanger further improves the system performance. 2
2
A R T I C LE I N FO
A B S T R A C T
Keywords: Water-to-water heat pump system CO2/R32 blends Thermal matching Internal heat exchanger
The purpose of this study is to evaluate the cooling and heating performance as well as the compressor operating conditions of CO2/R32 blends in a water-to-water heat pump system. Experiments were conducted to reveal the effects of the refrigerant compositions, heat source temperatures, and internal heat exchanger (IHX) on the performance of the system. The results show that 23.3% and 65.2% maximum improvement in heating coefficient of performance (COPh) and cooling coefficient of performance (COPc) of the system were obtained when the mass fraction of CO2 was 0.6, respectively. Meanwhile, heating capacity, cooling capacity, and the compressor discharge pressure of CO2/R32 blends were decreased significantly. The thermal matching between the heat source and CO2/R32 blends significantly affected the overall performance of the system. In addition, IHX is recommended in the water-to-water heat pump to improve the system performance.
1. Introduction CO2 is considered to be the most promising refrigerant [1] due to its excellent properties: 0 ozone depletion potential (ODP), low global warming potential (GWP), non-toxic, non-flammable, low viscosity, and high volumetric capacity [2], which promotes the applications of CO2 in heat pump water heater [3], commercial refrigeration and heat pump systems [4,5], automobile air conditioning [6], freezing and cold storage [7]. However, the large-scale applications of CO2 are limited by the two inevitable issues [8]: (1) high initial costs and the safety issues due to the significant high operating pressure of CO2. (2) low energy efficiency as a result of the large throttling losses during the expansion process. Many solutions are used for addressing these two issues, such as improving the expansion process by using expander or ejector instead
⁎
of throttle expansion valve [9], improving the compression process by adopting multi-stage compression intermediate cooling technology [10], and improving the system performance by mechanical subcooling [11]. While the system performance is improved and the system operating pressure is reduced, these solutions inevitably have a limitation, the assisted equipment result in additional costs. In view of this situation, numerous studies had been conducted in CO2 blends by some researchers. In the field of heating, ventilation, air-conditioning (HVAC), and heat pump, the researches of CO2 blends are mainly focused on CO2/ propane. Yu et al. [12] conducted a theoretical analysis and experimental study on the CO2/propane blends, the results show that the system coefficient of performance (COP) of the CO2/propane blends is 29.4% higher than that of the pure CO2, and the high pressure is reduced during the research range. Zhang et al. [13] carried out an
Corresponding author at: Tianjin Key Laboratory of Refrigeration Technology, Tianjin University of Commerce, Tianjin 300134, China. E-mail address: [email protected] (Z. Sun).
https://doi.org/10.1016/j.applthermaleng.2019.114303 Received 30 January 2019; Received in revised form 17 August 2019; Accepted 22 August 2019 Available online 23 August 2019 1359-4311/ © 2019 Elsevier Ltd. All rights reserved.
Applied Thermal Engineering 162 (2019) 114303
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Nomenclature Q W Cp m T/t s
e,wout g,w g,win g,wout
cooling capacity or heating capacity (kW) the compressor input power (kW) specific heat (kJ kg−1 K−1) mass flow rate (kg s−1) temperature (°C) specific entropy (kJ kg−1 K−1)
Acronyms COP IHX ODP GWP HVAC
Subscripts c h e g e,w e,win
the water at the evaporator outlet hot water in the gas cooler/condenser the water at the gas cooler/condenser inlet the water at the gas cooler/condenser outlet
cooling heating evaporator gas cooler/condenser chilling water in the evaporator the water at the evaporator inlet
coefficient of performance internal heat exchanger ozone depletion potential global warming potential heating, ventilation, and air-conditioning
Greek letters δ
uncertainty
district heating in cold region. The results show that as the refrigerant mass fraction of R32 increases, the system COP increases. The above studies about CO2-based and R32-based blends indicated that they have excellent properties and application prospects. However, when it comes to CO2/R32 blends, there are only a few studies, of which Xia et al. [24] carried out the thermo-economic analysis and comparative study of trans-critical power cycles using CO2/R32 blends as working fluids. The results show that among the CO2-based blends and pure CO2, CO2/R32 presents the highest exergy efficiency of 52.85%. Di Nicola et al. [25] analyzed the performances of a cascade refrigeration cycle operated with CO2/R32 blends as the low-temperature working fluid and concluded that CO2/R32 blends can be considered an attractive option for the low-temperature-circuit in cascade systems. Dai et al. [18] theoretically concluded that CO2/R32 is more suitable for its fairly high COP and lower high-side pressure in the case of moderate to high external temperature glide. Hakkaki-Fard et al. [26] developed a detailed screening heat pump theoretical model and assessed the performance of CO2/R32 blends. The results show that the blends of CO2/R32 (20/80 mass fraction) have the best performance among the mixtures studied. However, these results of the above studies about CO2/R32 blends have not been verified by experiments, a gap remains between the theoretical analyses and experimental study. Therefore, it is critical for experiments to be conducted and the cooling and heating performance as well as the compressor operating conditions of CO2/R32 blends to be evaluated. This study aims to contribute to filling the knowledge gap about the experimental study on CO2/R32 blends in the heat pump system. This paper is the first step of this research project, in which experiments are conducted to reveal the effects of the refrigerant compositions, heat source temperatures, and internal heat exchanger (IHX) on the performance of the system. The research on the optimal amounts of CO2/R32 blends charge, the heat matching characteristics of refrigerants and high-temperature and low-temperature heat sources, the blends composition shift, leakage and safety analyses will be conducted later by our research group.
experimental and theoretical investigation on the performance of CO2/ propane auto-cascade refrigerator and concluded that the cycle performance will be improved by increasing CO2 mass fraction. Zhu et al. [14] experimentally investigated the flow evaporation heat transfer characteristics of CO2/propane blends. The results show that the heat transfer coefficient of blends is between that of pure CO2 and pure propane for the same conditions, and increases with increasing CO2 proportion. Zhang et al. [15] carried out simulation and experimental investigations to research the relationships between optimum heat rejection pressure and related operating parameters for a trans-critical system using CO2/propane mixture. The results show that for a transcritical cycle working with CO2/propane mixture in which the mass fraction of CO2 is greater than 0.78, there is an optimal heat rejection pressure, under which a maximum system COP can be reached. There are also some studies on CO2/R41 blends. Yu et al. [16] carried out an experimental energetic analysis of CO2/R41 blends in automobile air conditioning and heat pump systems. The results demonstrate that the system COP in the heating and cooling modes can be improved up to a maximum of 14.5% and 25.7%, respectively. Wang et al. [17] conducted a thermodynamic analysis of CO2 blends with R41 applied in a small refrigerated cabinet and heat pump water heater. The results show that the CO2/R41 blends reduce the optimal high pressure of pure CO2 system, improve the system performance. Dai et al. [18] conducted a thermodynamic analysis of CO2/R41 blends in heat pump water heater, the results show that CO2/R41 is a suitable candidate for heat pump water heaters because of its high COP and low high-side pressure. Propane and R41 are environmentally friendly refrigerants with 0 ODP and low GWP. However, R32, which is also an environmental refrigerant especially in the transition stage of refrigerant substitution, is seldom paid attention to CO2 blends. Moreover, R32 is superior to propane in flammability and applicability. A great number of R32 blends have been recently proposed to replace R22 and R410A [19]. He et al. [20] carried out an experimental investigation on flow boiling heat transfer performance of propane/R32 in horizontal tubes and concluded that the new refrigerant mixture is near azeotropic, 0 ODP, and low GWP, which will be a competitive R22 replacement in household air conditioners. Tian et al. [21] conducted an experimental investigation of refrigerant mixture R32/propane as a drop-in replacement for 410A in household air conditioners. The results show that cooling and heating capacities are increased by 14.0–23.7%. Wang et al. [22] carried out a numerical study of gas injected heat pump using zeotropic R32/R1234ze(E) blends, and concluded that R32/R1234ze(E) zeotropic mixture is a potential replacement in the field of air conditioning and heat pumps. Fan et al. [23] conducted a theoretical study on a modified heat pump cycle with zeotropic mixture R32/propane for
2. Experiments 2.1. Experiment setup Fig. 1 illustrates a schematic diagram of the experiment setup of CO2/R32 blends in a water-to-water heat pump system. Table. 1 presents the main equipment parameters and accuracies of the devices. This system was composed of a compressor, oil separator, gas cooler/ condenser, IHX, hand expansion valve, evaporator, gas-liquid separator, hot water tank, chilling water tank, two water pumps, 2
Applied Thermal Engineering 162 (2019) 114303
Z. Sun, et al. TP
TP
1
2
Table 2 Test conditions. T
7
4 T
3
TP
TP
TP
TP
6
M
5 T
T
T
8
M
11
T
9
M
10
NO.
Mass fraction of CO2
Refrigerant charge (kg)
tg,win (°C)
tg,wout (°C)
te,win (°C)
te,wout (°C)
1 2 3 4 5 6 7
100% 95% 90% 80% 70% 60% 50%
2.0
20 20 40
45 55 45
12
7
evaporator outlet could be superheated, which leads to the increase in cooling capacity and heating capacity. The variation trend of cooling coefficient of performance (COPc) and heating coefficient of performance (COPh) are determined by the increasing degree of the compressor input power, cooling capacity and heating capacity, which are closely related to the thermophysical properties of refrigerants. When the increasing degree of cooling capacity and heating capacity are greater than the increasing degree of the compressor input power, COPc and COPh will increase, vice versa. The test methods are as follows. Firstly, the experiments of each group without IHX were conducted by opening the bypass valve ① and closing the bypass valve ②. The opening degree of the hand expansion valve was controlled to maintain the temperatures of CO2/R32 blends after throttling at 0 °C. Water temperature in the hot water tank was maintained at 20 °C by means of the thermostatic control system, and it was heated to 45 °C in the gas cooler/condenser by changing its volume flow rate, which achieving the working condition that the water temperature at the gas cooler/condenser inlet was 20 °C and the water temperature outlet was 45 °C. Similarly, the water temperatures at the evaporator inlet and outlet were maintained at 12 °C/7 °C. The experimental data were recorded after 30 min of stable operation of the system. Thereafter, the water temperatures at the gas cooler/condenser inlet and outlet were changed to 20 °C/55 °C and 40 °C/45 °C, respectively and the experimental data were recorded, respectively. Finally, the bypass valve ① was closed and the bypass valve ② was opened to carry out the experiments with IHX and then the above experimental steps without IHX were repeated to complete experiments of this group.
