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Experimental investigation of startup characteristics of R290 rotary compressor under low ambient temperature heating condition
Accepted Manuscript Title: Experimental investigation of startup characteristics of R290 rotary compressor under low ambient temperature heating condition Author: Jie Lin, Jianhua Wu, Ze Zhang, Zhenhua Chen, Jing Xie, Jun Lu PII: DOI: Reference:
S0140-7007(17)30092-0 http://dx.doi.org/doi: 10.1016/j.ijrefrig.2017.02.029 JIJR 3573
To appear in:
International Journal of Refrigeration
Received date: Revised date: Accepted date:
29-4-2016 21-2-2017 24-2-2017
Please cite this article as: Jie Lin, Jianhua Wu, Ze Zhang, Zhenhua Chen, Jing Xie, Jun Lu, Experimental investigation of startup characteristics of R290 rotary compressor under low ambient temperature heating condition, International Journal of Refrigeration (2017), http://dx.doi.org/doi: 10.1016/j.ijrefrig.2017.02.029. This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting proof before it is published in its final form. Please note that during the production process errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.
Experimental investigation of startup characteristics of R290 rotary compressor under low ambient temperature heating condition Jie Lin a ,Jianhua Wu a, *, Ze Zhang a, Zhenhua Chen a, b, Jing Xieb, Jun Lub a School of Energy and Power Engineering, Xi’an Jiaotong University, Xi’an 710049, PR China b Guangdong Meizhi Compressor Co. Ltd, Guangdong 528333, PR China *
Corresponding author. Tel.: +82 029 8266 3786, +82 13689206050. E-mail address: [email protected] (Jianhua Wu). Postal address: No. 28, Xian-ning West Rd., Xi’an City 710049, Shaanxi, PR China. Highlights
Startup characteristics of a rotary compressor in the R290 heat pump system were experimentally investigated Pressures and temperatures in the R290 heat pump system during cold and warm startup processes were measured. Temperatures in the rotary compressor shell and the pressures in cylinder during two startup processes were measured
Abstract: The cold and warm startup characteristics of an R290 heat pump system and its rotary compressor were experimentally investigated under a low ambient temperature heating condition. These startup characteristics included the temperatures in the compressor shell and system and the pressures in the cylinder and system. The results showed that the minimum suction gauge pressure (-48 kPa) during a cold startup was lower than that during a warm startup (155 kPa) under a low ambient temperature heating condition. In addition, the time required for the R290 heat pump system to reach a steady pressure was much longer than that of a system using R410A as the refrigerant. Compared with that of the cooling condition, smaller amounts of liquid were generated in the cylinder at the beginnings of both the cold and warm startup processes under the low ambient temperature heating condition. Keywords：R290; Experimental Investigation; Rotary Compressor; Heating Condition; Startup 1. Introduction Air conditioners, heat pumps, and other commercial refrigeration equipment are widely used in contemporary society. On–off cycling operation is a common way to control the capacity of indoor heating and cooling equipment. The startup of a heat pump is a complicated transient process, in which the refrigerant mass, heat load, torque, and rotational speed of the compressor vary. Thus, the startup process for a heat pump has an important effect on the reliability and efficiency of the system. Murphy and Goldschmidt (1984, 1985, and 1986) conducted several studies on the startup 1 Page 1 of 13
dynamics of a 3 ton split-type air conditioning system. A steady-state correlation was used for the compressor startup in the startup model. Tanaka et al. (1982) experimentally studied the dynamic characteristics of a heat pump during the startup and shutdown processes. The refrigerant mass distributions in various components, including the evaporator, condenser, and rotary compressor, were measured in their experiment. Based on their results, they made some useful technical suggestions to improve the startup performance. The Montreal Protocol was enacted in 1987 and included the decision to phase out R22 because it causes environmental problems (Protocol, 1987). This began a new period in transient characteristics research, and many alternatives to R22 have appeared, such as R134a, R32, and R290. Haberschill et al. (2003), Judge et al. (1996), and Sami and Dahmani (1996) presented a series of dynamic simulations of air conditioning or heat pump systems with different kinds of refrigerants. Fukuda and Hayano (1996) experimentally evaluated the lubricating characteristics of halogenated fluorocarbon/polyolester (HFC/POE) pairs consisting of R407C, R410A, and two kinds of POEs on the bearings of a rotary compressor in comparison with an R22 and mineral oil (MO) combination during the startup process. Judge, Hwang, and Radermacher (1996) experimentally and theoretically investigated the steady state and transient characteristics of a household heat pump/air conditioning system, and the results showed that on–off cycling of the system decreased the coefficients of performance (COPs) to values smaller than 75% of the steady state values. Kim and Bullard (2001) experimentally investigated the shutdown and startup performances of a residential split R410A heat pump, but only considered the warm startup processes. The cooling capacity, dehumidification capacity, power consumption, and cycle characteristics during the startup process were analyzed. Kapadia, Jain, and Agarwal (2009) presented the detailed dynamic characteristics of household air conditioning systems using R410A and R22 as refrigerants based on dynamic models. These dynamic characteristics included the pressures, temperatures, and mass flow rates. As substitutes for hydrochlorofluorocarbons (HCFCs), HFCs would aggravate the effects of global warming to some extent. For instance, although the ozone depletion potential (ODP) of the substitutes R32 and R410A is zero, they have high global warming potential (GWP) values. As a natural refrigerant, R290 could take the place of R22 because it is environmentally friendly and has excellent thermophysical properties. Moreover, the use of R290 would assist in decreasing the expenditure of electric energy by the heat pump system (Palm, 2008; Zhou & Zhang, 2010). A series of studies on R290 air conditioning and heat pump systems has been performed in recent years. Hrnjak and Hoehne (2004) experimentally investigated a low-charge propane refrigeration system using microchannel condensers, discussed the thermal model for this system, and made a comparison between the results of the experiments and the thermal and charge modeling. Cai et al. (2015) simulated and compared the leakage characteristics of a rotary compressor using R290 and compressors using other kinds of refrigerants. The results indicated that in order to obtain a relatively high efficiency, the radial clearance for an R290 rotary compressor should be smaller 2 Page 2 of 13
