1 Introduction

The use of both hydrochlorofluorocarbons (HCFCs) and hydrofluorocarbons (HFCs) in the world will gradually enter the period of reduction according to Montreal Protocol and Kigali Amendment [1, 2]. Therefore, natural refrigerants with zero ozone depletion potential (ODP) and low global warming potential (GWP), such as R290, have been considered as substitutes for HCFCs and HFCs in household air conditioners. With the support and guidance of relevant policies from various countries, the relationship between compressors and mainstream refrigerants will be increasingly close. Therefore, the development of compressor suitable for new refrigerants is also an important research direction.

Much research has been carried out to promote the market application of R290. Zhang et al. [3] analyzed the feasibility of R290 replacing R22 by comparing the thermo-physical properties of R290 and R22. The results are as follows. The standard boiling point, freezing point, critical temperature and other parameters of R290 are very close to R22. The mass charge of R290 is less than that of R22 in the same volume. Under the same evaporator load, both the suction and exhaust temperatures of R290 are lower than that of R22, which can decrease the irreversible losses by weakening the heat exchange between fluid and cylinders. The saturated vapor dynamic viscosity and saturated liquid dynamic viscosity of R290 are smaller than these of R22, which can reduce the friction loss between R290 and pipe wall. The latent heat of vaporization of R290 is 1.84 times that of R22, and the thermal conductivity of R290 is 1.75 times that of R22, which is helpful to reduce the size of heat exchanger and improve the overall performance of air conditioning system.

In the early phase, many scholars have done a lot of research on R290 from two aspects: phase change heat transfer [4,5,6,7] and expansion valve [8,9,10]. In terms of compressor, Zou et al. [11] charged R290 for linear compressor instead of traditional reciprocating compressor and conducted performance test and power consumption analysis of linear compressor. The results show that in order to improve the energy efficiency of linear compressor, it is necessary to optimize the performance coefficient of linear motor and reduce the fit clearance between piston and cylinder at the same time. Yuan et al. [12] compared the thermodynamic performance of R290, R32 and R22 by establishing a thermodynamic simulation model of the rolling piston compressor, which shows that the increase in suction superheat has a great impact on the R290 compressor. Wu et al. [13, 14] made experimental studies on the starting characteristics of the rolling piston compressor under the refrigeration and heating mode. The results show that the mixing viscosity of oil and refrigerant and the oil level of the compressor oil pan are within reasonable range after the R290 system is started, which ensures the stable start-up of air conditioning system. Pilla et al. [15] evaluated the compressor performance of the mixed refrigerant R290/R600a and estimated the temperature distribution during the operation. The results show that 60% R290 and 40% R600a have the best system performance. Chen et al. [16] tested the friction coefficient, bite force and wear amount of the sliding plate piston friction pair under the R290/mineral oil mixture combination under the sealed high-pressure environment and compared it with the R410A/POE oil mixture combination which is widely used at present. The results show that under the combination of R290/mineral oil mixture, the anti-wear ability of the compressor is enhanced, but the friction power consumption is also increased, which needs further improvement. Sotomayor et al. [17] established a simulation model of open piston compressor for automotive air conditioning based on the existing experimental data and simulated the operation of R1234yf and R290 compressors based on the simulation model of R134a, and the predicted values were in good agreement with the experimental values, which proved the applicability of the model to the study of new refrigerants. Cai et al. [18] simulated and analyzed the leakage characteristics of R290 rolling piston compressor and compared the leakage characteristics of R22 and R410A. The results show that compared with R22 and R410A compressors under the same conditions, R290 compressor needs smaller radial clearance to obtain higher efficiency. The reduction of mass charge must be accompanied by the decrease in lubricating oil. The oil-free compressor for solving the problem of compressor oil supply under low oil level has also been developed [19].

In order to further promote the use of R290 in room air conditioner, it is necessary to study the rolling piston compressor for R290. At present, there are few experimental studies on the off-design characteristics of R290 rolling piston compressor. The purpose of this study is to investigate the influence of multiple operating conditions on the performance parameters of the compressor. Combined with the properties of R290 and rolling piston compressor, the experimental results are analyzed, which provides thermodynamic reference for the design and improvement of R290 rolling piston compressor and promotes the efficient utilization of R290. Through the performance test under variable operating conditions, it provides reference for the improvement of rolling piston compressors using R290 or other flammable refrigerants and promotes the commercialization of R290 rolling piston compressor.