12
1-Compressor; 2-Oil separator; 3-Gas cooler/Condenser; 4- IHX; 5-Hand expansion valve; 6Evaporator; 7-Gas-liquid separator; 8,9-Pump; 10-Thermostatic control system; 11-Hot water tank; 12-Chilling water tank; T-Temperature sensor; P-Pressure sensor; M-water volume flow meter or refrigerant mass flow meter
Fig. 1. Schematic diagram of the experiment setup of CO2/R32 blends in a water-to-water heat pump system.
thermostatic control system (a refrigeration unit and electric heater), and several bypass valves. 2.2. Test conditions and test method The experiments were conducted under various test conditions as presented in Table. 2. Before experiments of each group were carried out, all refrigerants in the system must be extracted first, and then the system should be vacuum treated. Because the charging pressure of CO2 is higher than that of R32, R32 was charged first, followed by CO2. It is informed here that the amounts of CO2/R32 blends charge were all set to 2.0 kg [27]. And the experimental study on the optimal amounts of CO2/R32 blends charge will be conducted later by our research group. IHX was used in the system to explore the possibility of improving system performance. Theoretically, the compressor suction temperature can be increased by the use of IHX, resulting in an increase in the compressor input power based on the constant compressor pressure ratio. However, at the same time, the refrigerant at the gas cooler/ condenser outlet could be subcooled and the refrigerant at the
2.3. Data verification All Pt100 sensors were calibrated before the experiment. First, they were bundled with a four-wire Pt25 standard platinum resistor and placed in a thermostatic liquid tank. The temperature measurement
Table 1 The main equipment parameters and accuracies of the devices. NO.
Equipment
Main parameters
Model
Accuracy
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
Compressor Power-meter Compressor oil Gas cooler/condenser Hand expansion valve Evaporator IHX Hot water tank Cool water tank Temperature sensor Pressure sensor Water volume flow meter Refrigerant mass flow meter Water pump Oil separator Gas-liquid separator Electric heater
50 Hz, 1450 rad min−1, Discharge volume 1.12 m3 h−1 Rated voltage 3X220/380 V Density 0.974 g cm−3 (20 °C), Open flash point 242 °C Heat transfer area 1.8 m2, Design pressure 16 MPa Specification for circulation area φ12 Heat transfer area 1.5 m2, Design pressure 10 MPa Heat transfer area 0.2 m2, Design pressure 16 MPa Tank volume 0.225 m3 Tank volume 0.225 m3 Measuring range of −50–450 °C Measuring range of 0–16 MPa Measuring range of 0.6–6 m3 h−1 Measuring range of 0–100 kg h−1 Max flow 2 m3 h−1, Head 15 m, Power 0.37 kW, 2900 r min−1 Diameter 11 cm, height 52 cm Diameter 11 cm, height 52 cm Input power 200 W
CD180H 7p-0.5 POE 130 Double-pipe
/ 0.5% FS / / / / / / / ± 0.1 °C 0.5% FS 0.5% FS 0.2% FS / / / /
3
Double-pipe Double-pipe
PT100 DTT0X LWGYC-15 CMFI CHL2-20 LSWSC
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of CO2. This is because the addition of R32 reduces the blends′ density, thus reducing the blends’ mass flow rate as shown in Fig. 5. That eventually reduces the heating capacity. It can be observed that when the mass fraction of CO2 decreased from 0.95 to 0.9, both COPh and heating capacity decreased significantly, which indicates that the mass fraction of CO2 at 0.9 is a significant point, where the heat transfer performance between the blends and the heat source of the system is poor. Fig. 2 shows that when the mass fraction of CO2 was 0.9, the cycle form of the system was the trans-critical cycle, where the heat transfer of the blends was in the pseudo-critical area (above the critical point). In the pseudo-critical area, the heat transfer particularity of supercritical fluid is mainly manifested in the fact that the specific heat of supercritical fluid is sometimes higher or lower than that of conventional single-phase forced convection, and reaching the maximum at the pseudo-critical point [29]. Besides, the maximum specific heat increases as the operating pressure decreases, which results in the maximum average specific heat in the heat transfer process of the gas cooler, when the mass fraction of CO2 is 0.9. Limited by the small heat transfer area of the gas cooler, the temperature of CO2/R32 (0.9/0.1) changed slightly, which accounts for the temperature at the gas cooler outlet was larger than other blends. The experimental data show that the temperature at the gas cooler outlet of CO2/R32 (0.9/0.1) was 3.8 °C higher on average than other blends for three working conditions. The higher the outlet temperature of the gas cooler, the greater the end temperature difference and an increase in entropy generation of the gas cooler, eventually leads to poor heat transfer performance. This situation can be improved by increasing the heat transfer area of the gas cooler. For most working conditions, when the mass fraction of CO2 was 0.6, the COPh of the system reached its maximum value. Compared with pure CO2, the COPh increased by 10.3%, 19.2%, and 12.1%, respectively, for three working conditions. Under the same baseline, this value would change to 16.1%, 23.3%, and 12.5%, respectively after using IHX, which indicates that the COPh can be further increased by the use of IHX as shown in Fig. 3. This is because, for the CO2/R32 blends in the subcritical cycle, the heating capacity is increased while the compressor input power is decreased (See Fig. 11) after using IHX, which inevitably leads to an increase in COPh. However, when the mass fraction of CO2 was 0.6, compared with pure CO2, the heating capacity decreased by 15.9%, 10.4%, and 14.7%, respectively, for three working conditions. Thus, the balance between heating capacity and COPh at three working conditions should be weighed deliberately. Although when the mass fraction of CO2 was greater than 0.8, the use of IHX reduced the mass flow rate of the blends presented in Fig. 5, but it still increased the heating capacity of all blends as a result of the superheating degree at the evaporator outlet increases. In summary, IHX is recommended in the water-to-water heat pump to improve the heating performance of
accuracy of the four-wire Pt25 standard platinum resistor is very high, and the measured values can be accurate to 5 mK that is acceptable for this experiment. The measured values with high accuracy of the fourwire Pt25 standard platinum resistor were read out by the multimeter as the “true value” of temperature. Then, the temperature of the liquid in the thermostatic tank was changed per 5 °C in the temperature range of −5 to 150 °C, and the measured values of all Pt100 sensors and the four-wire Pt25 standard platinum resistor were recorded. Finally, the linear relationship between measured values by all Pt100 sensors and measured values by the four-wire Pt25 standard platinum resistor were fitted by linear regression, and each Pt100 sensor was calibrated. Consequently, the temperatures measured by the calibrated Pt100 sensors are convincing. The sensor locations were presented in Fig. 1. The accuracies of the measuring devices were shown in Table. 1. The cooling and heating capacities of the water-to-water heat pump system were determined by the water-side heat transfer, using Eqs. (1) and (2). The errors between the right side and the left side of Eq. (3) were within ± 10%, indicated that the tested system exhibits an effective heat balance between the high-temperature heat source and low-temperature heat source sides. COPc and COPh were determined by Eqs. (4) and (5). Uncertainty analysis was performed to verify the cooling/heating capacity and COP, the indirect measurements, using Eq. (6) [28], where R represents the targeted capacity or COP, Xi represents the uncertainty of its affecting factors, and N is the number of affecting factors. Based on that, the relative uncertainties of the system cooling/heating capacity and COP were calculated as 0.6–2.0% and 0.8–2.1%, respectively.