than those of R410A and R22 compressors under the same conditions. Wu et al. (2016) conducted a study on the cold startup performance of a heat pump system with an R290 rotary compressor under refrigerating conditions, and the results showed that a small amount of liquid was generating at the discharge port of the cylinder at the beginning of the startup process. Because the vapor-saturation curve of R290 is close to the isentropic curve on the P-h graph, with only a small angle between them, liquid is likely to be generated during the isentropic compression of R290, especially in a low-temperature environment. The transient characteristics of R290 heat pumps and rotary compressors using other kinds of refrigerants have been widely discussed in recent papers (Kapadia et al., 2009; Lect et al., 2014; Cai et al., 2015). However, few studies have been conducted on the startup behaviors of R290 rotary compressors under low environmental temperature conditions. The purpose of the experiment reported here was to investigate both the cold and warm startup characteristics of an R290 heat pump system with a rotary compressor under low-temperature heating conditions. The heat pump system was an air conditioning system operated in the reverse cycle. The dynamic temperatures and pressures in the R290 heat pump system as functions of time were measured under low ambient temperature heating conditions. Meanwhile, to determine the temperature changes in the compressor and identify the generation of liquid in the cylinder, the pressures in the cylinder and temperatures in the shell were measured. The results reported here will be helpful for the potential development of heat pump systems that use R290. 2. Experiment 2.1 Experimental setup The startup characteristics of an R290 rotary compressor and its corresponding heat pump system were examined in this study. These characteristics included the cold startup characteristics and warm startup characteristics under heating conditions with a low ambient temperature, where the outdoor and indoor dry-bulb temperatures were 0 °C and 15 °C, respectively. Cold startup means the compressor was kept in the outdoor environment long enough for its temperature to equalize with that of the environment before it was started. Warm startup means the compressor was shut down for a short time before being restarted. In this study, the startup characteristics were determined using an experimental apparatus similar to that described in a previous study (Wu et al., 2016). As seen in Fig.1, the experimental setup was composed of a residential split heat pump system and psychrometric rooms. The heat pump system consisted of an evaporator, a condenser, an electronic expansion valve, a rotary compressor, and some attachments like fans and an accumulator. The psychrometric rooms consisted of two independent fixed humidity and temperature chambers: an indoor space and outdoor space. To meet the exact humidity and temperature demands during the experiments, an air handling unit was installed in each chamber. Some air-measuring devices were installed in the 3 Page 3 of 13
indoor space to measure the pressure, temperature, and other variables to allow the experimental conditions to be adjusted. Four K-type sheathed thermocouples were installed in the shell of the compressor to monitor the temperature changes in various locations during startup. These temperatures included the temperature of the suction pipe between the accumulator and compressor, cylinder temperature, oil mixture temperature at the bottom of the oil sump, and motor temperature at the top of the stator. As shown in Fig.2, two miniature piezo-resistive dynamic pressure sensors (#1 and #2) were installed in the sub-bearing (positions #1 and #2) to measure the suction chamber and compression chamber pressures in the cylinder. The locations of positions #1 and #2 formed angles of 142° and 10° with the center line, respectively, as seen in Fig.2, where position #1 is close to the discharge port. Because our previously published research (Wu et al., 2016) contained a detailed description of the experimental apparatus, detailed descriptions of the compressor, heat exchangers, etc. will not be provided in this paper. The next section will describe the experimental procedure and conditions. 2.2 Experimental procedure and conditions For the cold startup characteristics investigation, before the heat pump was started, it was kept in the low-temperature outdoor space for 12 h until the component temperatures were the same as the environment temperature. For the warm startup characteristics investigation, the heat pump system was restarted 7 min after being shut down. The experiments were performed under a low ambient temperature heating capacity condition. The outdoor air dry-bulb temperature was 0 °C, and the indoor air dry-bulb temperature was 15 °C. The air volume flow rate for the outdoor unit was 655 m3·min-1, and the compressor operating frequency was 108 Hz. A conventional mineral oil (MO) was selected as the lubricant. The temperatures in the compressor shell were acquired every second. The temperatures and pressures in the heat pump system were measured every 5 s during the startup processes. Because of the rapid pressure change in the cylinder, the sampling rate of the data acquisition system was 30 kHz per channel, which meant 30,000 pressure values were collected per second. Before the startup of the system, all of the temperature and pressure transducers and data-collecting devices were working. 3. Results and discussion 3.1 Variation of pressures in system Fig.3 shows how the pressures (gauge pressures) changed over time during the cold and warm startups under a low ambient temperature heating condition. Before the cold startup, as shown in Fig.3(a), all four of the previously discussed pressures reached an equilibrium state (390 kPa). Before the warm startup, as shown in Fig.3(b), these four pressures, which were equal, gradually decreased over time to 410 kPa before the compressor was 4 Page 4 of 13
restarted. After the cold startup, as shown in Fig.3(a), it took 7 min and 11 min for the suction and discharge sides of the heat pump to reach equilibrium, respectively. The suction pressure reached a minimum of -48 kPa after 65 s, with a corresponding saturation temperature of -56.2 °C. Then, it slowly increased to the steady state value of 210 kPa in 420 s. The discharge pressure first showed a drastic increase, and then slowly increased to the steady state value of 1530 kPa. The inlet pressure of the expansion valve showed a variation trend similar to that of the discharge pressure. The only difference was that the equilibrium value was 60 kPa lower than that of the discharge pressure. The outlet pressure of the expansion valve had a variation trend similar to that of the suction pressure, except the equilibrium value was 110 kPa higher than that of the suction pressure. After the warm startup, as shown in Fig.3(b), the equilibrium states for the suction and discharge sides of the heat pump were reached in 5 min and 8 min, respectively. The suction pressure reached a minimum of 155 kPa after 130 s, with a corresponding saturation temperature of -18.8 °C. Then, it slowly increased to the steady state value of 210 kPa in 305 s. The discharge pressure sharply increased in the first 100 s, and then slowly increased to the equilibrium value of 1530 kPa. The inlet pressure of the expansion valve had a variation trend similar to that of the discharge pressure. The only difference was that the equilibrium value was 60 kPa lower than that of the discharge pressure. The outlet pressures of the expansion valve had a variation trend similar to that of the suction pressure, except the equilibrium value was 110 kPa higher than that of the suction pressure as a result of the flow resistance loss. Under a low ambient temperature heating condition, the pressure changes over time in the system during the cold startup were much different from those during the warm startup. During the cold startup, the minimum suction pressure (-48 kPa) was lower than the atmospheric pressure (100 kPa) because little refrigerant was contained in the suction pipe (Li et al., 2015), and a dramatic pressure drop occurred with the suction of the cylinder. Consequently, ambient air may have leaked into the heat pump system from the low-pressure side if the system was poorly sealed. In contrast, during the warm startup, the minimum suction pressure (155 kPa) was higher than the atmospheric pressure (100 kPa) because of the large quantity of refrigerant in the suction pipe (Li et al., 2015). Thus, no air could enter the heat pump system even if it was poorly sealed. Because of the low oil temperature, a large amount of refrigerant dissolved into the oil sump under the cold startup condition. Thus, only a small amount of refrigerant was contained on the suction side before the startup. When the compressor started, little refrigerant was drawn into the compressor, which resulted in a large pressure drop in the suction pipe. In contrast, under the warm startup condition, little refrigerant was dissolved in the oil sump because of the high oil temperature. Thus, there was more refrigerant on the low-pressure side. When the compressor started, a large quantity 5 Page 5 of 13