2 Experimental Apparatuses and Operating Conditions

2.1 Descriptions of the Test Platform

In order to study the off-design characteristics of R290 rolling piston compressor, the test platform based on the secondary refrigerant calorimeter method [20] was set up. Figure 1 is the schematic diagram of system. After being compressed by the compressor, R290 (vapor) with high temperature and high pressure (exhaust pressure Pc) enters the oil separator, and the lubricating oil is separated. Then, R290 (vapor) enters the condenser for condensation and becomes liquid with medium temperature and high pressure. R290 (liquid) passes through the dryer and the visual liquid lens in turn and then enters the subcooler for subcooling. The supercooled R290 (liquid, responding to the temperature after subcooling tasc and subcooling temperature tsc) flows through the filter and then enters the electronic expansive valve. After being cooled and depressurized, it becomes low-temperature (evaporating temperature te) and low-pressure (suction pressure Pe) liquid. Then, it absorbs the heat of the secondary refrigerant in the calorimeter and turns into low-temperature and low-pressure R290 (vapor). Finally, it flows into the compressor (corresponding to the suction temperature ts) for circulation.

Fig. 1
figure 1

The schematic diagram of system (1) Calorimeter. (2) Electronic expansive valve (EEV). (3) Filter. (4) Subcooler (water thermostat system). (5) Chiller for subcooler. (6) Electric heater for subcooler. (7) Visual liquid lens. (8) Dryer. (9) Condenser (water thermostat system). (10) Chiller for condenser. (11) Electric heater for condenser. (12) Oil separator. (13) Compressor. (14) Compressor environmental control room. (15) Electric heater for calorimeter. (16)–(19) Temperature sensor. (20)–(23) Pressure sensor

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The schematic diagram of the measurement and control for the test platform is shown in Fig. 2. The experimental parameters are adjusted by the refrigerating cycle system (e.g., Fig. 1) and the electrical measurement and control system (e.g., Fig. 2). For the monitoring and control of ten, Pe, ts, Pc and other important parameters, special PID instrument, actuators and drivers like step driver or silicon-controlled rectifier are used. For the control of other electrical components, programmable logic controller (PLC) is used for analog and digital control. To set the Pe (te) of the compressor, the opening of the electronic expansion valve (EEV) can be controlled by adjusting the output pulse of the special step driver, thus adjusting the mass flow of R290 and stabilize the Pe at the set value. The ts of the compressor is controlled by monitoring the temperature of R290 at the outlet of the calorimeter (evaporator) in the refrigerating cycle system to control heat input by adjusting the power of electric heater in the calorimeter. While ensuring that the ts reaches the set value, the sum of the heating amount of the electric heater and the heat leakage, and the heat absorption capacity of the evaporating coil achieve equilibrium. For the control of Pc, the condensing coil of the tested compressor system, an electric heater and a group of cooling water coils from another refrigerating cycle system are installed in a water thermostat system. The device achieves dynamic balance by controlling the electric heating capacity, the heat released by the cooling coil of another refrigerating cycle system and the condensing coil of the tested compressor system to maintain the constant Pc of the tested compressor. The control system of tsc is basically the same as that of Pc control system, except that it monitors the tsc. The PID control instrument receives the signal from the compressor ambient temperature (ten) sensor and adjusts the operating frequency of the electric heater according to the error between the environmental setting value and the actual value, thus balancing the electric heating capacity and the cooling capacity of the fan coil air conditioning system. The voltage Vcc and frequency f of the compressor are controlled by the regulated power supply. Some known parameters of the tested compressor are shown in Table 1. The maximum allowable deviation between the measured value and the set value δmax and the allowable deviation of the measured value relative to the average value δavg of the operating parameters in the experiment are shown in Table 2.

Fig. 2
figure 2

Measurement and control schematic diagram of compressor test platform

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Table 1 Some known parameters of the rolling piston compressor to be tested
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Table 2 Parameter error allowable settings
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In addition, considering the flammability of R290, the corresponding adjustment and design of R290 rolling piston compressor are also made, such as the electronic joint of fireproofing design, which makes the compressor run more reliable; the special overload specially developed to avoid the danger of use; the special internal materials with good compatibility are used to effectively prevent aging, etc.