Qc = Cp me, w (te, win − te, wout )
(1)
Qh = Cp mg, w (tg, wout − tg, win )
(2)
Qh = Qc + W
(3)
COPc = Qc / W
(4)
COPh = Qh/ W
(5)
δR = R
N
2
δXi ⎞ ⎝ Xi ⎠
∑⎛
⎜
i=1
⎟
(6)
3. Results and discussion 3.1. Effect of CO2/R32 blends on the system performance The critical temperatures of CO2 and R32 are 31.1 °C and 78.1 °C, respectively. When the temperature of high-temperature heat source reaches a certain high value, the cooling process without phase change of CO2 is completed above the critical point, where the cycle form of the system is called the trans-critical cycle and fluids above the critical point are called supercritical fluids. As R32 with higher critical temperature blended with CO2, the critical temperature of the CO2 blends increased. That indicates that there is a mixing ratio where the CO2 blends condensate below the critical point, the cycle form of the system is called the subcritical cycle. Fig. 2 presents the T-s diagrams of the system at mass fraction of CO2 with 0.8 and 0.9, under the experimental conditions. When the mass fraction of CO2 was 0.9, the cycle form of the system was the trans-critical cycle, and the high-temperature heat source heat exchanger is called gas cooler. However, when the mass fraction of CO2 was 0.8, the cycle form of the system was the subcritical cycle, and the high-temperature heat source heat exchanger is called condenser. Figs. 3 and 4 illustrate the variation of COPh and heating capacity with mass fraction of CO2. As can be seen from Fig. 3, the addition of R32 increased the COPh of pure CO2, which means R32 is the refrigerant having a higher heating effect than CO2. However, Fig. 4 presents the heating capacity decreased with reducing the mass fraction
380 360
Mass fraction of CO2 at 0.8 Mass fraction of CO2 at 0.9
T (K)
340 320 300 280 260 240 0.8
1.0
1.2
1.4
1.6
1.8
2.0
2.2
s (kJ/(kg·K)) Fig. 2. T-s diagram of the system at mass fraction of CO2 with 0.8 and 0.9. 4
Applied Thermal Engineering 162 (2019) 114303
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6.0 5.5 5.0
COPh
4.5 4.0
among the three working conditions. The experimental data show that the temperature glides of refrigerant in the gas cooler/condenser increased in turn under the working conditions where the water temperatures at the gas cooler/condenser inlet and outlet were 20 °C/45 °C, 20 °C/55 °C, and 40 °C/45 °C, respectively (whether IHX was used or not). For the three working conditions, the average temperature glides of refrigerant were 66.9 °C, 80.1 °C, and 91.9 °C, respectively, and this value would change to 74.3 °C, 86.2 °C, and 93.9 °C, respectively after using IHX. When the water temperatures at the gas cooler/condenser inlet and outlet were 20 °C/45 °C, the temperature glides of refrigerant in the gas cooler/condenser were the smallest, and that leads to the better thermal matching with the heat source, which results in the largest COPh and heating capacity among the three working conditions. Compared with the other two working conditions, the COPh increased by 10.9% and 49.5%, respectively and the heating capacity increased by 2.0% and 32.3%, respectively, when the mass fraction of CO2 was 0.6. When the water temperatures at the gas cooler/condenser inlet and outlet were 40 °C/45 °C, the COPh and heating capacity of the system were the worst, due to the slight temperature changes of hot water, which results in a mismatch with the larger temperature variations of the blends. As can been observed form Figs. 6 and 7, the effects of the refrigerant compositions, heat source temperatures, and IHX on the COPc and cooling capacity were basically the same as similar to COPh and heating capacity. The addition of R32 increased the COPc of pure CO2, which means R32 is also the refrigerant having a higher cooling effect than CO2. The cooling capacity decreased with reducing the mass fraction of CO2, which as a result of the reduction in the mass flow rate. For most working conditions, when the mass fraction of CO2 was 0.6, the COPc of the system also reached its maximum value. Compared with pure CO2, the COPc increased by 22.6%, 20.1%, and 31.2%, respectively, for three working conditions. Under the same baseline, this value would change to 29.3%, 29.8%, and 65.2%, respectively after using IHX, which indicates that the COPc can be further increased by the use of IHX because the cooling capacity is increased while the compressor input power is decreased. And the influence of IHX on cooling performance is greater than that of heating performance. Moreover, when the mass fraction of CO2 was 0.6, compared with pure CO2, the cooling capacity decreased by 6.5%, 9.7%, and 0.1%, respectively, for three working conditions. Similarly, the balance between cooling capacity and COPc at three working conditions should be weighed deliberately. IHX also increased the cooling capacity of all blends as a result of the subcooling degree at the gas cooler/condenser outlet increases. To sum up, IHX is recommended in the water-to-water heat pump to improve the cooling performance of the system. Obviously, the thermal matching between the high-temperature heat source and blends affected the heat transfer between the lowtemperature heat source and blends. When the water temperatures at
inlet 20 oC/outlet 45 oC inlet 20 oC/outlet 55 oC inlet 40 oC/outlet 45 oC inlet 20 oC/outlet 45 oC with IHX inlet 20 oC/outlet 55 oC with IHX inlet 40 oC/outlet 45 oC with IHX
3.5 3.0 2.5 2.0 1.5 0.4
0.5
0.6 0.7 0.8 0.9 Mass fraction of CO 2
1.0
1.1
Fig. 3. COPh variation with mass fraction of CO2.
5.0
Heating capacity (kW)
4.5 4.0
inlet 20 oC/outlet 45 oC with IHX inlet 20 oC/outlet 55 oC with IHX inlet 40 oC/outlet 45 oC with IHX
3.5 3.0 2.5
inlet 20 oC/outlet 45 oC inlet 20 oC/outlet 55 oC inlet 40 oC/outlet 45 oC
2.0 1.5 0.4
0.5
0.6 0.7 0.8 0.9 Mass fraction of CO 2
1.0
1.1
Fig. 4. Heating capacity variation with mass fraction of CO2.
24
inlet 20oC/outlet 45oC inlet 20oC/outlet 55oC inlet 40oC/outlet 45oC inlet 20oC/outlet 45oC with IHX inlet 20oC/outlet 55oC with IHX inlet 40oC/outlet 45oC with IHX
Mass flow rate (10-3 kg/s)
22 20 18 16 14 12 10 8 6
0.4
0.5
0.6 0.7 0.8 0.9 Mass fraction of CO2
1.0
4.0
1.1
inlet 20 oC/outlet 45 oC with IHX inlet 20 oC/outlet 55 oC with IHX inlet 40 oC/outlet 45 oC with IHX
3.6 3.2
Fig. 5. Mass flow rate of refrigerant variation with mass fraction of CO2.
2.8 COPc
the system. The experiments were carried out at three different temperatures of high-temperature heat sources. The CO2/R32 blends are non-azeotropic, which indicates that the temperature glide occurs during its phase change. If the temperature of the heat source and the blends reach the thermal matching, that is, the gradients of temperature change are similar, and the system performance will be greatly improved. Obviously, as can be seen from Figs. 3 and 4, when the water temperatures at the gas cooler/condenser inlet and outlet were 20 °C/ 45 °C, the COPh and heating capacity of the system were the largest
2.4 2.0 1.6 inlet 20 oC/outlet 45 oC inlet 20 oC/outlet 55 oC inlet 40 oC/outlet 45 oC
1.2 0.8 0.4
0.4
0.5
0.6 0.7 0.8 0.9 Mass fraction of CO 2
1.0
Fig. 6. COPc variation with mass fraction of CO2. 5
1.1
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Z. Sun, et al.
5.0 Cooling capacity (kW)
4.5 4.0 3.5 3.0
at the high-temperature side is cooled more fully and the high pressure is reduced. As mentioned above, when the water temperatures at the gas cooler/condenser inlet and outlet are 20 °C/45 °C, the system has the best heating and cooling performance among the three working conditions due to the better thermal matching, which results in the lowest compressor discharge pressure. Compared with the other two working conditions, the compressor discharge pressure decreased by 10.6% and 19.2%, respectively, when the mass fraction of CO2 was 0.6. Fig. 9 indicates that for most working conditions, blended R32 with CO2 reduced the compressor discharge temperature, and it further reduced with the increase of R32 composition, but the reduction was only within 20 °C. For most working conditions, the use of IHX increased the compressor discharge temperature, which is the inevitable result of the increase in the compressor suction temperature. The increasing range of the compressor discharge temperature is acceptable. When the water temperatures at the gas cooler/condenser inlet and outlet were 20 °C/ 45 °C, the compressor discharge temperature was the lowest, as a result of the compressor pressure ratio is the smallest under this working condition, as shown in Fig. 10. Compared with the other two working conditions, the compressor discharge temperature decreased by 10.9% and 28.3%, respectively, when the mass fraction of CO2 was 0.6. Figs. 10 and 11 show the variation of the compressor pressure ratio and the compressor input power with mass fraction of CO2. The compressor pressure ratio is the ultimate parameter reflecting the compressor operating conditions. Fig. 10 illustrates that for most working conditions, the compressor pressure ratio had an increasing trend with the increase of R32 composition, but the amplitude was too small except for the third working condition, which indicated that for any CO2/ R32 blends, the compressor operating conditions are stable. For most working conditions, the use of IHX had little influence on the compressor pressure ratio, especially in the trans-critical cycle. When the water temperatures at the gas cooler/condenser inlet and outlet were 20 °C/45 °C, the compressor pressure ratio was the smallest, which indicated that under this working condition, the compressor operating conditions are the best. Compared with the other two working conditions, the compressor pressure ratio decreased by 10.9% and 24.3%, respectively, when the mass fraction of CO2 was 0.6. Fig. 11 presents that blended R32 with CO2 reduced the compressor input power, and it reduced further with the increase of R32 composition, as a result of the reduction of mass flow rate indicated in Fig. 5. It can be observed that when mass fraction of CO2 decreased from 0.95 to 0.8, the compressor input power decreased slightly. The cause of this phenomenon is the same as similar to the compressor discharge pressure. When the mass fraction of CO2 was 0.6, compared with pure CO2, the compressor input power decreased by 23.7%, 24.8%, and 23.9%, respectively, for three working conditions. Under the same baseline, this value would change to 24.6%, 26.0%, and 24.4%, respectively after
inlet 20 oC/outlet 45 oC inlet 20 oC/outlet 55 oC inlet 40 oC/outlet 45 oC inlet 20 oC/outlet 45 oC with IHX inlet 20 oC/outlet 55 oC with IHX inlet 40 oC/outlet 45 oC with IHX
2.5 2.0 1.5 1.0 0.4
0.5
0.6 0.7 0.8 0.9 Mass fraction of CO 2
1.0
1.1
Fig. 7. Cooling capacity variation with mass fraction of CO2.
the gas cooler/condenser inlet and outlet were 20 °C/45 °C, the COPc and cooling capacity of the system were the largest among the three working conditions. Compared with the other two working conditions, the COPc increased by 20.5% and 88.1%, respectively and the cooling capacity increased by 10.9% and 66.5%, respectively, when the mass fraction of CO2 was 0.6. Essentially, the thermal matching between blends and chilling water is the key to affect the cooling performance of the system. However, this experiment only focused on the thermal matching between high-temperature heat source and blends, and there was only one group of the low-temperature heat source. The temperatures of CO2/R32 blends after throttling were maintained at 0 °C, and the experimental data show that the temperatures of CO2/R32 blends of the evaporator outlet were all close to 12 °C, the inlet chilling water temperature. Although the temperature glide of blends in the evaporator increases with the decrease of evaporating pressure. However, it can be found that the heat transfer between all blends and chilling water were still mismatched. In view of this, Kim et al. [30] concluded that this mismatch of temperature profiles may lead to pinching at the evaporator exit, and it was difficult to decrease the mean temperature difference between two heat transfer fluids by changing the heat exchanger area. Therefore, to improve the system efficiency, CO2/R32 blends must necessarily be used in applications with appropriate large temperature glides of the low-temperature heat source. The above results lay a foundation for the thermal matching experiment between the low-temperature heat source and blends conducted later by our research group.