of refrigerant was drawn into the compressor, which resulted in a small pressure drop in the suction pipe. Under a low ambient temperature heating condition, the equilibrium time (the time required for the temperature or pressure to achieve stability and remain almost unchanged) for the high-pressure side was greater than that for the low-pressure side during both the cold and warm startup processes, as shown in Fig.3(a) and Fig.3(b). This was because it took longer to increase the internal pressure of the shell during the startup processes. The equilibrium times for the high-pressure side of the heat pump during both the cold and warm startup processes under the low ambient temperature heating condition were longer than those under the standard cooling condition, as shown in a previous study (Wu et al., 2016), because of the higher pressure increase needed on the high-pressure side under the low ambient temperature heating condition. The equilibrium times for the suction side (5 min) and discharge side (8 min) of the R290 heat pump during the warm startup in this study were much longer than those of the R410A heat pump system (low-pressure side: 2 min; high-pressure side: 3 min) (Kim & Bullard, 2001). This might have been because the pressure ratio (approximately 7.3) of the R290 heat pump system in this study was much larger than that of the R410A system (approximately 3) under the same operating conditions. As an alternative, it could have been because the R290 charge was much smaller than the R410A charge for the same cooling capacity as a result of the higher latent heat of vaporization for R290. Thus, less refrigerant was drawn into the cylinder. 3.2 Variation of temperatures in system Fig.4 shows how the temperatures changed over time in the R290 heat pump system during the cold and warm startups under the low ambient temperature heating condition. Before the cold startup, as shown in Fig.4(a), all four of the previously discussed temperatures had reached an equilibrium state value of 0.5 °C. Before the warm startup, as shown in Fig.4(b), the four previously discussed temperatures had not reached an equilibrium state. The discharge temperature decreased rapidly, whereas the other three temperatures decreased slowly over time. After the cold startup, as shown in Fig.4(a), the suction temperature reached a minimum of -23 °C after 80 s, which was higher than the saturation temperature (-56.2 °C) (see Fig.3(a)). Thus, no condensation would occur in the suction pipe. After that, the suction temperature gradually increased until reaching the equilibrium state value of -3 °C in 500 s. The discharge temperature first rapidly increased and then slowly increased to the equilibrium value of 61 °C after 900 s. The inlet temperature of the expansion valve had a variation trend similar to that of the discharge pressure within the first 250 s. Then, it slowly decreased to the equilibrium value of 24 °C in 350 s. The outlet temperature of the expansion valve was higher than the suction temperature within the first 500 s, while it was lower than the suction temperature after 500 s. 6 Page 6 of 13
After the warm startup, as shown in Fig.4(b), the suction pressure reached a minimum of -7 °C after 130 s, which was higher than the saturation temperature (-18.8 °C) (see Fig.3(b)). Thus, no condensation would occur in the suction pipe. Then, it slowly increased to the steady state value of -3 °C in 350 s. The discharge temperature increased rapidly at first then slowly increased to the equilibrium state value of 61 °C after 600 s. The inlet temperature of the expansion valve had a variation trend similar to that of the discharge pressure within the first 95 s. Then, it slowly decreased to the equilibrium state value of 24 °C in 290 s. The outlet temperature of the expansion valve was higher than the suction temperature within the first 280 s, while it was lower than the suction temperature after 280 s because of the flow resistance loss in the outdoor heat exchanger. Under the low ambient temperature heating condition, the temperature change in the system during the cold startup was quite different from that during the warm startup. The equilibrium times for the four measured temperatures during the cold startup were longer than those during the warm startup (see Fig.4) because it took longer to heat or cool the components of the heat pump system during the cold startup process. The minimum suction temperature (-23 °C) during the cold startup was lower than that of the warm startup (-7 °C) as a result of the larger pressure drop in the suction pipe during the cold startup. 3.3 Variation of temperature in compressor Fig.5 shows the temperatures change in the compressor during the cold and warm startup processes under the low ambient temperature heating condition. Before the cold startup, as shown in Fig.5(a), all four of the previously discussed temperatures (the temperatures of the suction pipe, cylinder, oil sump, and motor stator) reached an equilibrium state value of 0.5 °C. In contrast, these four temperatures slowly decreased over time before the warm startup, as shown in Fig.5(b). After the cold startup, as shown in Fig.5(a), the motor stator temperature immediately increased until reaching the equilibrium state. The oil temperature began to increase after 90 s during the cold startup. Meanwhile, the cylinder temperature first decreased slightly and then slowly increased to the steady state. One reason for this was the low-temperature refrigerant being drawn into the cylinder. Another reason was the fact that the high pressure had not been established in the shell, which resulted in a low discharge temperature. Thus, the cylinder temperature first decreased slightly. After the high pressure was established, which resulted in a high discharge temperature, the cylinder temperature gradually increased to the equilibrium state. Under the cooling condition, the cylinder temperature showed a sharp drop during the cold startup, as previously reported (Wu et al., 2016). The suction pipe temperature showed a rapid decrease to -26 °C in 68 s. Then, it transiently peaked at about -0.2 °C because a much larger quantity of refrigerant was drawn into the cylinder from the suction pipe. After this transient peak, the suction 7 Page 7 of 13