2.2 Descriptions of Operating Conditions

By controlling the parameters of operating conditions, the test conditions of compressor can be determined. The standard test conditions of compressor specified in GB/T 15765-2014 [21] are shown in Table 3. Based on the standard experimental conditions, in order to further study the off-design characteristics of R290 rolling piston compressor, the experimental conditions shown in Table 4 are designed by using the variable-controlling method. Conditions of No. 1–7 are variable suction temperature tests, which is called Vts; No. 8–28 are variable compression ratio tests called Vπ, in which No. 8–18 are variable compression ratio (variable evaporating pressure) tests called VPe and No. 19–28 are variable compression ratio (variable condensing pressure) tests called VPc. In addition, variable compression ratio conditions can be divided into two types: suction temperature of 18.3 ℃ and suction temperature of 35 ℃. As a measuring parameter of the system, tsc is not the test condition of the compressor and has no effect on several performance parameters, such as the volumetric efficiency of the compressor. Therefore, in this experiment, tsc is set to be 8.3 ℃ in all conditions, and the f of the compressor is 50 Hz.

Table 3 Compressor standard test conditions
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Table 4 Experimental conditions
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3 Data Processing

Electric power W, coefficient of performance COP, exhaust temperature td can be measured or calculated indirectly from the experiments. The measurement and calculation methods of the above parameters refer to the existing test standards and related literature of compressor [13, 14, 20, 21]. Cooling capacity Q is calculated by Eq. (1)

$$Q = \frac{{Q_{1} + Q_{2} }}{2}$$
(1)

where Q1 is the sum of the electricity input to the electric heater and the heat leakage of the calorimeter, and Q2 is the product of the actual mass flow rate and enthalpy difference of R290 at the inlet and outlet of the calorimeter. The detailed calculation methods for both are given in the references [20].

Volumetric efficiency ηV is a parameter describing the utilization degree of working volume of compressor cylinder, reflecting the volume loss caused by clearance volume, suction resistance, suction heating, vapor leakage and suction reflux [22], which can be calculated by Eq. (2)

$$\eta_{{\text{V}}} = \lambda_{{\text{V}}} \lambda_{{\text{p}}} \lambda_{{\text{T}}} \lambda_{{\text{l}}} \lambda_{{\text{h}}}$$
(2)

where λV is the volume coefficient, it can be calculated by empirical Eq. (3)

$$\lambda_{{\text{V}}} = 1 - c\left[ {\left( \pi \right)^{{\frac{1}{\kappa }}} - 1} \right]$$
(3)

where c is the relative clearance volume, and κ is the expansion adiabatic index.

λP is the pressure coefficient, which indicates the influence of relative pressure loss on volumetric efficiency at the end of suction, it can be considered as 1 because the rolling piston compressor has no suction valve; λT is the temperature coefficient, which reflects the reduction of the vapor transmission volume caused by the heating of the gas inhaled by the compressor. Its size is equal to the ratio of the density outside the suction port and in the cylinder at the end of the suction; λl is the leakage coefficient, which is related to the compression ratio, compressor structure, lubricating oil, etc; λh is the reflux coefficient, which is approximately equal to 1 because of the small edge angle [22].

From the experimental point of view, ηV can also be calculated by Eq. (4)

$$\eta_{{\text{V}}} = q_{{{\text{va}}}} /q_{{{\text{vt}}}} = v_{{{\text{suc}}}} q_{{\text{m}}} /(V_{0} f)$$
(4)

where qm is the tested inspiratory mass flow of R290 compressor [kg s−1]; V0 is the theoretical displacement of the compressor [m3]; f is the frequency of the compressor [Hz]; vsuc is the specific volume of R290 at the suction port of the compressor [m3 kg−1]; qva is the actual volume flow of R290 compressor [m3 s−1], while qvt is the theoretical volume flow of R290 compressor [m3 s−1].

