Thermodynamic Analysis of Comprehensive Performance of Carbon Dioxide(R744) and Its Mixture With Ethane(R170) Used in Refrigeration and Heating System at Low Evaporation Temperature

Traditional refrigerants have been phased out due to the destruction of the ozone layer and the greenhouse effect [1], while R744, a natural working fluid, has been paid more and more attention because of its good environmental friendliness (ozone depleting potential (ODP) = 0, global warming potential (GWP) = 1). The critical temperature of R744 is lower than the ambient temperature of typical summer conditions (35 ℃). Transcritical cycle mode is suitable for refrigeration and heating [2]. R744 at low temperature has very low viscosity and good heat transfer performance, which is suitable for low ambient conditions. However, there are two problems in the R744 transcritical cycle system. One is that the pressure on the high-pressure side of the system is too high, which makes the system unstable. Second, the efficiency of the cycle is low, so the above problems can be solved by improving the system itself and mixing other low GWP working fluids with R744.

In the aspect of the system, in order to improve the cycle coefficient of the R744 system cycle, Lorentzen [3] proposed a two-stage compression cycle in the research of the R744 cycle. Nickl et al. [4] used an expander to drive the compressor and acted as a high-pressure stage compressor of a two-stage compression system. In order to improve the overall efficiency by recovering the expansion work. Gu et al. [5] analyzed the R744 two-stage compression system at low evaporation temperature. Cecchinato et al. [6] analyzed that the coefficient of performance (COP) of the R744 refrigeration system with a two-stage compression system increased by 9%. Yang et al. [7] found that using an expansion valve instead of a throttle valve can improve the system performance by more than 30%. Yang et al. [8] found that the COP of the indirect subcooling system is 22% higher than that of the single-stage R744 transcritical system. Pitarch et al. [9] showed that the COP of the R744 two-stage air-source heat pump system was higher than that of the single-stage R744 system. Wang and Zhang [10] analyzed the performance of single-stage R744 heat pump systems under different subcooling modes. The results showed that the overall efficiency is the best when the internal heat exchanger is used for subcooling in winter. Wang et al. [11] showed that the COP of the R744 system is 57.4% higher than that of single-stage compression when a two-stage compression system is used. Zhang et al. [12] established five thermodynamic models for improving the COP of a single-stage compression system. The results showed that the COP of the improved system can be increased by 14–21%. Zhu et al. [13] conducted an experimental study on the ejector expansion transcritical R744 heat pump system. The results showed that the COP of the system is about 10.3% higher than that of the basic system. Llopis et al. [14] studied the thermodynamic performance of the mechanical subcooling system and the basic system. The results showed that the maximum COP of the mechanical subcooling system can be increased by about 20%, and the maximum cooling capacity can be increased by about 28.8%.