pipe temperature gradually decreased to the equilibrium state value of -7.5 °C. After the warm startup, as shown in Fig.5(b), all four temperatures in the compressor first decreased and then increased to the equilibrium state value, which was much different from the results for the cold startup under the low ambient temperature heating condition and the cold startup under the cooling condition, as shown in a previous report (Wu et al., 2016). At the initial stage of the warm startup, as shown in Fig.5(b), the suction pipe temperature sharply decreased with a decrease in the suction pressure, as shown in Fig.3(b). The refrigerant discharged from the muffler had a low temperature because of the low discharge pressures. Thus, the motor and cylinder were cooled by the discharged refrigerant in the shell and compressed refrigerant in the cylinder, respectively. The oil temperature continued to decrease at the same rate as before the startup and reached a minimum of 31 °C after 100 s. With the discharge temperature increase (see Fig.4(b)), the temperatures of the motor, cylinder, and oil sump increased. After reaching the equilibrium state, the four previously discussed temperatures for the cold startup were equal to that of the warm startup. The minimum suction pipe temperature (-13 °C) during the warm startup was higher than that during the cold startup (-26 °C). It is likely that there was much more refrigerant in the suction pipe and outdoor heat exchanger during the warm startup than during the cold startup (Li et al., 2015). 3.4 Pressures change at position #1 Fig.6 shows the pressure changes at position #1 during both the cold and warm startup processes under the low ambient temperature heating condition. The angle between position #1 and the center line was 142°. Thus, the pressure sensor at position #1 was used to measure the pressures in both the compression chamber and suction chamber. Before the cold startup, as shown in Fig.6(a), the pressures reached an equilibrium state (350 kPa). In contrast, before the warm startup, as shown in Fig.6(b), the pressures gradually decreased over time to 350 kPa before the compressor was restarted. As Fig.6(a) shows, the maximum pressure measured by sensor #1 gradually increased to 510 kPa in 10 s during the cold startup. In contrast, during the warm startup, the maximum pressure at position #1 rose dramatically to 600 kPa in 1 s. Then, the pressure in the compressor gradually increased, as shown in Fig.6(b). These results indicate that the time required for the pressure to increase at position #1 during the warm startup was shorter than that during the cold startup under the low ambient temperature heating condition. No liquid slugging occurred during the initial compression process for both the cold and warm startup processes. In addition, the pressure increase during the warm startup was quicker than that during the cold startup. 3.5 Pressures changes at position #2 8 Page 8 of 13
Fig.7 shows the pressures changes over time at position #2 during the cold and warm startup processes under the low ambient temperature heating condition. The angle between position #2 and the center line was 10°. Thus, the pressure sensor at position #2 was mainly used to measure the pressure just before discharge. Before the cold startup, as shown in Fig.7(a), the pressures reached an equilibrium state (350 kPa). In contrast, before the warm startup, as shown in Fig.7(b), the pressures gradually decreased over time to 350 kPa before the compressor was restarted. As Fig.7(a) shows, the discharge pressure at position #2 dramatically increased to 1070 kPa in 1 s, and then rapidly decreased to 550 kPa during the cold startup. Subsequently, the pressure difference between the suction chamber and compression chamber in the cylinder increased over time. In contrast, as shown in Fig.7(b), during the warm startup, the discharge pressure soared to 930 kPa in 1 s before rapidly decreasing to 600 kPa. Then, the pressure gradually increased again. As Fig.7 shows, the pressure increase in the cylinder during the warm startup was faster than that during the cold startup. This could have been because there was much more refrigerant in the suction pipe and outdoor heat exchanger during the warm startup than during the cold startup (Li et al., 2015). Thus, it took longer for the cold startup to transfer the refrigerant from the discharge side to the suction side. With less refrigerant on the low-pressure side, it would have taken more time for the pressure to increase. The maximum pressures at position #2 during the cold and warm startup processes (1070 kPa and 930 kPa, respectively) were both much higher than their corresponding discharge pressures, as shown in Fig.3, because of a slight liquid slugging. However, they were both lower than that during a cold startup process (2210 kPa) under a cooling condition, as previously reported (Wu et al., 2016). As previously mentioned, no condensation occurred in the suction pipe during the cold and warm startup processes. Moreover, many compression processes occur in the cylinder within 1 s with a compressor operating frequency of 108 Hz. Thus, any liquid that existed in the suction pipe before the startup process must have been discharged from the cylinder within 1 s. For these two reasons, the slight liquid slugging may have occurred because of R290 condensation in the cylinder.
4
Conclusions Experimental investigations of the cold and warm startup characteristics of an R290 rotary
compressor in a heat pump system were conducted under a low ambient temperature heating condition. The experiments were conducted in psychrometric rooms with fixed humidity and temperature values. We measured four different temperatures in the compressor shell and pressures in the cylinder during the cold and warm startup processes. The pressures and temperatures in the R290 heat pump system during the startup processes were also measured. The following conclusions may be drawn: 1) Under the low ambient temperature heating condition, the minimum suction pressure (-48 9 Page 9 of 13
kPa) during the cold startup was lower than that during the warm startup (155 kPa). Air may have been sucked into the system as a result of the sub-atmospheric pressure under the cold startup process, which could easily lead to security risks because of the flammability of R290. The equilibrium times for the suction side (5 min) and discharge side (8 min) of the R290 heat pump during the warm startup in this study were much longer than those of an R410A heat pump system (low-pressure side: 2 min; high-pressure side: 3 min)(Kim & Bullard, 2001). This resulted in a greater electric power consumption for the R290 heat pump to build up these pressures. 2) Under the low ambient temperature heating condition, the equilibrium time for the four discussed temperatures were longer than those during the warm startup. The minimum suction temperature (-23 °C) during the cold startup was lower than that during the warm startup (-7 °C). The very low suction temperature during the cold startup had adverse effects on the solubility of R290 in oil, and thus on the flow characteristic of the R290/oil mixture. 3) No liquid slugging occurred in the cylinder during the initial compression process for both startup processes, whereas it did occur in the late stage of the discharge process. The maximum forces in the cylinder during the cold and warm startup processes under the heating condition were lower than those under the cooling condition (Wu et al., 2016), which indicated that the heating condition is beneficial for improving the reliability and prolonging the life of an R290 compressor. Acknowledgement This work was supported by the HCFC phase out management plane in room heat pump sector technical assistance fund (China). References Cai, D., He, G., Yokoyama, T., Tian, Q., Yang, X., & Pan, J. (2015). Simulation and comparison of leakage characteristics of R290 in rolling piston type rotary compressor. International Journal of Refrigeration, 53, 42-54. doi:10.1016/j.ijrefrig.2015.02.001 Fukuda;, T., & Hayano;, M. (1996). HFC/POE Lubricity Evaluation on the Rotary Compressor in System Operation. International Compressor Engineering Conference at Purdue, 121-126. Haberschill, P., Gay, L., Aubouin, P., & Lallemand, M. (2003). Dynamic model of a vapor-compression refrigerating machine using R-407C. HVAC&R Research, 9(4), 451-466. Hrnjak, P., & Hoehne, M. (2004). Charge minimization in systems and components using hydrocarbons as a refrigerant. Urbana, 51, 61801. Judge, J., Hwang, Y., & Radermacher, R. (1996). A transient and steady state study of pure and mixed refrigerants in a residential heat pump (Doctoral dissertation, research directed by Dept. of Mechanical Engineering.University of Maryland at College Park). Kim, M.-H., & Bullard, C. W. (2001). Dynamic characteristics of a R-410A split air-conditioning system.