Discharge pressure of compressor (MPa)
3.2. Effect of CO2/R32 blends on the compressor operating conditions Figs. 8 and 9 illustrate the variation of the compressor discharge pressure and the compressor discharge temperature with mass fraction of CO2. As shown in Fig. 8, blended R32 with CO2 reduced the compressor discharge pressure, and it reduced further with the increase of R32 composition, as a result of the normal boiling point of CO2 (−78.4 °C) is lower than that of R32 (−51.7 °C). It can be observed that when the mass fraction of CO2 decreased from 0.95 to 0.8, the compressor discharge pressure decreased slightly. As mentioned above, the heat transfer performance between the blends and the heat source of the system is poor when the mass fraction of CO2 is 0.9, which results in relatively high compressor discharge pressure. When the mass fraction of CO2 was 0.6, compared with pure CO2, the compressor discharge pressure decreased by 26.2%, 26.9%, and 21.4%, respectively, for three working conditions. Under the same baseline, this value would change to 27.5%, 28.4%, and 23.9%, respectively after using IHX, which indicates that IHX can further reduce the compressor discharge pressure in the subcritical cycle. This is because after using IHX, the refrigerant
11 10
inlet 20oC/outlet 45 oC with IHX inlet 20oC/outlet 55 oC with IHX inlet 40oC/outlet 45 oC with IHX
9 8 7 6
inlet 20oC/outlet 45 oC inlet 20oC/outlet 55 oC inlet 40oC/outlet 45 oC
5 4 0.4
0.5
0.6 0.7 0.8 0.9 Mass fraction of CO 2
1.0
1.1
Fig. 8. Compressor discharge pressure variation with mass fraction of CO2. 6
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Discharge temperature of compressor (oC)
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200
input power was the smallest, as a result of the compressor pressure ratio was the smallest. Compared with the other two working conditions, the compressor input power decreased by 8.0% and 11.5%, respectively, when the mass fraction of CO2 was 0.6.
inlet 20oC/outlet 45oC inlet 20oC/outlet 55oC inlet 40oC/outlet 45oC inlet 20oC/outlet 45oC with IHX inlet 20oC/outlet 55oC with IHX inlet 40oC/outlet 45oC with IHX
180 160 140
4. Conclusions Experiments were conducted to reveal the effects of the refrigerant compositions, heat source temperatures, and internal heat exchanger (IHX) on the cooling and heating performance as well as the compressor operating conditions of CO2/R32 blends in a water-to-water heat pump system. Based on the results and preceding discussion, the following conclusions can be obtained.
120 100 80 0.4
0.5
0.6 0.7 0.8 0.9 Mass fraction of CO2
1.0
1.1
(1). R32 is the refrigerant having a higher heating and cooling effect than CO2, 23.3% and 65.2% maximum improvement in heating coefficient of performance (COPh) and cooling coefficient of performance (COPc) were obtained when the mass fraction of CO2 was 0.6, respectively. Meanwhile, heating capacity and cooling capacity of CO2/R32 blends were decreased, as a result of the mass flow rate decreased due to the smaller vapor density of R32. R32 blended with CO2, the compressor discharge pressure was reduced significantly, and the compressor operating conditions were stable due to the compressor pressure ratio changed slightly. (2). The high-temperature heat source and the CO2/R32 blends reached the better thermal matching when the water temperatures at the gas cooler/condenser inlet and outlet were 20 °C/45 °C, which results in the optimum overall performance of the system. The lowtemperature heat source and the CO2/R32 blends all mismatched when the water temperatures at the evaporator inlet and outlet were 12 °C/7 °C, and thus CO2/R32 blends must necessarily be used in applications with appropriate large temperature glides of the low-temperature heat source. (3). IHX is recommended in the water-to-water heat pump to improve the heating and cooling performance of the system. IHX further reduced the compressor discharge pressure in the subcritical cycle. Due to the compressor pressure ratio was reduced using IHX, the stability of the compressor operating was enhanced.
Fig. 9. Compressor discharge temperature variation with mass fraction of CO2.
Pressure ratio of compressor
3.5 3.0 2.5 inlet 20 oC/outlet 45 oC inlet 20 oC/outlet 55 oC inlet 40 oC/outlet 45 oC inlet 20 oC/outlet 45 oC with IHX inlet 20 oC/outlet 55 oC with IHX inlet 40 oC/outlet 45 oC with IHX
2.0 1.5 1.0 0.5 0.4
0.5
0.6 0.7 0.8 0.9 Mass fraction of CO 2
1.0
1.1
Fig. 10. Compressor pressure ratio variation with mass fraction of CO2.
Input power of compressor (kW)
1.6
inlet 20 oC/outlet 45 oC with IHX inlet 20 oC/outlet 55 oC with IHX inlet 40 oC/outlet 45 oC with IHX
1.5 1.4
Acknowledgments
1.3
The authors are grateful for the support of the TD12-5048, 17JCYBJC29600 and 18JCYBJC22200 projects supplied by the Tianjin Natural Science Foundation.
1.2 1.1
References:
1.0
o
o
inlet 20 C/outlet 45 C inlet 20 oC/outlet 55 oC inlet 40 oC/outlet 45 oC
0.9 0.8 0.4
0.5
0.6
0.7
0.8
0.9
1.0
[1] G. Lorentzen, J. Pettersen, A new efficient and environmentally benign system for car air-conditioning, Int. J. Refrig 16 (1993) 4–12. [2] N. Abas, A.R. Kalair, N. Khan, A. Haider, Z. Saleem, M.S. Saleem, Natural and synthetic refrigerants, global warming: a review, Renew. Sustain. Energy Rev. 90 (2018) 557–569. [3] K. Nawaz, B. Shen, A. Elatar, V. Baxter, O. Abdelaziz, Performance optimization of CO2 heat pump water heater, Int. J. Refrig. 85 (2018) 213–228. [4] Á.Á. Pardiñas, A. Hafner, K. Banasiak, Novel integrated CO2 vapour compression racks for supermarkets thermodynamic analysis of possible system configurations and influence of operational conditions, Appl. Therm. Eng. 131 (2018) 1008–1025. [5] D. Wang, B. Yu, W. Li, J. Shi, J. Chen, Heating performance evaluation of a CO2 heat pump system for an electrical vehicle at cold ambient temperatures, Appl. Therm. Eng. 142 (2018) 656–664. [6] J. Li, J. Jia, L. Huang, S. Wang, Experimental and numerical study of an integrated fin and micro-channel gas cooler for a CO2 automotive air-conditioning, Appl. Therm. Eng. 116 (2017) 636–647. [7] R. Llopis, C. Sanz-Kock, R. Cabello, D. Sánchez, L. Nebot-Andrés, J. Catalán-Gil, Effects caused by the internal heat exchanger at the low temperature cycle in a cascade refrigeration plant, Appl. Therm. Eng. 103 (2016) 1077–1086. [8] P. Bansal, A review–status of CO2 as a low temperature refrigerant: fundamentals and R&D opportunities, Appl. Therm. Eng. 41 (2012) 18–29. [9] O. Joneydi Shariatzadeh, S.S. Abolhassani, M. Rahmani, M. Ziaee Nejad, Comparison of transcritical CO2 refrigeration cycle with expander and throttling valve including/excluding internal heat exchanger: exergy and energy points of
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Mass fraction of CO 2 Fig. 11. Compressor input power variation with mass fraction of CO2.
using IHX, which indicates that IHX can slightly reduce the compressor input power in the subcritical cycle. Theoretically, the compressor suction temperature can be increased by the use of IHX, resulting in an increase in the compressor input power, but this must be based on the constant compressor pressure ratio. Fig. 10 presents that the compressor pressure ratio was decreased significantly after using IHX when the mass fraction of CO2 was 0.6, which leads to a significant reduction in the specific power of the compression process. As can be seen from Figs. 10 and 11, the influence of the compressor pressure ratio on the compressor input power is obvious. When the water temperatures at the gas cooler/condenser inlet and outlet were 20 °C/45 °C, the compressor 7
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[20] G. He, F. Liu, D. Cai, J. Jiang, Experimental investigation on flow boiling heat transfer performance of a new near azeotropic refrigerant mixture R290/R32 in horizontal tubes, Int. J. Heat Mass Transf. 102 (2016) 561–573. [21] Q. Tian, D. Cai, L. Ren, W. Tang, Y. Xie, G. He, F. Liu, An experimental investigation of refrigerant mixture R32/R290 as drop-in replacement for HFC410A in household air conditioners, Int. J. Refrig 57 (2015) 216–228. [22] B. Wang, Z. Cheng, W. Shi, X. Li, Numerical study of gas injected heat pump using zeotropic R32/R1234ze(E) mixture: comparison of two type economizers, Appl. Therm. Eng. 142 (2018) 410–420. [23] C. Fan, G. Yan, J. Yu, Theoretical study on a modified heat pump cycle with zeotropic mixture R32/R290 for district heating in cold region, Appl. Therm. Eng. 156 (2019) 702–707. [24] J. Xia, J. Wang, G. Zhang, J. Lou, P. Zhao, Y. Dai, Thermo-economic analysis and comparative study of transcritical power cycles using CO2-based mixtures as working fluids, Appl. Therm. Eng. 144 (2018) 31–44. [25] G. Di Nicola, G. Giuliani, F. Polonara, R. Stryjek, Blends of carbon dioxide and HFCs as working fluids for the low-temperature circuit in cascade refrigerating systems, Int. J. Refrig. 28 (2005) 130–140. [26] A. Hakkaki-Fard, Z. Aidoun, M. Ouzzane, Applying refrigerant mixtures with thermal glide in cold climate air-source heat pumps, Appl. Therm. Eng. 62 (2014) 714–722. [27] Z. Sun, S. Liu, Y. Liang, M. Song, J. Guo, The optimal charge of carbon dioxide in water–water heat pump systems with and without an internal heat exchanger, HKIE Trans. 24 (2017) 99–106. [28] R.J. Moffat, Describing the uncertainties in experimental results, Exp. Therm Fluid Sci. 1 (1) (1988) 3–17. [29] B. Dai, M. Li, C. Dang, W. Yu, Y. Ma, Effects of lubricating oil on thermal performance of water-cooled carbon dioxide gas cooler, Appl. Therm. Eng. 80 (2015) 288–300. [30] J.H. Kim, J.M. Cho, M.S. Kim, Cooling performance of several CO2/propane mixtures and glide matching with secondary heat transfer fluid, Int. J. Refrig. 31 (2008) 800–806.