In the aspect of using the R744 refrigerant mixture for the transcritical system, Kim et al. [15] studied the performance of the R744/R290 refrigerant mixture in an air conditioning system through experiments, and the results showed that the discharge pressure of R744/R290 mixture decreased with the increase of propane mole fraction. Niu and Zhang [16] found that the R744/R290 (71/29) refrigerant mixture can replace R13 when the evaporation temperature exceeds 201 K. Zhang et al. [17] showed that for the transcritical cycle using the R744/R290 mixture. The results showed that when the mass fraction of R744 is greater than 0.78, there is an optimal discharge pressure, under which the maximum COP of the system can be achieved. Dai et al. [18] analyzed the application of R744 mixed with 10 kinds of low GWP refrigerants in a heat pump water heater. The results showed that R744/R41 was the best choice because of its high COP and low pressure on the high-pressure side. Kim et al. [19] carried out an experimental study on the performance of the R744/R290 mixture in a transcritical refrigeration cycle. The results showed that the pressure of the system was reduced, which confirmed the feasibility of reducing the pressure of the system by using a mixture of working fluids. Koyama et al. [20] used an R744/DME mixture in the transcritical refrigeration cycle. Compared with pure R744, the discharge pressure in the refrigeration and heating system decreased by 2 MPa and 1.9 MPa, respectively, while the COP of the system remained basically unchanged [20]. Yu et al. [21] applied an R744/R290 refrigerant mixture to an automobile air conditioning system. The experimental results showed that under the same operating conditions, when the mixing ratio is 60/40, the system performance is the best; COP is 29.4% higher than that of pure R744 system, and the discharge temperature is also reduced by 47 ℃. Sarkar et al. [22,23] used R600 and R600a refrigerant mixed with R744 in a certain proportion in the heat pump system. The results showed that the COP of the system has been significantly improved. Yu et al. [24] increased the mass fraction of R41; the coefficient of performance of pure R744 in heating and cooling mode can be increased by 14.5% and 25.7%, respectively, and the optimal pressure correlation of R744/R41 is fitted. Dubey et al. [25] theoretically analyzed the R744/R1270 cascade refrigeration system and reported that the R744/R270 system has higher performance than critical R744 cycle and N₂O/R744 transcritical cycle. Ju et al. [26] found that R744/R290 with a mass fraction of 12/88 was the most suitable natural substitute for R22 in heat pump water heaters, and its optimal heating cop and capacity were 11.00% and 17.50% higher than R22, respectively. Sun et al. [27] found that when the mass fraction of R744 was 0.6, the coefficient of thermal performance and coefficient of refrigeration performance of the system increased by 23.3% and 65.2%, respectively. Nicola et al. [28] analyzed the performance of the cascade refrigeration cycle with R744/R32 mixture as a low-temperature working fluid. The results showed that the R744/R32 mixture can be considered as an attractive choice for a low-temperature loop of cascade system. Xia et al. [29] used the R744/R32 mixture as a working fluid to analyze and compare the thermal economy of the transcritical power cycle. The results showed that among the R744/R32 mixture and pure R744, R744/R32 has the highest exergy efficiency (52.85%). Bhattacharyya et al. [30] optimized R744/C₃H₈ refrigeration and heating two-stage compression system. The results showed that the high-temperature cycle part of the system is very effective in improving the overall performance. Getu and Bansal [31] calculated the R134a/R744 cascade refrigeration system by linear regression analysis method and obtained that with the increase of high-temperature condensation temperature, the coefficient of performance of the system decreased and the refrigerant mass flow increased. Dopazo et al. [32] designed a 9-kw NH₃/R744 cascade horizontal plate refrigerator system under the ambient temperature of −50 ℃ and determined the optimal intermediate temperature to maximize the coefficient of performance of the system through the experiments of varying low-temperature evaporation temperature and high-temperature condensation temperature. Compared with the two-stage compression system, the COP of the cascade cycle system was obtained when the low-temperature evaporation temperature was lower than −40 ℃. It is larger than the conclusion of the two-stage compression cycle.

Because the refrigerant mixture of R744 can be divided into azeotropic and non-azeotropic types, the cooling water in the condensation process of the non-azeotropic refrigerant mixture is constantly changing, and the temperature of the object to be cooled in the evaporation process is constantly decreasing. According to the characteristics of variable temperature, the heat transfer temperature difference in the phase change process is reduced, the irreversible loss in the process is reduced, and the heat transfer irreversible loss in the condenser and evaporator is reduced; the efficiency of the refrigeration cycle can be improved. But in the operation of the system, if there is leakage, it will lead to a change in the ratio and then affect the overall system performance. If the azeotropic refrigerant mixture leaks, the impact on the system performance is very small. At the same time, whether the azeotropic refrigerant can improve the system performance compared with its pure working fluid needs to be analyzed according to different working fluids.

At present, there are few studies on the performance analysis of R744/170 in transcritical cycle systems [11,33] and low evaporation temperature conditions. On the one hand, compared to pure R744, the mixing of R744 and R170 provides a certain possibility for reducing system pressure. On the other hand, currently, almost all mixtures of R744 are non-azeotropic; non-azeotropic mixtures are prone to leakage during system operation (due to the presence of R744, the high-pressure side pressure of the system is too high). Therefore, this paper focuses on mixing R744 with different proportions of R170 to form an azeotropic and non-azeotropic mixture and makes a comprehensive comparative analysis of its refrigeration and heating performance in a transcritical cycle system at low-temperature conditions.