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Lect, K. A. A. J. A., & Al-Amir, Q. R. A. (2014). Performance evaluation of small scale air-conditioning system using R22 and alternative refrigerants. Journal of Engineering, 20(1). Murphy, W. E., & Goldschmidt, V. W. (1984). Transient response of air conditioners―A qualitative interpretation through a sample case. ASHRAE transactions, 90(1B), 997-1008. Murphy, W. E., & Goldschmidt, V. W. (1985). Cyclic characteristics of a typical residential air conditioner-modeling of start-up transients. ASHRAE transactions, 91(2A), 427-437. Murphy, W. E., & Goldschmidt, V. W. (1986). Cycling characteristics of a residential air conditioner-modeling of shutdown transients. ASHRAE Transactions, 92(1A), 186-202. Protocol, M. (1987). Montreal protocol on substances that deplete the ozone layer. Sept, 16(1987), 26. Palm, B. (2008). Hydrocarbons as refrigerants in small heat pump and refrigeration systems–a review. International Journal of Refrigeration, 31(4), 552-563. doi:10.1016/j.ijrefrig.2007.11.016 Sami, S. M., & Dahmani, A. (1996). Numerical prediction of dynamic performance of vapour-compression heat pump using new HFC alternatives to HCFC-22. Applied thermal engineering, 16(8), 691-705. Tanaka, N., Ikeuchi, M., & Yamanaka, G. (1982). Experimental study on the dynamic characteristics of a heat pump. ASHRAE transactions, 88, 323-331. Wu, J., Lin, J., Zhang, Z., Chen, Z., Xie, J., & Lu, J. (2016). Experimental investigation on cold startup characteristics of a rotary compressor in the R290 air-conditioning system under cooling condition. International Journal of Refrigeration, 65, 209-217. Zhou, G., & Zhang, Y. (2010). Performance of a split-type air conditioner matched with coiled adiabatic capillary tubes using HCFC22 and HC290. Applied Energy, 87(5), 1522-1528. doi:10.1016/j.apenergy.2009.10.005
Fig.1 Schematic of experimental apparatus
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Fig.2 Pressure sensors installed in sub-bearing
（a）cold startup （b）warm startup Fig.3 Variation of gauge pressures in heat pump system during startup
（b）warm startup （a）cold startup Fig.4 Variation of temperatures in heat pump system
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（a）cold startup （b）warm startup Fig.5 Temperatures change in compressor during startup
（a）cold startup （b）warm startup Fig.6 Variation of pressures at position #1
（b）warm startup （a）cold startup Fig.7 Variation of pressure at position #2 during startup
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S0140-7007(17)30092-0 http://dx.doi.org/doi: 10.1016/j.ijrefrig.2017.02.029 JIJR 3573
To appear in:
International Journal of Refrigeration
Received date: Revised date: Accepted date:
29-4-2016 21-2-2017 24-2-2017
Please cite this article as: Jie Lin, Jianhua Wu, Ze Zhang, Zhenhua Chen, Jing Xie, Jun Lu, Experimental investigation of startup characteristics of R290 rotary compressor under low ambient temperature heating condition, International Journal of Refrigeration (2017), http://dx.doi.org/doi: 10.1016/j.ijrefrig.2017.02.029. This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting proof before it is published in its final form. Please note that during the production process errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.
Experimental investigation of startup characteristics of R290 rotary compressor under low ambient temperature heating condition Jie Lin a ,Jianhua Wu a, *, Ze Zhang a, Zhenhua Chen a, b, Jing Xieb, Jun Lub a School of Energy and Power Engineering, Xi’an Jiaotong University, Xi’an 710049, PR China b Guangdong Meizhi Compressor Co. Ltd, Guangdong 528333, PR China *
Corresponding author. Tel.: +82 029 8266 3786, +82 13689206050. E-mail address: [email protected] (Jianhua Wu). Postal address: No. 28, Xian-ning West Rd., Xi’an City 710049, Shaanxi, PR China. Highlights
Startup characteristics of a rotary compressor in the R290 heat pump system were experimentally investigated Pressures and temperatures in the R290 heat pump system during cold and warm startup processes were measured. Temperatures in the rotary compressor shell and the pressures in cylinder during two startup processes were measured
Abstract: The cold and warm startup characteristics of an R290 heat pump system and its rotary compressor were experimentally investigated under a low ambient temperature heating condition. These startup characteristics included the temperatures in the compressor shell and system and the pressures in the cylinder and system. The results showed that the minimum suction gauge pressure (-48 kPa) during a cold startup was lower than that during a warm startup (155 kPa) under a low ambient temperature heating condition. In addition, the time required for the R290 heat pump system to reach a steady pressure was much longer than that of a system using R410A as the refrigerant. Compared with that of the cooling condition, smaller amounts of liquid were generated in the cylinder at the beginnings of both the cold and warm startup processes under the low ambient temperature heating condition. Keywords：R290; Experimental Investigation; Rotary Compressor; Heating Condition; Startup 1. Introduction Air conditioners, heat pumps, and other commercial refrigeration equipment are widely used in contemporary society. On–off cycling operation is a common way to control the capacity of indoor heating and cooling equipment. The startup of a heat pump is a complicated transient process, in which the refrigerant mass, heat load, torque, and rotational speed of the compressor vary. Thus, the startup process for a heat pump has an important effect on the reliability and efficiency of the system. Murphy and Goldschmidt (1984, 1985, and 1986) conducted several studies on the startup 1 Page 1 of 13
dynamics of a 3 ton split-type air conditioning system. A steady-state correlation was used for the compressor startup in the startup model. Tanaka et al. (1982) experimentally studied the dynamic characteristics of a heat pump during the startup and shutdown processes. The refrigerant mass distributions in various components, including the evaporator, condenser, and rotary compressor, were measured in their experiment. Based on their results, they made some useful technical suggestions to improve the startup performance. The Montreal Protocol was enacted in 1987 and included the decision to phase out R22 because it causes environmental problems (Protocol, 1987). This began a new period in transient characteristics research, and many alternatives to R22 have appeared, such as R134a, R32, and R290. Haberschill et al. (2003), Judge et al. (1996), and Sami and Dahmani (1996) presented a series of dynamic simulations of air conditioning or heat pump systems with different kinds of refrigerants. Fukuda and Hayano (1996) experimentally evaluated the lubricating characteristics of halogenated fluorocarbon/polyolester (HFC/POE) pairs consisting of R407C, R410A, and two kinds of POEs on the bearings of a rotary compressor in comparison with an R22 and mineral oil (MO) combination during the startup process. Judge, Hwang, and Radermacher (1996) experimentally and theoretically investigated the steady state and transient characteristics of a household heat pump/air conditioning system, and the results showed that on–off cycling of the system decreased the coefficients of performance (COPs) to values smaller than 75% of the steady state values. Kim and Bullard (2001) experimentally investigated the shutdown and startup performances of a residential split R410A heat pump, but only considered the warm startup processes. The cooling