view, Appl. Therm. Eng. 93 (2016) 779–787. [10] M. Goodarzi, A. Gheibi, M. Motamedian, Comparative analysis of an improved twostage multi-inter-cooling ejector-expansion trans-critical CO2 refrigeration cycle, Appl. Therm. Eng. 81 (2015) 58–65. [11] D. Astrain, A. Merino, L. Catalán, P. Aranguren, M. Araiz, D. Sánchez, R. Cabello, R. Llopis, Improvements in the cooling capacity and the COP of a transcritical CO2 refrigeration plant operating with a thermoelectric subcooling system, Appl. Therm. Eng. 155 (2019) 110–122. [12] B. Yu, D. Wang, C. Liu, F. Jiang, J. Shi, J. Chen, Performance improvements evaluation of an automobile air conditioning system using CO2-propane mixture as a refrigerant, Int. J. Refrig. 88 (2018) 172–181. [13] L. Zhang, S. Xu, P. Du, H. Liu, Experimental and theoretical investigation on the performance of CO2/propane auto-cascade refrigerator with a fractionation heat exchanger, Appl. Therm. Eng. 87 (2015) 669–677. [14] Y. Zhu, X. Wu, Z. Wei, Heat transfer characteristics and correlation for CO2/propane mixtures flow evaporation in a smooth mini tube, Appl. Therm. Eng. 81 (2015) 253–261. [15] X.P. Zhang, F. Wang, X.W. Fan, X.L. Wei, F.K. Wang, Determination of the optimum heat rejection pressure in transcritical cycles working with R744/R290 mixture, Appl. Therm. Eng. 54 (2013) 176–184. [16] B. Yu, J. Yang, D. Wang, J. Shi, Z. Guo, J. Chen, Experimental energetic analysis of CO2/R41 blends in automobile air-conditioning and heat pump systems, Appl. Energy 239 (2019) 1142–1153. [17] D. Wang, Y. Lu, L. Tao, Thermodynamic analysis of CO2 blends with R41 as an azeotropy refrigerant applied in small refrigerated cabinet and heat pump water heater, Appl. Therm. Eng. 125 (2017) 1490–1500. [18] B. Dai, C. Dang, M. Li, H. Tian, Y. Ma, Thermodynamic performance assessment of carbon dioxide blends with low-global warming potential (GWP) working fluids for a heat pump water heater, Int. J. Refrig. 56 (2015) 1–14. [19] A. Mota-Babiloni, J. Navarro-Esbrí, P. Makhnatch, F. Molés, Refrigerant R32 as lower GWP working fluid in residential air conditioning systems in Europe and the USA, Renew. Sustain. Energy Rev. 80 (2017) 1031–1042.
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Applied Thermal Engineering journal homepage: www.elsevier.com/locate/apthermeng
Experimental study on CO2/R32 blends in a water-to-water heat pump system
T
⁎
Zhili Sun , Qi Cui, Qifan Wang, Jinghong Ning, Jianghe Guo, Baomin Dai, Yongqiang Liu, Yibo Xu Tianjin Key Laboratory of Refrigeration Technology, Tianjin University of Commerce, Tianjin 300134, China
H I GH L IG H T S
experiment of CO /R32 blends in a water-to-water heat pump system is conducted. • The and 65.2% maximum efficiency improvement in heating and cooling are obtained. • 23.3% optimal mixing ratio of CO /R32 (0.6/0.4) is found. • The pressure, cooling capacity, and heating capacity are reduced. • System • The internal heat exchanger further improves the system performance. 2
2
A R T I C LE I N FO
A B S T R A C T
Keywords: Water-to-water heat pump system CO2/R32 blends Thermal matching Internal heat exchanger
The purpose of this study is to evaluate the cooling and heating performance as well as the compressor operating conditions of CO2/R32 blends in a water-to-water heat pump system. Experiments were conducted to reveal the effects of the refrigerant compositions, heat source temperatures, and internal heat exchanger (IHX) on the performance of the system. The results show that 23.3% and 65.2% maximum improvement in heating coefficient of performance (COPh) and cooling coefficient of performance (COPc) of the system were obtained when the mass fraction of CO2 was 0.6, respectively. Meanwhile, heating capacity, cooling capacity, and the compressor discharge pressure of CO2/R32 blends were decreased significantly. The thermal matching between the heat source and CO2/R32 blends significantly affected the overall performance of the system. In addition, IHX is recommended in the water-to-water heat pump to improve the system performance.
1. Introduction CO2 is considered to be the most promising refrigerant [1] due to its excellent properties: 0 ozone depletion potential (ODP), low global warming potential (GWP), non-toxic, non-flammable, low viscosity, and high volumetric capacity [2], which promotes the applications of CO2 in heat pump water heater [3], commercial refrigeration and heat pump systems [4,5], automobile air conditioning [6], freezing and cold storage [7]. However, the large-scale applications of CO2 are limited by the two inevitable issues [8]: (1) high initial costs and the safety issues due to the significant high operating pressure of CO2. (2) low energy efficiency as a result of the large throttling losses during the expansion process. Many solutions are used for addressing these two issues, such as improving the expansion process by using expander or ejector instead
⁎
of throttle expansion valve [9], improving the compression process by adopting multi-stage compression intermediate cooling technology [10], and improving the system performance by mechanical subcooling [11]. While the system performance is improved and the system operating pressure is reduced, these solutions inevitably have a limitation, the assisted equipment result in additional costs. In view of this situation, numerous studies had been conducted in CO2 blends by some researchers. In the field of heating, ventilation, air-conditioning (HVAC), and heat pump, the researches of CO2 blends are mainly focused on CO2/ propane. Yu et al. [12] conducted a theoretical analysis and experimental study on the CO2/propane blends, the results show that the system coefficient of performance (COP) of the CO2/propane blends is 29.4% higher than that of the pure CO2, and the high pressure is reduced during the research range. Zhang et al. [13] carried out an
Corresponding author at: Tianjin Key Laboratory of Refrigeration Technology, Tianjin University of Commerce, Tianjin 300134, China. E-mail address: [email protected] (Z. Sun).
https://doi.org/10.1016/j.applthermaleng.2019.114303 Received 30 January 2019; Received in revised form 17 August 2019; Accepted 22 August 2019 Available online 23 August 2019 1359-4311/ © 2019 Elsevier Ltd. All rights reserved.
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Nomenclature Q W Cp m T/t s
e,wout g,w g,win g,wout
cooling capacity or heating capacity (kW) the compressor input power (kW) specific heat (kJ kg−1 K−1) mass flow rate (kg s−1) temperature (°C) specific entropy (kJ kg−1 K−1)
Acronyms COP IHX ODP GWP HVAC
Subscripts c h e g e,w e,win
the water at the evaporator outlet hot water in the gas cooler/condenser the water at the gas cooler/condenser inlet the water at the gas cooler/condenser outlet
cooling heating evaporator gas cooler/condenser chilling water in the evaporator the water at the evaporator inlet
coefficient of performance internal heat exchanger ozone depletion potential global warming potential heating, ventilation, and air-conditioning
Greek letters δ
uncertainty
district heating in cold region. The results show that as the refrigerant mass fraction of R32 increases, the system COP increases. The above studies about CO2-based and R32-based blends indicated that they have excellent properties and application prospects. However, when it comes to CO2/R32 blends, there are only a few studies, of which Xia et al. [24] carried out the thermo-economic analysis and comparative study of trans-critical power cycles using CO2/R32 blends as working fluids. The results show that among the CO2-based blends and pure CO2, CO2/R32 presents the highest exergy efficiency of 52.85%. Di Nicola et al. [25] analyzed the performances of a cascade refrigeration cycle operated with CO2/R32 blends as the low-temperature working fluid and concluded that CO2/R32 blends can be considered an attractive option for the low-temperature-circuit in cascade systems. Dai et al. [18] theoretically concluded that CO2/R32 is more suitable for its fairly high COP and lower high-side pressure in the case of moderate to high external temperature glide. Hakkaki-Fard et al. [26] developed a detailed screening heat pump theoretical model and assessed the performance of CO2/R32 blends. The results show that the blends of CO2/R32 (20/80 mass fraction) have the best performance among the mixtures studied. However, these results of the above studies about CO2/R32 blends have not been verified by experiments, a gap remains between the theoretical analyses and experimental study. Therefore, it is critical for experiments to be conducted and the cooling and heating performance as well as the compressor operating conditions of CO2/R32 blends to be evaluated. This study aims to contribute to filling the knowledge gap about the experimental study on CO2/R32 blends in the heat pump system. This paper is the first step of this research project, in which experiments are conducted to reveal the effects of the refrigerant compositions, heat source temperatures, and internal heat exchanger (IHX) on the performance of the system. The research on the optimal amounts of CO2/R32 blends charge, the heat matching characteristics of refrigerants and high-temperature and low-temperature heat sources, the blends composition shift, leakage and safety analyses will be conducted later by our research group.