Electric efficiency ηel is a parameter that characterizes the perfection degree of the input work of the motor used in the compressor, which is the product of indicative efficiency ηi, heating efficiency ηt, leakage efficiency ηl, mechanical efficiency ηm and motor efficiency ηmo. When it comes to indicated efficiency ηi, a mathematical analysis model is proposed in reference [22], which can be calculated by Eq. (5)

$$\eta_{i} = \frac{{\lambda_{{\text{T}}} \lambda_{{\text{l}}} }}{{1 - \frac{{1.5(\Delta p_{{{\text{sm}}}} + \Delta p_{{{\text{dm}}}} \pi^{1/\kappa } )}}{{(h_{{{\text{dis}}}} - h_{{{\text{suc}}}} )/v_{{{\text{suc}}}} }}}}$$
(5)

where \(\Delta\)psm and \(\Delta\)pdm are the average pressure drop of suction and exhaust valves, respectively, \(\Delta\)psm can be ignored because there is no suction valve in rolling piston compressor; hsuc and hdis are the specific enthalpies of suction and exhaust of compressor [kJ kg−1]; π is the compression ratio; κ is the adiabatic index of R290.

Based on the experiment, the electric efficiency ηel is usually calculated by Eq. (6)

$$\eta_{{{\text{el}}}} = q_{{\text{m}}} \left( {h_{{\text{dis,is}}} - h_{{{\text{suc}}}} } \right)/W$$
(6)

where hsuc is the specific enthalpy of R290 at the suction port of the compressor [kJ kg−1]; hdis,is is the specific enthalpy of isentropic compressed exhaust [kJ kg−1].

Comprehensive efficiency coefficient ηcom is the cycle efficiency, also known as the perfection degree of thermodynamics. It is usually calculated by Eq. (7)

$$\eta_{{{\text{com}}}} = \frac{{{\text{COP}}}}{{{\text{COP}}_{0} }} = \frac{{{\text{COP}}}}{{\left( {h_{{{\text{suc}}}} - h_{{{\text{el}}}} } \right)/\left( {h_{{\text{dis,is}}} - h_{{{\text{suc}}}} } \right)}}$$
(7)

where COP0 is the theoretical coefficient of performance; hel is the specific enthalpy at the inlet of evaporator [kJ kg−1].

4 Results

4.1 Results at Vts

The superheat of this experiment is controlled by adjusting the power of the electric heater in the calorimeter, and the W is obtained by measuring the electric heating power, so the overheating belongs to useful superheat [23]. By consulting REFPROP version 10.0 [24], the unit mass cooling capacity (theo-q) and theoretical work per unit mass (theo-w) under the theoretical cycle are calculated, as shown in Fig. 3. It is worth mentioning that the pipeline resistance, suction and exhaust resistance and other factors are ignored in the theoretical cycle. When the ts increases from 15 to 41 ℃, theo-q increases from 272 to 320 kJ/kg by 17.5%, while theo-w increases from 56.7 to 64.6 kJ/kg by 13.9%. Compared with the theoretical system, the changes of Q and W of R290 rolling piston compressor at Vts in the actual system are shown in Fig. 4. With the increase of ts, Q increased from 3356 to 3637 W, an increase of 8.4%. It can be concluded that both Q and theo-q benefit from useful superheat. However, considering that with the increase of ts, vsuc will increase, resulting in the decrease of qm, so with the increase of ts (especially 22 ℃), the growth rate of Q gradually slows down and lower than that of theo-q. Different from theo-w, W decreased from 1075 to 1061 W by 1.26%. The value of W is the product of qm and w. Obviously, in this experiment, the effect of the decrease of qm on W is greater than that of the increase of w, resulting in a small decrease of W. In conclusion, the Q increases, while the W decreases with the increase of ts, and they both show a trend of slowing down when the ts reaches 22 ℃.

Fig. 3
figure 3

Changes of theo-q, theo-w at Vts

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Fig. 4
figure 4

Changes of Q, W at Vts

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In the theoretical system (in Fig. 3), the growth rate of theo-w is slightly lower than that of theo-q, which also leads to the increase of theo-COP by 3.22%. As shown in Fig. 5, with the increase of ts, COP increases from 3.13 to 3.39 by 8.3%. However, the increase rate will gradually slow down when the ts reaches 22 ℃. The reason for it is that the td also increases from 74.8 to 95.2 ℃ by 27.3% with the increase of ts. Although in theory, the td of R290 is lower than some common refrigerants (like R22) at household air conditioning conditions, the td of the last two groups of experiments is above 90 ℃, which will have adverse effects on the operation of the system. If td is too high, it may bring about the decrease in viscosity of lubricating oil, cause poor lubrication, accelerate wear and increase power consumption. Meanwhile, high td will also reduce the cooling effect of exhaust on the motor. Therefore, when it comes to using R290 that is flammable in a refrigerating cycle system, we should try our best to avoid operating the system under the conditions with high td.