capacity, dehumidification capacity, power consumption, and cycle characteristics during the startup process were analyzed. Kapadia, Jain, and Agarwal (2009) presented the detailed dynamic characteristics of household air conditioning systems using R410A and R22 as refrigerants based on dynamic models. These dynamic characteristics included the pressures, temperatures, and mass flow rates. As substitutes for hydrochlorofluorocarbons (HCFCs), HFCs would aggravate the effects of global warming to some extent. For instance, although the ozone depletion potential (ODP) of the substitutes R32 and R410A is zero, they have high global warming potential (GWP) values. As a natural refrigerant, R290 could take the place of R22 because it is environmentally friendly and has excellent thermophysical properties. Moreover, the use of R290 would assist in decreasing the expenditure of electric energy by the heat pump system (Palm, 2008; Zhou & Zhang, 2010). A series of studies on R290 air conditioning and heat pump systems has been performed in recent years. Hrnjak and Hoehne (2004) experimentally investigated a low-charge propane refrigeration system using microchannel condensers, discussed the thermal model for this system, and made a comparison between the results of the experiments and the thermal and charge modeling. Cai et al. (2015) simulated and compared the leakage characteristics of a rotary compressor using R290 and compressors using other kinds of refrigerants. The results indicated that in order to obtain a relatively high efficiency, the radial clearance for an R290 rotary compressor should be smaller 2 Page 2 of 13
than those of R410A and R22 compressors under the same conditions. Wu et al. (2016) conducted a study on the cold startup performance of a heat pump system with an R290 rotary compressor under refrigerating conditions, and the results showed that a small amount of liquid was generating at the discharge port of the cylinder at the beginning of the startup process. Because the vapor-saturation curve of R290 is close to the isentropic curve on the P-h graph, with only a small angle between them, liquid is likely to be generated during the isentropic compression of R290, especially in a low-temperature environment. The transient characteristics of R290 heat pumps and rotary compressors using other kinds of refrigerants have been widely discussed in recent papers (Kapadia et al., 2009; Lect et al., 2014; Cai et al., 2015). However, few studies have been conducted on the startup behaviors of R290 rotary compressors under low environmental temperature conditions. The purpose of the experiment reported here was to investigate both the cold and warm startup characteristics of an R290 heat pump system with a rotary compressor under low-temperature heating conditions. The heat pump system was an air conditioning system operated in the reverse cycle. The dynamic temperatures and pressures in the R290 heat pump system as functions of time were measured under low ambient temperature heating conditions. Meanwhile, to determine the temperature changes in the compressor and identify the generation of liquid in the cylinder, the pressures in the cylinder and temperatures in the shell were measured. The results reported here will be helpful for the potential development of heat pump systems that use R290. 2. Experiment 2.1 Experimental setup The startup characteristics of an R290 rotary compressor and its corresponding heat pump system were examined in this study. These characteristics included the cold startup characteristics and warm startup characteristics under heating conditions with a low ambient temperature, where the outdoor and indoor dry-bulb temperatures were 0 °C and 15 °C, respectively. Cold startup means the compressor was kept in the outdoor environment long enough for its temperature to equalize with that of the environment before it was started. Warm startup means the compressor was shut down for a short time before being restarted. In this study, the startup characteristics were determined using an experimental apparatus similar to that described in a previous study (Wu et al., 2016). As seen in Fig.1, the experimental setup was composed of a residential split heat pump system and psychrometric rooms. The heat pump system consisted of an evaporator, a condenser, an electronic expansion valve, a rotary compressor, and some attachments like fans and an accumulator. The psychrometric rooms consisted of two independent fixed humidity and temperature chambers: an indoor space and outdoor space. To meet the exact humidity and temperature demands during the experiments, an air handling unit was installed in each chamber. Some air-measuring devices were installed in the 3 Page 3 of 13
indoor space to measure the pressure, temperature, and other variables to allow the experimental conditions to be adjusted. Four K-type sheathed thermocouples were installed in the shell of the compressor to monitor the temperature changes in various locations during startup. These temperatures included the temperature of the suction pipe between the accumulator and compressor, cylinder temperature, oil mixture temperature at the bottom of the oil sump, and motor temperature at the top of the stator. As shown in Fig.2, two miniature piezo-resistive dynamic pressure sensors (#1 and #2) were installed in the sub-bearing (positions #1 and #2) to measure the suction chamber and compression chamber pressures in the cylinder. The locations of positions #1 and #2 formed angles of 142° and 10° with the center line, respectively, as seen in Fig.2, where position #1 is close to the discharge port. Because our previously published research (Wu et al., 2016) contained a detailed description of the experimental apparatus, detailed descriptions of the compressor, heat exchangers, etc. will not be provided in this paper. The next section will describe the experimental procedure and conditions. 2.2 Experimental procedure and conditions For the cold startup characteristics investigation, before the heat pump was started, it was kept in the low-temperature outdoor space for 12 h until the component temperatures were the same as the environment temperature. For the warm startup characteristics investigation, the heat pump system was restarted 7 min after being shut down. The experiments were performed under a low ambient temperature heating capacity condition. The outdoor air dry-bulb temperature was 0 °C, and the indoor air dry-bulb temperature was 15 °C. The air volume flow rate for the outdoor unit was 655 m3·min-1, and the compressor operating frequency was 108 Hz. A conventional mineral oil (MO) was selected as the lubricant. The temperatures in the compressor shell were acquired every second. The temperatures and pressures in the heat pump system were measured every 5 s during the startup processes. Because of the rapid pressure change in the cylinder, the sampling rate of the data acquisition system was 30 kHz per channel, which meant 30,000 pressure values were collected per second. Before the startup of the system, all of the temperature and pressure transducers and data-collecting devices were working. 3. Results and discussion 3.1 Variation of pressures in system Fig.3 shows how the pressures (gauge pressures) changed over time during the cold and warm startups under a low ambient temperature heating condition. Before the cold startup, as shown in Fig.3(a), all four of the previously discussed pressures reached an equilibrium state (390 kPa). Before the warm startup, as shown in Fig.3(b), these four pressures, which were equal, gradually decreased over time to 410 kPa before the compressor was 4 Page 4 of 13