experimental and theoretical investigation on the performance of CO2/ propane auto-cascade refrigerator and concluded that the cycle performance will be improved by increasing CO2 mass fraction. Zhu et al. [14] experimentally investigated the flow evaporation heat transfer characteristics of CO2/propane blends. The results show that the heat transfer coefficient of blends is between that of pure CO2 and pure propane for the same conditions, and increases with increasing CO2 proportion. Zhang et al. [15] carried out simulation and experimental investigations to research the relationships between optimum heat rejection pressure and related operating parameters for a trans-critical system using CO2/propane mixture. The results show that for a transcritical cycle working with CO2/propane mixture in which the mass fraction of CO2 is greater than 0.78, there is an optimal heat rejection pressure, under which a maximum system COP can be reached. There are also some studies on CO2/R41 blends. Yu et al. [16] carried out an experimental energetic analysis of CO2/R41 blends in automobile air conditioning and heat pump systems. The results demonstrate that the system COP in the heating and cooling modes can be improved up to a maximum of 14.5% and 25.7%, respectively. Wang et al. [17] conducted a thermodynamic analysis of CO2 blends with R41 applied in a small refrigerated cabinet and heat pump water heater. The results show that the CO2/R41 blends reduce the optimal high pressure of pure CO2 system, improve the system performance. Dai et al. [18] conducted a thermodynamic analysis of CO2/R41 blends in heat pump water heater, the results show that CO2/R41 is a suitable candidate for heat pump water heaters because of its high COP and low high-side pressure. Propane and R41 are environmentally friendly refrigerants with 0 ODP and low GWP. However, R32, which is also an environmental refrigerant especially in the transition stage of refrigerant substitution, is seldom paid attention to CO2 blends. Moreover, R32 is superior to propane in flammability and applicability. A great number of R32 blends have been recently proposed to replace R22 and R410A [19]. He et al. [20] carried out an experimental investigation on flow boiling heat transfer performance of propane/R32 in horizontal tubes and concluded that the new refrigerant mixture is near azeotropic, 0 ODP, and low GWP, which will be a competitive R22 replacement in household air conditioners. Tian et al. [21] conducted an experimental investigation of refrigerant mixture R32/propane as a drop-in replacement for 410A in household air conditioners. The results show that cooling and heating capacities are increased by 14.0–23.7%. Wang et al. [22] carried out a numerical study of gas injected heat pump using zeotropic R32/R1234ze(E) blends, and concluded that R32/R1234ze(E) zeotropic mixture is a potential replacement in the field of air conditioning and heat pumps. Fan et al. [23] conducted a theoretical study on a modified heat pump cycle with zeotropic mixture R32/propane for
2. Experiments 2.1. Experiment setup Fig. 1 illustrates a schematic diagram of the experiment setup of CO2/R32 blends in a water-to-water heat pump system. Table. 1 presents the main equipment parameters and accuracies of the devices. This system was composed of a compressor, oil separator, gas cooler/ condenser, IHX, hand expansion valve, evaporator, gas-liquid separator, hot water tank, chilling water tank, two water pumps, 2
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TP
1
2
Table 2 Test conditions. T
7
4 T
3
TP
TP
TP
TP
6
M
5 T
T
T
8
M
11
T
9
M
10
NO.
Mass fraction of CO2
Refrigerant charge (kg)
tg,win (°C)
tg,wout (°C)
te,win (°C)
te,wout (°C)
1 2 3 4 5 6 7
100% 95% 90% 80% 70% 60% 50%
2.0
20 20 40
45 55 45
12
7
evaporator outlet could be superheated, which leads to the increase in cooling capacity and heating capacity. The variation trend of cooling coefficient of performance (COPc) and heating coefficient of performance (COPh) are determined by the increasing degree of the compressor input power, cooling capacity and heating capacity, which are closely related to the thermophysical properties of refrigerants. When the increasing degree of cooling capacity and heating capacity are greater than the increasing degree of the compressor input power, COPc and COPh will increase, vice versa. The test methods are as follows. Firstly, the experiments of each group without IHX were conducted by opening the bypass valve ① and closing the bypass valve ②. The opening degree of the hand expansion valve was controlled to maintain the temperatures of CO2/R32 blends after throttling at 0 °C. Water temperature in the hot water tank was maintained at 20 °C by means of the thermostatic control system, and it was heated to 45 °C in the gas cooler/condenser by changing its volume flow rate, which achieving the working condition that the water temperature at the gas cooler/condenser inlet was 20 °C and the water temperature outlet was 45 °C. Similarly, the water temperatures at the evaporator inlet and outlet were maintained at 12 °C/7 °C. The experimental data were recorded after 30 min of stable operation of the system. Thereafter, the water temperatures at the gas cooler/condenser inlet and outlet were changed to 20 °C/55 °C and 40 °C/45 °C, respectively and the experimental data were recorded, respectively. Finally, the bypass valve ① was closed and the bypass valve ② was opened to carry out the experiments with IHX and then the above experimental steps without IHX were repeated to complete experiments of this group.
12
1-Compressor; 2-Oil separator; 3-Gas cooler/Condenser; 4- IHX; 5-Hand expansion valve; 6Evaporator; 7-Gas-liquid separator; 8,9-Pump; 10-Thermostatic control system; 11-Hot water tank; 12-Chilling water tank; T-Temperature sensor; P-Pressure sensor; M-water volume flow meter or refrigerant mass flow meter
Fig. 1. Schematic diagram of the experiment setup of CO2/R32 blends in a water-to-water heat pump system.
thermostatic control system (a refrigeration unit and electric heater), and several bypass valves. 2.2. Test conditions and test method The experiments were conducted under various test conditions as presented in Table. 2. Before experiments of each group were carried out, all refrigerants in the system must be extracted first, and then the system should be vacuum treated. Because the charging pressure of CO2 is higher than that of R32, R32 was charged first, followed by CO2. It is informed here that the amounts of CO2/R32 blends charge were all set to 2.0 kg [27]. And the experimental study on the optimal amounts of CO2/R32 blends charge will be conducted later by our research group. IHX was used in the system to explore the possibility of improving system performance. Theoretically, the compressor suction temperature can be increased by the use of IHX, resulting in an increase in the compressor input power based on the constant compressor pressure ratio. However, at the same time, the refrigerant at the gas cooler/ condenser outlet could be subcooled and the refrigerant at the
2.3. Data verification All Pt100 sensors were calibrated before the experiment. First, they were bundled with a four-wire Pt25 standard platinum resistor and placed in a thermostatic liquid tank. The temperature measurement
Table 1 The main equipment parameters and accuracies of the devices. NO.
Equipment
Main parameters
Model
Accuracy
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
Compressor Power-meter Compressor oil Gas cooler/condenser Hand expansion valve Evaporator IHX Hot water tank Cool water tank Temperature sensor Pressure sensor Water volume flow meter Refrigerant mass flow meter Water pump Oil separator Gas-liquid separator Electric heater
50 Hz, 1450 rad min−1, Discharge volume 1.12 m3 h−1 Rated voltage 3X220/380 V Density 0.974 g cm−3 (20 °C), Open flash point 242 °C Heat transfer area 1.8 m2, Design pressure 16 MPa Specification for circulation area φ12 Heat transfer area 1.5 m2, Design pressure 10 MPa Heat transfer area 0.2 m2, Design pressure 16 MPa Tank volume 0.225 m3 Tank volume 0.225 m3 Measuring range of −50–450 °C Measuring range of 0–16 MPa Measuring range of 0.6–6 m3 h−1 Measuring range of 0–100 kg h−1 Max flow 2 m3 h−1, Head 15 m, Power 0.37 kW, 2900 r min−1 Diameter 11 cm, height 52 cm Diameter 11 cm, height 52 cm Input power 200 W
CD180H 7p-0.5 POE 130 Double-pipe
/ 0.5% FS / / / / / / / ± 0.1 °C 0.5% FS 0.5% FS 0.2% FS / / / /
3
Double-pipe Double-pipe
PT100 DTT0X LWGYC-15 CMFI CHL2-20 LSWSC
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of CO2. This is because the addition of R32 reduces the blends′ density, thus reducing the blends’ mass flow rate as shown in Fig. 5. That eventually reduces the heating capacity. It can be observed that when the mass fraction of CO2 decreased from 0.95 to 0.9, both COPh and heating capacity decreased significantly, which indicates that the mass fraction of CO2 at 0.9 is a significant point, where the heat transfer performance between the blends and the heat source of the system is poor. Fig. 2 shows that when the mass fraction of CO2 was 0.9, the cycle form of the system was the trans-critical cycle, where the heat transfer of the blends was in the pseudo-critical area (above the critical point). In the pseudo-critical area, the heat transfer particularity of supercritical fluid is mainly manifested in the fact that the specific heat of supercritical fluid is sometimes higher or lower than that of conventional single-phase forced convection, and reaching the maximum at the pseudo-critical point [29]. Besides, the maximum specific heat increases as the operating pressure decreases, which results in the maximum average specific heat in the heat transfer process of the gas cooler, when the mass fraction of CO2 is 0.9. Limited by the small heat transfer area of the gas cooler, the temperature of CO2/R32 (0.9/0.1) changed slightly, which accounts for the temperature at the gas cooler outlet was larger than other blends. The experimental data show that the temperature at the gas cooler outlet of CO2/R32 (0.9/0.1) was 3.8 °C higher on average than other blends for three working conditions. The higher the outlet temperature of the gas cooler, the greater the end temperature difference and an increase in entropy generation of the gas cooler, eventually leads to poor heat transfer performance. This situation can be improved by increasing the heat transfer area of the gas cooler. For most working conditions, when the mass fraction of CO2 was 0.6, the COPh of the system reached its maximum value. Compared with pure CO2, the COPh increased by 10.3%, 19.2%, and 12.1%, respectively, for three working conditions. Under the same baseline, this value would change to 16.1%, 23.3%, and 12.5%, respectively after using IHX, which indicates that the COPh can be further increased by the use of IHX as shown in Fig. 3. This is because, for the CO2/R32 blends in the subcritical cycle, the heating capacity is increased while the compressor input power is decreased (See Fig. 11) after using IHX, which inevitably leads to an increase in COPh. However, when the mass fraction of CO2 was 0.6, compared with pure CO2, the heating capacity decreased by 15.9%, 10.4%, and 14.7%, respectively, for three working conditions. Thus, the balance between heating capacity and COPh at three working conditions should be weighed deliberately. Although when the mass fraction of CO2 was greater than 0.8, the use of IHX reduced the mass flow rate of the blends presented in Fig. 5, but it still increased the heating capacity of all blends as a result of the superheating degree at the evaporator outlet increases. In summary, IHX is recommended in the water-to-water heat pump to improve the heating performance of
accuracy of the four-wire Pt25 standard platinum resistor is very high, and the measured values can be accurate to 5 mK that is acceptable for this experiment. The measured values with high accuracy of the fourwire Pt25 standard platinum resistor were read out by the multimeter as the “true value” of temperature. Then, the temperature of the liquid in the thermostatic tank was changed per 5 °C in the temperature range of −5 to 150 °C, and the measured values of all Pt100 sensors and the four-wire Pt25 standard platinum resistor were recorded. Finally, the linear relationship between measured values by all Pt100 sensors and measured values by the four-wire Pt25 standard platinum resistor were fitted by linear regression, and each Pt100 sensor was calibrated. Consequently, the temperatures measured by the calibrated Pt100 sensors are convincing. The sensor locations were presented in Fig. 1. The accuracies of the measuring devices were shown in Table. 1. The cooling and heating capacities of the water-to-water heat pump system were determined by the water-side heat transfer, using Eqs. (1) and (2). The errors between the right side and the left side of Eq. (3) were within ± 10%, indicated that the tested system exhibits an effective heat balance between the high-temperature heat source and low-temperature heat source sides. COPc and COPh were determined by Eqs. (4) and (5). Uncertainty analysis was performed to verify the cooling/heating capacity and COP, the indirect measurements, using Eq. (6) [28], where R represents the targeted capacity or COP, Xi represents the uncertainty of its affecting factors, and N is the number of affecting factors. Based on that, the relative uncertainties of the system cooling/heating capacity and COP were calculated as 0.6–2.0% and 0.8–2.1%, respectively.