In this paper, R744/170 is composed of mixtures with different proportions. According to the temperature slip of the mixed refrigerant under specific evaporation pressure, azeotropic or non-azeotropic refrigerant mixtures are defined. The temperature slip of 0 or very small represents azeotropic refrigerant, and the rest is non-azeotropic refrigerant. The temperature slip of R744/170 mixtures with different proportions under specific evaporation pressure passes through refprop 9.0 [34], as shown in Tables 1–3; it can be found from Table 3 that R744/170 (77.6/22.4) is an azeotropic refrigerant if its temperature slip is very small at low evaporation temperature. In addition, the ODP and GWP of R744 (ODP = 0, GWP = 1), R170 (ODP = 0, GWP = 20), and R744/170 mixtures with different proportions are calculated by [35], and the results are shown in Table 4. Table 5 shows the critical temperature and pressure of each working medium.

Because the critical evaporation temperature of R744 and its mixture is low, the transcritical cycle is adopted in both refrigeration and heating systems. Fig. 1(a) is the system schematic diagram of refrigeration in a low-temperature environment studied in this paper, and Figs. 1(b) and 1(c) are the system schematic diagrams of heating. Figures 2 and 3 are the corresponding diagrams of the enthalpy of pressure and the temperature entropy. The whole system consists of a compressor, air cooler, throttle valve, evaporator, and internal heat exchanger. In Fig. 2, “2S” and “2” represent the state points of the working medium after isentropic compression and actual compression, respectively. The R744 transcritical cycle and R744/R170 cycle studied in this paper are the energy thermodynamics models established under optimal pressure.

All the analyses in this paper are based on the following assumptions:

1. The system is in a steady-state.

2. When the working fluid flows in the heat exchanger and connecting pipe, the pressure drop and heat loss are not considered.

3. The compression process is adiabatic but not isentropic.

4. At the outlet of the evaporator, the refrigerant is saturated with steam.

5. The refrigerant does not contain lubricating oil.

The design conditions of the transcritical system used for refrigeration and heating are shown in Tables 6 and 7.

Energy thermodynamic model

• Work input to compressor:

  $wcomp=h2−h1$

  (1)
• Heating effect of gas cooler:

  $qg=h2−h3$

  (2)
• Refrigerating effect of evaporator:

  $qe=h0−h5$

  (3)
• Energy balance in the internal heat exchanger:

  $h1−h0=h3−h4$

  (4)
• Isentropic efficiency of compressor:

  $εis=(h2s−h1)/(h2−h1)$

  (5)
• The effectiveness of the internal beat exchanger is given by Sarkar et al. [36] as 0.6:

  $β=(h1−h0)/(h3−h0)$

  (6)
• COP_h is given by:

  $COPh=(h2−h3)/(h2−h1)$

  (7)
• COP_e is given by:

  $COPe=(h0−h5)/(h2−h1)$

  (8)

For exergy analysis of a refrigeration and heating system, the parameters of the internal and external environment of the system should be considered, so each system has a corresponding exergy thermodynamic model. Based on the studied three systems, the respective exergy thermodynamic model is developed.

Exergy thermodynamic model

• Compressor irreversibility:

  $Icomp,re=Ten,re(S2−S1)$

  (9)

  $Icomp,h=Ten,h(S2−S1)$

  (10)

  $Icomp,air=Ten,air(S2−S1)$

  (11)

  where T_en,re is the external environment temperature of refrigeration (308 K), T_en,h is the external environment temperature of heating (263 K), and T_en,air is the external environment temperature of air-source heat pump water heater (274 K).
• Capillary irreversibility:

  $Icap,re=Ten,re(S5−S4)$

  (12)

  $Icap,h=Ten,h(S5−S4)$

  (13)

  $Icap,air=Ten,air(S5−S4)$

  (14)
• Internal heat exchanger irreversibility:

  $Iihe,re=Tihe,re[(S1−S0)+(S4−S3)]$

  (15)

  $Iihe,h=Tihe,h[(S1−S0)+(S4−S3)]$

  (16)