restarted. After the cold startup, as shown in Fig.3(a), it took 7 min and 11 min for the suction and discharge sides of the heat pump to reach equilibrium, respectively. The suction pressure reached a minimum of -48 kPa after 65 s, with a corresponding saturation temperature of -56.2 °C. Then, it slowly increased to the steady state value of 210 kPa in 420 s. The discharge pressure first showed a drastic increase, and then slowly increased to the steady state value of 1530 kPa. The inlet pressure of the expansion valve showed a variation trend similar to that of the discharge pressure. The only difference was that the equilibrium value was 60 kPa lower than that of the discharge pressure. The outlet pressure of the expansion valve had a variation trend similar to that of the suction pressure, except the equilibrium value was 110 kPa higher than that of the suction pressure. After the warm startup, as shown in Fig.3(b), the equilibrium states for the suction and discharge sides of the heat pump were reached in 5 min and 8 min, respectively. The suction pressure reached a minimum of 155 kPa after 130 s, with a corresponding saturation temperature of -18.8 °C. Then, it slowly increased to the steady state value of 210 kPa in 305 s. The discharge pressure sharply increased in the first 100 s, and then slowly increased to the equilibrium value of 1530 kPa. The inlet pressure of the expansion valve had a variation trend similar to that of the discharge pressure. The only difference was that the equilibrium value was 60 kPa lower than that of the discharge pressure. The outlet pressures of the expansion valve had a variation trend similar to that of the suction pressure, except the equilibrium value was 110 kPa higher than that of the suction pressure as a result of the flow resistance loss. Under a low ambient temperature heating condition, the pressure changes over time in the system during the cold startup were much different from those during the warm startup. During the cold startup, the minimum suction pressure (-48 kPa) was lower than the atmospheric pressure (100 kPa) because little refrigerant was contained in the suction pipe (Li et al., 2015), and a dramatic pressure drop occurred with the suction of the cylinder. Consequently, ambient air may have leaked into the heat pump system from the low-pressure side if the system was poorly sealed. In contrast, during the warm startup, the minimum suction pressure (155 kPa) was higher than the atmospheric pressure (100 kPa) because of the large quantity of refrigerant in the suction pipe (Li et al., 2015). Thus, no air could enter the heat pump system even if it was poorly sealed. Because of the low oil temperature, a large amount of refrigerant dissolved into the oil sump under the cold startup condition. Thus, only a small amount of refrigerant was contained on the suction side before the startup. When the compressor started, little refrigerant was drawn into the compressor, which resulted in a large pressure drop in the suction pipe. In contrast, under the warm startup condition, little refrigerant was dissolved in the oil sump because of the high oil temperature. Thus, there was more refrigerant on the low-pressure side. When the compressor started, a large quantity 5 Page 5 of 13
of refrigerant was drawn into the compressor, which resulted in a small pressure drop in the suction pipe. Under a low ambient temperature heating condition, the equilibrium time (the time required for the temperature or pressure to achieve stability and remain almost unchanged) for the high-pressure side was greater than that for the low-pressure side during both the cold and warm startup processes, as shown in Fig.3(a) and Fig.3(b). This was because it took longer to increase the internal pressure of the shell during the startup processes. The equilibrium times for the high-pressure side of the heat pump during both the cold and warm startup processes under the low ambient temperature heating condition were longer than those under the standard cooling condition, as shown in a previous study (Wu et al., 2016), because of the higher pressure increase needed on the high-pressure side under the low ambient temperature heating condition. The equilibrium times for the suction side (5 min) and discharge side (8 min) of the R290 heat pump during the warm startup in this study were much longer than those of the R410A heat pump system (low-pressure side: 2 min; high-pressure side: 3 min) (Kim & Bullard, 2001). This might have been because the pressure ratio (approximately 7.3) of the R290 heat pump system in this study was much larger than that of the R410A system (approximately 3) under the same operating conditions. As an alternative, it could have been because the R290 charge was much smaller than the R410A charge for the same cooling capacity as a result of the higher latent heat of vaporization for R290. Thus, less refrigerant was drawn into the cylinder. 3.2 Variation of temperatures in system Fig.4 shows how the temperatures changed over time in the R290 heat pump system during the cold and warm startups under the low ambient temperature heating condition. Before the cold startup, as shown in Fig.4(a), all four of the previously discussed temperatures had reached an equilibrium state value of 0.5 °C. Before the warm startup, as shown in Fig.4(b), the four previously discussed temperatures had not reached an equilibrium state. The discharge temperature decreased rapidly, whereas the other three temperatures decreased slowly over time. After the cold startup, as shown in Fig.4(a), the suction temperature reached a minimum of -23 °C after 80 s, which was higher than the saturation temperature (-56.2 °C) (see Fig.3(a)). Thus, no condensation would occur in the suction pipe. After that, the suction temperature gradually increased until reaching the equilibrium state value of -3 °C in 500 s. The discharge temperature first rapidly increased and then slowly increased to the equilibrium value of 61 °C after 900 s. The inlet temperature of the expansion valve had a variation trend similar to that of the discharge pressure within the first 250 s. Then, it slowly decreased to the equilibrium value of 24 °C in 350 s. The outlet temperature of the expansion valve was higher than the suction temperature within the first 500 s, while it was lower than the suction temperature after 500 s. 6 Page 6 of 13
After the warm startup, as shown in Fig.4(b), the suction pressure reached a minimum of -7 °C after 130 s, which was higher than the saturation temperature (-18.8 °C) (see Fig.3(b)). Thus, no condensation would occur in the suction pipe. Then, it slowly increased to the steady state value of -3 °C in 350 s. The discharge temperature increased rapidly at first then slowly increased to the equilibrium state value of 61 °C after 600 s. The inlet temperature of the expansion valve had a variation trend similar to that of the discharge pressure within the first 95 s. Then, it slowly decreased to the equilibrium state value of 24 °C in 290 s. The outlet temperature of the expansion valve was higher than the suction temperature within the first 280 s, while it was lower than the suction temperature after 280 s because of the flow resistance loss in the outdoor heat exchanger. Under the low ambient temperature heating condition, the temperature change in the system during the cold startup was quite different from that during the warm startup. The equilibrium times for the four measured temperatures during the cold startup were longer than those during the warm startup (see Fig.4) because it took longer to heat or cool the components of the heat pump system during the cold startup process. The minimum suction temperature (-23 °C) during the cold startup was lower than that of the warm startup (-7 °C) as a result of the larger pressure drop in the suction pipe during the cold startup. 