Fig. 5
figure 5

Changes of COP, td at Vts

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Figure 6 shows the trend of efficiency η at Vts. When ts increases, ηV increases slowly in the range of 0.86–0.87. The reasons for the changes are as follows. According to Eq. (2), for rolling piston compressor, λP and λh are approximately equal to 1, whose effect on ηV is negligible. According to Eq. (3), when π is constant, λV causes little effect on ηV. For λT, the vsuc increases after being heated by the cylinder. Therefore, at the beginning of compression, the ratio of refrigerant density outside the suction port to that in the cylinder decreases at the end of compression suction, and λT decreases. However, with the increase of ts, the temperature difference between the outside of the suction port and the inside of the cylinder decreases at the end of the compression suction, that is, the ratio of density increases and λT increases; For λl, R290 is in the vapor state when it is superheated, so there is no phenomenon that the liquid refrigerant dilutes the lubricating oil. Besides, the compression ratio remains unchanged, which basically has no effect on the refrigerant leakage of R290. Therefore, λl causes negligible effect on ηV. In general, ηV is mainly affected by λT and increases slightly. Obviously, the effective way to increase ηV is to increase ts properly.

Fig. 6
figure 6

Changes of \(\eta\) at Vts

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As shown in Fig. 6, ηel increases from 0.672 to 0.694, with an increase of 3.2%. The reasons can be analyzed from both theoretical formula and experimental results. First, when π is constant, with the increase of ts, λl almost remains unchanged, while λT increases. Second, according to the experimental results and Eq. (5), the influence of the increase of λT and vsuc on ηi is greater than that of w on ηi, so ηel is increasing continuously. However, due to the increase of td, the cooling effect of exhaust on the motor is reduced, and the growth trend of ηel is gradually slowing down. With the increase of ts, ηcom increases from 0.655 to 0.691 by 5.6%. For the R290 rolling piston compressor system, the useful superheat promotes the growth of COP and COP0. The increase in the ratio of the two indicates that the increment of COP is greater than that of COP0. All in all, η increases with the increase of ts and shows a trend of slowing down when the ts reaches 22 ℃.

4.2 Results at Vπ

Similar to Vts, theo-q and theo-w under Vπ were also studied theoretically. The variation of theo-q and theo-w under Vπ is shown in Figs. 7 and 8, respectively. With the decrease of Pe from 0.673 MPa to 0.519 MPa, theo-q at ts of 18.3 ℃ increases from 274.7 to 280.6 kJ/kg by 2.1%, while theo-w at ts of 18.3 ℃ increases sharply from 49.6 to 65.47 kJ/kg by 32.1%. By contrast, with the increase of Pc from 1.569 to 2.117 MPa, theo-q at ts of 18.3 ℃ decreases from 302.4 to 261.2 kJ/kg by 13.6%, while theo-w at ts of 18.3 ℃ increases sharply from 54.2 to 70.9 kJ/kg by 30.3%. It can be concluded that the effect of the increase of Pc on theo-q is greater than that of the decrease of Pe, which attributes to the increase in theoretical enthalpy after subcooling (theo-hasc) when the Pc is increasing and the tsc is constant. In addition, the growth trends of theo-q and theo-w at ts of 35 ℃ are similar to those at ts of 18.3 ℃.