  $Iihe,air=Tihe,air[(S1−S0)+(S4−S3)]$

  (17)
• Evaporator irreversibility:

  $Ie,re=Ten,re[(S0−S5)+(h0−h5)/Tin]$

  (18)

  $Ie,h=Tin,h[(S0−S5)+(h0−h5)/Ten,h]$

  (19)

  $Ie,air=Ten,air[(S0−S5)+(h0−h5)/Ten,air]$

  (20)

  where T_in is the internal temperature of refrigeration (276 K), and T_in,h is the internal temperature of refrigeration (274 K).
• Gas cooler irreversibility:

  $Ig,re=Ten,re[(h2−h3)/Ten,re−(S2−S3)]$

  (21)

  $Ig,h=Tin,h[(h2−h3)/Tin,h−(S2−S3)]$

  (22)

  $Ig,air=Ten,air[(h2−h3)/Tevef,g,air−(S2−S3)]$

  (23)

  $Tevef,g,air=(Tout,w−Tin,w)/ln(Tout,w−Tin,w)$

  (24)

  where T_evef,g,air is the external fluid thermodynamic average temperature of the gas cooler of air-source heat pump water heater, which can be calculated by Eq. (24). T_in,w is the hot water inlet temperature of heat pump water heaters (290 K), and T_out,w is the hot water outlet temperature of heat pump water heaters (338 K).
• Total system irreversibility:

  $Itot,re=Icomp,re+Icap,re+Iihe,re+Ie,re+Ig,re$

  (25)

  $Itot,h=Icomp,h+Icap,h+Iihe,h+Ie,h+Ig,h$

  (26)

  $Itot,air=Icomp,air+Icap,air+Iihe,air+Ie,air+Ig,air$

  (27)
• The exergy efficiency for the overall system:

  $ηre=1−Itot,re/wcomp,re$

  (28)

  $ηh=1−Itot,h/wcomp,h$

  (29)

  $ηair=1−Itot,air/wcomp,air$

  (30)

As we all know, in the transcritical cycle, there is an optimal high pressure, which corresponds to the maximum COP of the system. For the pure R744 cycle, many scholars have done a lot of research on optimal high pressure, and the summarized formula can be used for accurate calculation. For the R744/R170 mixed refrigerant transcritical cycle, there is no mature theoretical formula to calculate the optimal high pressure. In this paper, an iterative method is used to determine the optimal high pressure of the R744/R170 transcritical cycle. The specific method is as shown in Fig. 4. First, a group of high-pressure ranges is given, in which each high pressure increases by 1 kPa, and the evaporation temperature and the outlet temperature of the gas cooler are input at the same time for continuous iteration. When the COP of the system is the maximum, the corresponding pressure is the optimal high pressure, and the COP, discharge temperature, compression ratio, refrigeration and heating capacity, power consumption, and other performance parameters are output.

The state parameters (such as enthalpy and entropy) in the calculation process are obtained by REFPROP9.0.

In order to verify the correctness of the optimal pressure calculated by the model, the optimal pressure values of pure R744 under different working conditions calculated by the above program are compared with formula (31) [36]. The results are shown in Table 8, in which the maximum error is 12.36%, and the minimum error is 1.6%; because the working condition of the correlation is different from that of this paper, even if a small part of the error is too large, it is within the acceptable range, so the correctness of the model can be verified.

Figure 5 shows the variation of the optimal discharge pressure with the outlet temperature of the gas cooler and the evaporation temperature. It can be seen from the figure that the P_opt of the system increases and decreases with the increase of the outlet temperature of the gas cooler and the evaporation temperature, respectively. Under the same conditions of the outlet temperature of the gas cooler and the evaporation temperature, the P_opt of R744/R170(25/75) system is lower than that of other refrigerants, and the maximum P_opt of R744/R170(25/75) is 2.43 MPa, 3.61 MPa, and 1.28 MPa lower at the ratio of R744/R170 (50/50), (77.6/22.4), and (100/0), respectively. The minimum P_opt of R744/R170(25/75) is 1.91 MPa, 2.96 MPa, and 1.2 MPa lower at the ratio of R744/R170(50/50), (77.6/22.4), and (100/0), respectively. In addition, it is found that the P_opt of the system increases from R744 at 50% to 77.6% and decreases from R744 at 77.6% to 100%. Therefore, it can be seen that P_opt does not increase monotonically with the increase of R744.