Qc = Cp me, w (te, win − te, wout )
(1)
Qh = Cp mg, w (tg, wout − tg, win )
(2)
Qh = Qc + W
(3)
COPc = Qc / W
(4)
COPh = Qh/ W
(5)
δR = R
N
2
δXi ⎞ ⎝ Xi ⎠
∑⎛
⎜
i=1
⎟
(6)
3. Results and discussion 3.1. Effect of CO2/R32 blends on the system performance The critical temperatures of CO2 and R32 are 31.1 °C and 78.1 °C, respectively. When the temperature of high-temperature heat source reaches a certain high value, the cooling process without phase change of CO2 is completed above the critical point, where the cycle form of the system is called the trans-critical cycle and fluids above the critical point are called supercritical fluids. As R32 with higher critical temperature blended with CO2, the critical temperature of the CO2 blends increased. That indicates that there is a mixing ratio where the CO2 blends condensate below the critical point, the cycle form of the system is called the subcritical cycle. Fig. 2 presents the T-s diagrams of the system at mass fraction of CO2 with 0.8 and 0.9, under the experimental conditions. When the mass fraction of CO2 was 0.9, the cycle form of the system was the trans-critical cycle, and the high-temperature heat source heat exchanger is called gas cooler. However, when the mass fraction of CO2 was 0.8, the cycle form of the system was the subcritical cycle, and the high-temperature heat source heat exchanger is called condenser. Figs. 3 and 4 illustrate the variation of COPh and heating capacity with mass fraction of CO2. As can be seen from Fig. 3, the addition of R32 increased the COPh of pure CO2, which means R32 is the refrigerant having a higher heating effect than CO2. However, Fig. 4 presents the heating capacity decreased with reducing the mass fraction
380 360
Mass fraction of CO2 at 0.8 Mass fraction of CO2 at 0.9
T (K)
340 320 300 280 260 240 0.8
1.0
1.2
1.4
1.6
1.8
2.0
2.2
s (kJ/(kg·K)) Fig. 2. T-s diagram of the system at mass fraction of CO2 with 0.8 and 0.9. 4
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6.0 5.5 5.0
COPh
4.5 4.0
among the three working conditions. The experimental data show that the temperature glides of refrigerant in the gas cooler/condenser increased in turn under the working conditions where the water temperatures at the gas cooler/condenser inlet and outlet were 20 °C/45 °C, 20 °C/55 °C, and 40 °C/45 °C, respectively (whether IHX was used or not). For the three working conditions, the average temperature glides of refrigerant were 66.9 °C, 80.1 °C, and 91.9 °C, respectively, and this value would change to 74.3 °C, 86.2 °C, and 93.9 °C, respectively after using IHX. When the water temperatures at the gas cooler/condenser inlet and outlet were 20 °C/45 °C, the temperature glides of refrigerant in the gas cooler/condenser were the smallest, and that leads to the better thermal matching with the heat source, which results in the largest COPh and heating capacity among the three working conditions. Compared with the other two working conditions, the COPh increased by 10.9% and 49.5%, respectively and the heating capacity increased by 2.0% and 32.3%, respectively, when the mass fraction of CO2 was 0.6. When the water temperatures at the gas cooler/condenser inlet and outlet were 40 °C/45 °C, the COPh and heating capacity of the system were the worst, due to the slight temperature changes of hot water, which results in a mismatch with the larger temperature variations of the blends. As can been observed form Figs. 6 and 7, the effects of the refrigerant compositions, heat source temperatures, and IHX on the COPc and cooling capacity were basically the same as similar to COPh and heating capacity. The addition of R32 increased the COPc of pure CO2, which means R32 is also the refrigerant having a higher cooling effect than CO2. The cooling capacity decreased with reducing the mass fraction of CO2, which as a result of the reduction in the mass flow rate. For most working conditions, when the mass fraction of CO2 was 0.6, the COPc of the system also reached its maximum value. Compared with pure CO2, the COPc increased by 22.6%, 20.1%, and 31.2%, respectively, for three working conditions. Under the same baseline, this value would change to 29.3%, 29.8%, and 65.2%, respectively after using IHX, which indicates that the COPc can be further increased by the use of IHX because the cooling capacity is increased while the compressor input power is decreased. And the influence of IHX on cooling performance is greater than that of heating performance. Moreover, when the mass fraction of CO2 was 0.6, compared with pure CO2, the cooling capacity decreased by 6.5%, 9.7%, and 0.1%, respectively, for three working conditions. Similarly, the balance between cooling capacity and COPc at three working conditions should be weighed deliberately. IHX also increased the cooling capacity of all blends as a result of the subcooling degree at the gas cooler/condenser outlet increases. To sum up, IHX is recommended in the water-to-water heat pump to improve the cooling performance of the system. Obviously, the thermal matching between the high-temperature heat source and blends affected the heat transfer between the lowtemperature heat source and blends. When the water temperatures at
inlet 20 oC/outlet 45 oC inlet 20 oC/outlet 55 oC inlet 40 oC/outlet 45 oC inlet 20 oC/outlet 45 oC with IHX inlet 20 oC/outlet 55 oC with IHX inlet 40 oC/outlet 45 oC with IHX
3.5 3.0 2.5 2.0 1.5 0.4
0.5
0.6 0.7 0.8 0.9 Mass fraction of CO 2
1.0
1.1
Fig. 3. COPh variation with mass fraction of CO2.
5.0
Heating capacity (kW)
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inlet 20 oC/outlet 45 oC with IHX inlet 20 oC/outlet 55 oC with IHX inlet 40 oC/outlet 45 oC with IHX
3.5 3.0 2.5
inlet 20 oC/outlet 45 oC inlet 20 oC/outlet 55 oC inlet 40 oC/outlet 45 oC
2.0 1.5 0.4
0.5
0.6 0.7 0.8 0.9 Mass fraction of CO 2
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Fig. 4. Heating capacity variation with mass fraction of CO2.
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inlet 20oC/outlet 45oC inlet 20oC/outlet 55oC inlet 40oC/outlet 45oC inlet 20oC/outlet 45oC with IHX inlet 20oC/outlet 55oC with IHX inlet 40oC/outlet 45oC with IHX
Mass flow rate (10-3 kg/s)
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inlet 20 oC/outlet 45 oC with IHX inlet 20 oC/outlet 55 oC with IHX inlet 40 oC/outlet 45 oC with IHX
3.6 3.2
Fig. 5. Mass flow rate of refrigerant variation with mass fraction of CO2.