Fig. 7
figure 7

Changes of theo-q at Vπ

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Fig. 8
figure 8

Changes of theo-w at Vπ

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Based on the analysis of the theoretical system, the changes of Q and W of tested R290 compressor at Vπ are shown in Figs. 9 and 10. With the decrease of Pe from 0.673 to 0.519 MPa, Q at ts of 18.3 ℃ (35 ℃) decreases from 4124 (4180 W) to 2981 W (3098 W) by 27.7% (25.9%), while W at ts of 18.3 ℃ (35 ℃) decreases slightly from 1060 (1061 W) to 1035 W (1038 W) by 2.4% (2.2%). With the increase of Pc from 1.569 to 2.117 MPa, Q at ts of 18.3 ℃ (35 ℃) decreases from 3881 (4129 W) to 3005 W (3312 W) by 22.6% (19.8%), while W at ts of 18.3 ℃ (35 ℃) increases from 885 (902 W) to 1224 W (1226 W) by 38.3% (35.9%). Obviously, different from theo-q at VPe, Q at Vπ shows a downward trend with the increase of π, and the variation range of Q is quite larger. The reason for it is that when the ts is constant and Pe is decreasing, vsuc is constantly increasing, greatly reducing the qm and Q. Besides, different from theo-w at Vπ, with the increase of π, W at VPe decreases slightly, while W at VPc increases sharply, which indicates that compared with the decrease of Pe, the effect of the increase of Pc on W is much greater.

Fig. 9
figure 9

Changes of Q at Vπ

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Fig. 10
figure 10

Changes of W at Vπ

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Figures 11 and 12 show the variation of COP and td of R290 rolling piston compressor system at Vπ, respectively. With the decrease of Pe, COP at ts of 18.3 ℃ (35 ℃) decreases from 3.76 (3.86) to 2.78 (2.87) by 26.1% (25.7%). With the increase of Pc, COP at ts of 18.3 ℃ (35 ℃) decreases from 3.98 (4.48) to 2.44 (2.73) by 38.7% (39.1%). When the ts is 18.3 ℃, the decrease rate of COP at VPc is 48.5% higher than that at VPe, and the decrease rate is 52.3% higher when the ts is 35 ℃, which mainly attributes to the difference of variation trend for W between VPe (decreased by about 2%) and VPc (increased by over 35%). On the other hands, with the decrease of Pe, td at ts of 18.3 ℃ (35 ℃) increases from 71 (85.1) to 80.2 (94.8) by 13.0% (11.4%). With the increase of Pc, td at ts of 18.3 ℃ (35 ℃) increases from 66.3 (80.2) to 85.8 (95.3) by 29.4% (18.8%). It can be concluded that the decline rate of td at VPc is 127.0% (95.2%) higher than that at VPe at ts of 18.3 ℃ (35 ℃), and the increase of td is unfavorable to the operation of the system.

Fig. 11
figure 11

Changes of COP at Vπ

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Fig. 12
figure 12

Changes of td at Vπ

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Figures 13 and 14 show the changes of η of R290 rolling piston compressor at Vπ. With the decrease of Pe, ηV at ts of 18.3 ℃ (35 ℃) decreases from 0.872 (0.878) to 0.848 (0.854) by 2.8% (2.7%). With the increase of Pc, ηV at ts of 18.3 ℃ (35 ℃) decreases from 0.887 (0.891) to 0.831 (0.843) by 6.3% (5.4%). The reasons for the changing trend of ηV are as follows. According to Eq. (3), λV increases with the increase of π. For λT, the vsuc increases after being heated by the cylinder. Therefore, at the beginning of compression, the ratio of density (R290) outside the suction port to that in the cylinder decreases at the end of compression suction, and λT decreases. However, when ts is higher (35 ℃ compared to 18.3 ℃), the temperature difference between the outside of the suction port and the inside of the cylinder decreases at the end of the compression suction, that is, the ratio of density increases and λT increases slightly, which is the reason why the downward trend of ηV at ts of 35 ℃ is lower than that of 18.3 ℃. For λl, R290 is vapor when it is overheated, so there is no phenomenon that the liquid refrigerant dilutes the lubricating oil. Although the leakage mainly depends on the structural clearance of the compressor, with the increase of π, λl will also decrease. In general, ηV decreases slightly with the increase of π.