Figure 6 shows the change of COP_e with the outlet temperature and evaporation temperature of the gas cooler under optimal pressure. It can be seen from the figure that under the same conditions of outlet temperature and evaporation temperature of the gas cooler, the COP_e of the R744 system is higher than that of its refrigerant mixture. Compared with its refrigerant mixture, the maximum COP_e of R744 was 21.2%, 29.6%, and 22% higher at the ratio of R744/R170(25/75), (50/50), and (77.6/22.4), respectively, and the minimum COP_e of R744 was 21.8%, 27.8%, and 18.2% higher at the ratio of R744/R170(25/75), (50/50), and (77.6/22.4), respectively. At the same time, because the difference between R744/R170 (25/75) and R744/R170(77.6/22.4) is very small, it cannot be shown intuitively in the front of Fig. 6. Therefore, as shown in Fig. 7, by intercepting the detail cloud image at the back of the coordinate to observe the difference of cope, it can be seen that R744/R170(77.6/22.4) is slightly higher than R744/R170(25/75). In addition, the COP_e increases with the increase of evaporation temperature. This is because the cooling capacity of the system increases with the increase of evaporation temperature when the outlet temperature of the gas cooler remains unchanged. At the same time, when the suction pressure increases, the discharge pressure, that is, the optimal pressure, decreases with the increase of evaporation temperature, resulting in a decrease in power consumption, so the COP_e of the system increases. The reason why the COP_e decreases with the increase of the outlet temperature of the gas cooler is that the optimal discharge pressure increases with the increase of the outlet temperature of the gas cooler when the refrigerating capacity and suction pressure remain unchanged, which leads to the increase of power consumption, so the COP_e of the system decreases. At the same time, it is found that the COP_e of the system is not a monotone function, which increases with the proportion of R744.

Figure 8 shows the change of discharge temperature with the outlet temperature and evaporation temperature of the gas cooler. It can be seen from the figure that the discharge temperature of the R744 system is significantly higher than that of mixed refrigerant under the same conditions of outlet temperature and evaporation temperature of the gas cooler. With the increase of the R744 ratio, the discharge temperature process increases. The maximum discharge temperature of R744/R170(25/75) system is 9.72 ℃, 22.94 ℃, and 49.62 ℃ lower than that of R744/R170(50/50), (77.6/22.4), and (100/0), respectively. The lowest discharge temperature of R744/R170(25/75) system is 6.69 ℃, 11.29 ℃, and 20.31 ℃ lower than that of R744/R170 (50/50), (77.6/22.4), and (100/0), respectively. In addition, the discharge temperature of the system increases and decreases with the increase of the outlet temperature and evaporation temperature of the gas cooler, respectively. This is because when the outlet temperature of the gas cooler is constant, the higher the evaporation temperature is, the lower the discharge pressure will reduce the discharge temperature. When the evaporation temperature is constant, the higher the discharge pressure is, the higher the discharge temperature will lead to the increase of discharge temperature.

Figure 9 shows the change of COP_h with the outlet temperature and evaporation temperature of the gas cooler under optimal pressure. It can be seen from the figure that the overall change trend is the same as that of COP, and COP_h of R744 system is higher than that of mixed refrigerant; the maximum COP_h was 10.3%, 13.8%, and 10.8% higher at the ratio of R744/R170(25/75), (50/50), and (77.6/22.4), respectively, and the minimum COP_h was 7.5%, 9.1%, and 6.6% higher at the ratio of R744/R170(25/75), (50/50), and (77.6/22.4), respectively. Figure 10 is similar to Fig. 7, and all are detailed cloud graphs of COP_h at the back of coordinate changing with different parameters. In addition, COP_h increases with the increase of evaporation temperature, which is because the higher the evaporation temperature makes the power consumption lower, and the heating capacity decreases, which results in the reduction of COP_h. COP_h decreases with the increase of the temperature at the outlet of the gas cooling, which is because the discharge pressure and temperature increase with the increase of the outlet temperature of the gas cooler, which leads to the increase in power consumption and the reduction of heating capacity, which makes the COP_h of the system decrease.