2.8 COPc
the system. The experiments were carried out at three different temperatures of high-temperature heat sources. The CO2/R32 blends are non-azeotropic, which indicates that the temperature glide occurs during its phase change. If the temperature of the heat source and the blends reach the thermal matching, that is, the gradients of temperature change are similar, and the system performance will be greatly improved. Obviously, as can be seen from Figs. 3 and 4, when the water temperatures at the gas cooler/condenser inlet and outlet were 20 °C/ 45 °C, the COPh and heating capacity of the system were the largest
2.4 2.0 1.6 inlet 20 oC/outlet 45 oC inlet 20 oC/outlet 55 oC inlet 40 oC/outlet 45 oC
1.2 0.8 0.4
0.4
0.5
0.6 0.7 0.8 0.9 Mass fraction of CO 2
1.0
Fig. 6. COPc variation with mass fraction of CO2. 5
1.1
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5.0 Cooling capacity (kW)
4.5 4.0 3.5 3.0
at the high-temperature side is cooled more fully and the high pressure is reduced. As mentioned above, when the water temperatures at the gas cooler/condenser inlet and outlet are 20 °C/45 °C, the system has the best heating and cooling performance among the three working conditions due to the better thermal matching, which results in the lowest compressor discharge pressure. Compared with the other two working conditions, the compressor discharge pressure decreased by 10.6% and 19.2%, respectively, when the mass fraction of CO2 was 0.6. Fig. 9 indicates that for most working conditions, blended R32 with CO2 reduced the compressor discharge temperature, and it further reduced with the increase of R32 composition, but the reduction was only within 20 °C. For most working conditions, the use of IHX increased the compressor discharge temperature, which is the inevitable result of the increase in the compressor suction temperature. The increasing range of the compressor discharge temperature is acceptable. When the water temperatures at the gas cooler/condenser inlet and outlet were 20 °C/ 45 °C, the compressor discharge temperature was the lowest, as a result of the compressor pressure ratio is the smallest under this working condition, as shown in Fig. 10. Compared with the other two working conditions, the compressor discharge temperature decreased by 10.9% and 28.3%, respectively, when the mass fraction of CO2 was 0.6. Figs. 10 and 11 show the variation of the compressor pressure ratio and the compressor input power with mass fraction of CO2. The compressor pressure ratio is the ultimate parameter reflecting the compressor operating conditions. Fig. 10 illustrates that for most working conditions, the compressor pressure ratio had an increasing trend with the increase of R32 composition, but the amplitude was too small except for the third working condition, which indicated that for any CO2/ R32 blends, the compressor operating conditions are stable. For most working conditions, the use of IHX had little influence on the compressor pressure ratio, especially in the trans-critical cycle. When the water temperatures at the gas cooler/condenser inlet and outlet were 20 °C/45 °C, the compressor pressure ratio was the smallest, which indicated that under this working condition, the compressor operating conditions are the best. Compared with the other two working conditions, the compressor pressure ratio decreased by 10.9% and 24.3%, respectively, when the mass fraction of CO2 was 0.6. Fig. 11 presents that blended R32 with CO2 reduced the compressor input power, and it reduced further with the increase of R32 composition, as a result of the reduction of mass flow rate indicated in Fig. 5. It can be observed that when mass fraction of CO2 decreased from 0.95 to 0.8, the compressor input power decreased slightly. The cause of this phenomenon is the same as similar to the compressor discharge pressure. When the mass fraction of CO2 was 0.6, compared with pure CO2, the compressor input power decreased by 23.7%, 24.8%, and 23.9%, respectively, for three working conditions. Under the same baseline, this value would change to 24.6%, 26.0%, and 24.4%, respectively after
inlet 20 oC/outlet 45 oC inlet 20 oC/outlet 55 oC inlet 40 oC/outlet 45 oC inlet 20 oC/outlet 45 oC with IHX inlet 20 oC/outlet 55 oC with IHX inlet 40 oC/outlet 45 oC with IHX
2.5 2.0 1.5 1.0 0.4
0.5
0.6 0.7 0.8 0.9 Mass fraction of CO 2
1.0
1.1
Fig. 7. Cooling capacity variation with mass fraction of CO2.
the gas cooler/condenser inlet and outlet were 20 °C/45 °C, the COPc and cooling capacity of the system were the largest among the three working conditions. Compared with the other two working conditions, the COPc increased by 20.5% and 88.1%, respectively and the cooling capacity increased by 10.9% and 66.5%, respectively, when the mass fraction of CO2 was 0.6. Essentially, the thermal matching between blends and chilling water is the key to affect the cooling performance of the system. However, this experiment only focused on the thermal matching between high-temperature heat source and blends, and there was only one group of the low-temperature heat source. The temperatures of CO2/R32 blends after throttling were maintained at 0 °C, and the experimental data show that the temperatures of CO2/R32 blends of the evaporator outlet were all close to 12 °C, the inlet chilling water temperature. Although the temperature glide of blends in the evaporator increases with the decrease of evaporating pressure. However, it can be found that the heat transfer between all blends and chilling water were still mismatched. In view of this, Kim et al. [30] concluded that this mismatch of temperature profiles may lead to pinching at the evaporator exit, and it was difficult to decrease the mean temperature difference between two heat transfer fluids by changing the heat exchanger area. Therefore, to improve the system efficiency, CO2/R32 blends must necessarily be used in applications with appropriate large temperature glides of the low-temperature heat source. The above results lay a foundation for the thermal matching experiment between the low-temperature heat source and blends conducted later by our research group.
Discharge pressure of compressor (MPa)
3.2. Effect of CO2/R32 blends on the compressor operating conditions Figs. 8 and 9 illustrate the variation of the compressor discharge pressure and the compressor discharge temperature with mass fraction of CO2. As shown in Fig. 8, blended R32 with CO2 reduced the compressor discharge pressure, and it reduced further with the increase of R32 composition, as a result of the normal boiling point of CO2 (−78.4 °C) is lower than that of R32 (−51.7 °C). It can be observed that when the mass fraction of CO2 decreased from 0.95 to 0.8, the compressor discharge pressure decreased slightly. As mentioned above, the heat transfer performance between the blends and the heat source of the system is poor when the mass fraction of CO2 is 0.9, which results in relatively high compressor discharge pressure. When the mass fraction of CO2 was 0.6, compared with pure CO2, the compressor discharge pressure decreased by 26.2%, 26.9%, and 21.4%, respectively, for three working conditions. Under the same baseline, this value would change to 27.5%, 28.4%, and 23.9%, respectively after using IHX, which indicates that IHX can further reduce the compressor discharge pressure in the subcritical cycle. This is because after using IHX, the refrigerant
11 10
inlet 20oC/outlet 45 oC with IHX inlet 20oC/outlet 55 oC with IHX inlet 40oC/outlet 45 oC with IHX
9 8 7 6
inlet 20oC/outlet 45 oC inlet 20oC/outlet 55 oC inlet 40oC/outlet 45 oC
5 4 0.4
0.5
0.6 0.7 0.8 0.9 Mass fraction of CO 2
1.0
1.1
Fig. 8. Compressor discharge pressure variation with mass fraction of CO2. 6
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Discharge temperature of compressor (oC)
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200
input power was the smallest, as a result of the compressor pressure ratio was the smallest. Compared with the other two working conditions, the compressor input power decreased by 8.0% and 11.5%, respectively, when the mass fraction of CO2 was 0.6.
inlet 20oC/outlet 45oC inlet 20oC/outlet 55oC inlet 40oC/outlet 45oC inlet 20oC/outlet 45oC with IHX inlet 20oC/outlet 55oC with IHX inlet 40oC/outlet 45oC with IHX
180 160 140
4. Conclusions Experiments were conducted to reveal the effects of the refrigerant compositions, heat source temperatures, and internal heat exchanger (IHX) on the cooling and heating performance as well as the compressor operating conditions of CO2/R32 blends in a water-to-water heat pump system. Based on the results and preceding discussion, the following conclusions can be obtained.
120 100 80 0.4
0.5
0.6 0.7 0.8 0.9 Mass fraction of CO2
1.0
1.1
(1). R32 is the refrigerant having a higher heating and cooling effect than CO2, 23.3% and 65.2% maximum improvement in heating coefficient of performance (COPh) and cooling coefficient of performance (COPc) were obtained when the mass fraction of CO2 was 0.6, respectively. Meanwhile, heating capacity and cooling capacity of CO2/R32 blends were decreased, as a result of the mass flow rate decreased due to the smaller vapor density of R32. R32 blended with CO2, the compressor discharge pressure was reduced significantly, and the compressor operating conditions were stable due to the compressor pressure ratio changed slightly. (2). The high-temperature heat source and the CO2/R32 blends reached the better thermal matching when the water temperatures at the gas cooler/condenser inlet and outlet were 20 °C/45 °C, which results in the optimum overall performance of the system. The lowtemperature heat source and the CO2/R32 blends all mismatched when the water temperatures at the evaporator inlet and outlet were 12 °C/7 °C, and thus CO2/R32 blends must necessarily be used in applications with appropriate large temperature glides of the low-temperature heat source. (3). IHX is recommended in the water-to-water heat pump to improve the heating and cooling performance of the system. IHX further reduced the compressor discharge pressure in the subcritical cycle. Due to the compressor pressure ratio was reduced using IHX, the stability of the compressor operating was enhanced.
Fig. 9. Compressor discharge temperature variation with mass fraction of CO2.
Pressure ratio of compressor
3.5 3.0 2.5 inlet 20 oC/outlet 45 oC inlet 20 oC/outlet 55 oC inlet 40 oC/outlet 45 oC inlet 20 oC/outlet 45 oC with IHX inlet 20 oC/outlet 55 oC with IHX inlet 40 oC/outlet 45 oC with IHX
2.0 1.5 1.0 0.5 0.4
0.5
0.6 0.7 0.8 0.9 Mass fraction of CO 2
1.0
1.1
Fig. 10. Compressor pressure ratio variation with mass fraction of CO2.
Input power of compressor (kW)
1.6
inlet 20 oC/outlet 45 oC with IHX inlet 20 oC/outlet 55 oC with IHX inlet 40 oC/outlet 45 oC with IHX
1.5 1.4
Acknowledgments
1.3
The authors are grateful for the support of the TD12-5048, 17JCYBJC29600 and 18JCYBJC22200 projects supplied by the Tianjin Natural Science Foundation.
1.2 1.1
References:
1.0
o
o
inlet 20 C/outlet 45 C inlet 20 oC/outlet 55 oC inlet 40 oC/outlet 45 oC
0.9 0.8 0.4
0.5
0.6
0.7
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