Fig. 13
figure 13

Changes of \(\eta\) at VPe

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Fig. 14
figure 14

Changes of \(\eta\) at VPc

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With the decrease of Pe, ηel at ts of 18.3 ℃ (35 ℃) decreases from 0.688 (0.695) to 0.672 (0.678) by 2.3% (2.5%). With the increase of Pc, ηel at ts of 18.3 ℃ (35 ℃) decreases from 0.703 (0.712) to 0.640 (0.667) by 9.0% (6.3%). The reasons for the variation of ηel are as follows: The increase of π may increase the clearance value of the moving components of the compressor, reducing the mechanical efficiency; the increase of π will increase the td, weaking the cooling effect of the exhaust on the motor, thus reducing the motor efficiency. At VPe or VPc, the slope of the change curve of ηel under the condition of ts of 35 ℃ is similar to that of 18.3 ℃, and the ηel of the former is greater than that of the latter (about 1.2% higher on average at VPe; 2.46% higher on average at VPc). In addition, it can be concluded that compared with the decrease of Pe, the increase of Pc has a greater impact on the decrease of ηel. With the decrease of Pe, ηcom at ts of 18.3 ℃ (35 ℃) decreases from 0.682 (0.692) to 0.662 (0.672) by 2.9% (2.9%). With the increase of Pc, ηcom at ts of 18.3 ℃ (35 ℃) decreases from 0.692 (0.702) to 0.628 (0.662) by 9.3% (5.7%). According to Eq. (7), the decrease of ηcom mainly attributes to the significant decrease of COP, and the decreasing trend of COP was greater than that of COP0. In addition, the decline rate of ηcom at VPc is about 3 (2) times that at VPe at ts of 18.3 ℃ (35 ℃). Therefore, avoiding the continuous fluctuation of Pc is an important measure to ensure the stability of ηcom.

5 Conclusions

In this paper, the test platform of R290 rolling piston compressor was set up. The effects of ts, π (Pe, Pc) on COP, Q, W, td, ηV, ηel and ηcom of R290 rolling piston compressor were studied. Besides, the change laws were analyzed, and the following conclusions could be made from the experiments.

With the increase of ts (from 15 to 41 ℃), Q increases from 3356 to 3637 W by 8.4%, while W decreases from 1075 to 1061 W by 1.26%, which attributes to the useful superheat and the decrease of qm. Based on Q and W, COP increases from 3.13 to 3.39 by 8.31% although the increase rate will gradually slow down when the ts reaches 22 ℃ because of the increase of td (increasing from 74.8 to 95.2 ℃ by 27.3%). For η, ηV increases slowly in the range of 0.86–0.87 (mainly affected by λT); ηel increases from 0.672 to 0.694 by 3.2%; ηcom increases from 0.655 to 0.691 by 5.6%, which shows the improvement of performance at Vts (limited to the temperature range studied in this paper). In general, increasing ts properly is an effective means to improve the performance of R290 rolling piston compressor.

With the decrease of Pe from 0.673 to 0.519 MPa, Q at ts of 18.3 ℃ (35 ℃) decreases by 27.7% (25.9%), while W at ts of 18.3 ℃ (35 ℃) decreases slightly by 2.4% (2.2%). With the increase of Pc from 1.569 to 2.117 MPa, Q at ts of 18.3 ℃ (35 ℃) decreases by 22.6% (19.8%), while W at ts of 18.3 ℃ (35 ℃) increases by 38.3% (35.9%), which shows that the downward trends of W at VPe and VPc are quite different. Meanwhile, the decline rate of td at VPc is 127.0% (95.2%) higher than that at VPe at ts of 18.3 ℃ (35 ℃). For η, the downward trend of ηV at ts of 35 ℃ is lower than that of 18.3 ℃ (mainly affected by λV and λT); ηel decreases with the increase of π slightly (mainly depends on clearance value and the cooling effect of the exhaust on the motor); the decline rate of ηcom at VPc is about 3 (2) times that at VPe at ts of 18.3 ℃ (35 ℃), which indicates that the frequent fluctuation of Pc will lead to the significant decrease of ηcom.

The construction of the experimental platform, the design of the experimental scheme and thermodynamic analysis of the changing trend of each variable provide thermodynamic and experimental references for the improvement of rolling piston compressor using R290 or other flammable refrigerants, promoting the commercialization of R290 rolling piston compressor. With that, the optimization of the suction and exhaust system of R290 rolling piston compressor is worth further research, which is closely related to the cooling capacity, electric power, volumetric efficiency, and the resistance in the flow process. The related structural improvement direction can focus on the thickness of the exhaust valve seat, the diameter of the exhaust hole, the height of the cylinder, etc. Based on these structural parameters, it is a good research method to use the simulation software to get the object with better performance and then carry out the whole machine experiment to verify the performance (such as cooling capacity and electric power) of the improved compressor.