Figure 11 shows the variation of the optimal discharge pressure with the outlet temperature and evaporation temperature of the gas cooler. It can be seen from the figure that the trend is the same as the optimal discharge pressure of the refrigeration. In addition, the P_opt of R744/R170(25/75) system is lower than that of other refrigerants, and the maximum P_opt of R744/R170(25/75) is 2.6 MPa, 3.6 MPa, and 1.35 MPa lower at the ratio of R744/R170(50/50), (77.6/22.4), and (100/0), respectively. The minimum P_opt of R744/R170(25/75) is 1.87 MPa, 2.93 MPa, and 1.23 MPa lower at the ratio of R744/R170(50/50), (77.6/22.4), and (100/0), respectively.

Figures 12 and 13 show the change in the compression ratio of the system with the evaporation temperature and the outlet temperature of the gas cooler. Because the compression ratio of the system for refrigeration and heating is basically the same, only the comparative analysis of the system for refrigeration is made to save space. It can be seen from the two figures that the compression ratio decreases with the increase of evaporation temperature and increases with the increase of outlet temperature of the gas cooler. In addition, it can be seen from Fig. 12 that the compression ratio of R744/R170 (25/75) is higher than that of R744, and the difference decreases with the increase of outlet temperature of the gas cooler, and even a small part of R744/R170 (25/75) has a lower compression ratio than R744. With the increase in outlet temperature of the gas cooler, the compression ratio of R744/R170 (77.6/22.4) is higher than that of R744. It can be found from Fig. 13 that the compression ratio of R744/R170 (25/75) is higher than that of R744, and the difference increases with the increase of evaporation temperature. The compression ratio of R744/R170 (77.6/22.4) is lower than that of R744, but with the increase of evaporation temperature, a small part of R744/R170 (77.6/22.4) has a higher compression ratio than R744.

Figure 14 shows the change in the coefficient of performance of the system used for refrigeration and heating with the proportion of R170. From the figure, we can find that the COP_e/COP_h of the system does not increase or decrease monotonously with the increase of the proportion of R170 but decreases first and increases and reaches the lowest when the proportion of R170 is 50%. This may be because the proportion of R744 dominates the COP_e/COP_h of the system when the proportion of R170 is relatively low. Therefore, with the decrease of R744, the COP_e/COP_h of the system also decreases. When the proportion of R170 is high, the proportion of R170 dominates the COP_e/COP_h of the system. Therefore, with the increase of R170, the COP_e/COP_h of the system increases.

Figure 15 shows the variation of the optimal pressure of the system for refrigeration and heating with the proportion of R170; Figs. 15(a) and 15(b) are the optimal pressures in the refrigeration system, and Figs. 15(c) and 15(d) are the optimal pressures in the heating system. It can be seen from the figure that the optimal pressure values are similar in both refrigeration and heating systems, and they first increase and then decrease with the increase of the R170 proportion. The highest pressure occurs when the proportion of R170 is 22.4%. At this moment, the mixture is azeotropic refrigerant. When the proportion of R170 is 75%, the optimal pressure is the lowest.

Figure 16 shows the variation of the compression ratio of the system used for refrigeration and heating with R170; Figs. 16(a) and 16(b) are the compression ratios in the refrigeration system, and Figs. 16(c) and 16(d) are the compression ratios in the heating system. It can be seen from Figs. 16(a) and 16(c) that when the outlet temperature of gas cooling is 308 k, the compression ratio in the refrigeration system increases with the increase of R170 in the mixture, but the increase is very small. The compression ratio in the heating system first increases and then decreases with the increase of R170 in the mixture and reaches the maximum when R170 accounts for 50%. It can be seen from Figs. 16(b) and 16(d) that the change trend of compression ratio with R170 proportion in refrigeration and heating systems is the same. It is worth noting that when the evaporation temperature is 245 k, the compression ratio of R744 is lower than that of other mixed refrigerants. With the increase in outlet temperature of the gas cooler, the compression ratio of R744 gradually exceeds that of R744/R170(77.6/22.4). The compression ratio of R744/R170(25/75) with the increase of outlet temperature of the gas cooler, the difference between R744/R170(50/50) decreases gradually, even exceeding its compression ratio. In addition, when the system is used for refrigeration and heating, although the compression ratio fluctuates with the change of R170 proportion, the overall analysis from the perspective of R170 proportion shows that the compression ratio of each working medium has little difference, so the influence of R170 proportion on the compression ratio of the system is not significant. Figure 17 shows the change in discharge temperature with an R170 proportion. It can be seen from the figure that the discharge temperature decreases with the increase of the R170 proportion, which is different from COP and optimal pressure.

Figures 18(a)–18(c), respectively, show the variation trend of exergy efficiency of refrigeration, heating, and air-source heat pump systems with R170 proportion. Exergy efficiency can reflect the irreversible loss of the system in the process of energy transfer from the perspective of the second law of thermodynamics. It can be seen from the figure that in the three systems, the exergy efficiency first decreases and increases with the increase of the R170 proportion and reaches the lowest when the R170 proportion is 50%, and the R744 system has the highest exergy efficiency. At the same time, the exergy efficiency of the heating system is the highest among the three systems, which indicates that the working fluids are suitable for indoor heating in a low external environment. In addition, exergy efficiency increases with the increase of evaporation temperature and decreases with the increase of outlet temperature of the gas cooler.

In this paper, the performance of R744 and R744/R170 mixed refrigerants in refrigeration and heating systems is studied by simulation method, and the changes of COP, optimal discharge pressure, discharge temperature, compression ratio, and other system energy parameters with evaporation temperature, gas cooler outlet temperature, and R170 proportion are analyzed. The main conclusions are as follows:

1. In heating systems, P_opt increases and decreases with the increase of gas cooler outlet temperature and evaporation temperature.

2. The P_opt of R744/R170 (25/75) system is lower than that of other refrigerants.

3. The P_opt first increases and then decreases with the proportion of R170.

4. COP_h increases with the increase of evaporation temperature and decreases with the increase of outlet temperature of the gas cooler. Meanwhile, COP_e/COP_h first decreases and then increases with the proportion of R170.

5. The discharge temperature of R744/R170 (25/75) system is the lowest. The discharge temperature increases and decreases with the increase of outlet temperature and evaporation temperature of the gas cooler and increases with the increase of R744 proportion.

6. The influence of the R170 proportion on the compression ratio of the system is not big.

7. The exergy efficiency of R744/R170 at the ratio of 25/75 is higher than that of other refrigerants.

Through the study of this paper, if the refrigerant is selected to replace R744, the comprehensive performance of each working medium in the system should be considered. In conclusion, R744/R170 is the best choice under the ratio of 25/75.

There are no conflicts of interest. This article does not include research in which human participants were involved. Informed consent not applicable. This article does not include any research in which animal participants were involved.

No data, models, or code were generated or used for this paper.

h =

specific enthalpy of state point, kJ/kg

q =

refrigerating/heating capacity per unit, kJ/kg

s =

specific entropy, kJ/(kg·K)

I =

irreversibility per unit, kJ/kg

P =

Pressure, MPa

T =

temperature, K

X =

R170 mass fractions

β =

efficiency of internal heat exchanger

$ε$ =

efficiency of compressor

η =

exergy efficiency

air =

air-source heat pump water heater

cap =

capillary

comp =

compressor

e =

evaporation

eq =

calculational

en =

environment

evef =

external fluid thermodynamic average

g =

condensation

h =

heating

is =

isentropic

ihe =

internal heat exchanger

in,w =

Inlet temperature of water

out,w =

out temperature of water

opt =

optimum

re =

refrigeration

sim =

simulation

tot =

total

0, 1, 2, 2s, 3, 4, 5 =

state point

COP =

coefficient of performance

     GWP =

global warming potential

ODP =

ozone depleting potential