3.3 Variation of temperature in compressor Fig.5 shows the temperatures change in the compressor during the cold and warm startup processes under the low ambient temperature heating condition. Before the cold startup, as shown in Fig.5(a), all four of the previously discussed temperatures (the temperatures of the suction pipe, cylinder, oil sump, and motor stator) reached an equilibrium state value of 0.5 °C. In contrast, these four temperatures slowly decreased over time before the warm startup, as shown in Fig.5(b). After the cold startup, as shown in Fig.5(a), the motor stator temperature immediately increased until reaching the equilibrium state. The oil temperature began to increase after 90 s during the cold startup. Meanwhile, the cylinder temperature first decreased slightly and then slowly increased to the steady state. One reason for this was the low-temperature refrigerant being drawn into the cylinder. Another reason was the fact that the high pressure had not been established in the shell, which resulted in a low discharge temperature. Thus, the cylinder temperature first decreased slightly. After the high pressure was established, which resulted in a high discharge temperature, the cylinder temperature gradually increased to the equilibrium state. Under the cooling condition, the cylinder temperature showed a sharp drop during the cold startup, as previously reported (Wu et al., 2016). The suction pipe temperature showed a rapid decrease to -26 °C in 68 s. Then, it transiently peaked at about -0.2 °C because a much larger quantity of refrigerant was drawn into the cylinder from the suction pipe. After this transient peak, the suction 7 Page 7 of 13
pipe temperature gradually decreased to the equilibrium state value of -7.5 °C. After the warm startup, as shown in Fig.5(b), all four temperatures in the compressor first decreased and then increased to the equilibrium state value, which was much different from the results for the cold startup under the low ambient temperature heating condition and the cold startup under the cooling condition, as shown in a previous report (Wu et al., 2016). At the initial stage of the warm startup, as shown in Fig.5(b), the suction pipe temperature sharply decreased with a decrease in the suction pressure, as shown in Fig.3(b). The refrigerant discharged from the muffler had a low temperature because of the low discharge pressures. Thus, the motor and cylinder were cooled by the discharged refrigerant in the shell and compressed refrigerant in the cylinder, respectively. The oil temperature continued to decrease at the same rate as before the startup and reached a minimum of 31 °C after 100 s. With the discharge temperature increase (see Fig.4(b)), the temperatures of the motor, cylinder, and oil sump increased. After reaching the equilibrium state, the four previously discussed temperatures for the cold startup were equal to that of the warm startup. The minimum suction pipe temperature (-13 °C) during the warm startup was higher than that during the cold startup (-26 °C). It is likely that there was much more refrigerant in the suction pipe and outdoor heat exchanger during the warm startup than during the cold startup (Li et al., 2015). 3.4 Pressures change at position #1 Fig.6 shows the pressure changes at position #1 during both the cold and warm startup processes under the low ambient temperature heating condition. The angle between position #1 and the center line was 142°. Thus, the pressure sensor at position #1 was used to measure the pressures in both the compression chamber and suction chamber. Before the cold startup, as shown in Fig.6(a), the pressures reached an equilibrium state (350 kPa). In contrast, before the warm startup, as shown in Fig.6(b), the pressures gradually decreased over time to 350 kPa before the compressor was restarted. As Fig.6(a) shows, the maximum pressure measured by sensor #1 gradually increased to 510 kPa in 10 s during the cold startup. In contrast, during the warm startup, the maximum pressure at position #1 rose dramatically to 600 kPa in 1 s. Then, the pressure in the compressor gradually increased, as shown in Fig.6(b). These results indicate that the time required for the pressure to increase at position #1 during the warm startup was shorter than that during the cold startup under the low ambient temperature heating condition. No liquid slugging occurred during the initial compression process for both the cold and warm startup processes. In addition, the pressure increase during the warm startup was quicker than that during the cold startup. 3.5 Pressures changes at position #2 8 Page 8 of 13
Fig.7 shows the pressures changes over time at position #2 during the cold and warm startup processes under the low ambient temperature heating condition. The angle between position #2 and the center line was 10°. Thus, the pressure sensor at position #2 was mainly used to measure the pressure just before discharge. Before the cold startup, as shown in Fig.7(a), the pressures reached an equilibrium state (350 kPa). In contrast, before the warm startup, as shown in Fig.7(b), the pressures gradually decreased over time to 350 kPa before the compressor was restarted. As Fig.7(a) shows, the discharge pressure at position #2 dramatically increased to 1070 kPa in 1 s, and then rapidly decreased to 550 kPa during the cold startup. Subsequently, the pressure difference between the suction chamber and compression chamber in the cylinder increased over time. In contrast, as shown in Fig.7(b), during the warm startup, the discharge pressure soared to 930 kPa in 1 s before rapidly decreasing to 600 kPa. Then, the pressure gradually increased again. As Fig.7 shows, the pressure increase in the cylinder during the warm startup was faster than that during the cold startup. This could have been because there was much more refrigerant in the suction pipe and outdoor heat exchanger during the warm startup than during the cold startup (Li et al., 2015). Thus, it took longer for the cold startup to transfer the refrigerant from the discharge side to the suction side. With less refrigerant on the low-pressure side, it would have taken more time for the pressure to increase. The maximum pressures at position #2 during the cold and warm startup processes (1070 kPa and 930 kPa, respectively) were both much higher than their corresponding discharge pressures, as shown in Fig.3, because of a slight liquid slugging. However, they were both lower than that during a cold startup process (2210 kPa) under a cooling condition, as previously reported (Wu et al., 2016). As previously mentioned, no condensation occurred in the suction pipe during the cold and warm startup processes. Moreover, many compression processes occur in the cylinder within 1 s with a compressor operating frequency of 108 Hz. Thus, any liquid that existed in the suction pipe before the startup process must have been discharged from the cylinder within 1 s. For these two reasons, the slight liquid slugging may have occurred because of R290 condensation in the cylinder.
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Conclusions Experimental investigations of the cold and warm startup characteristics of an R290 rotary
compressor in a heat pump system were conducted under a low ambient temperature heating condition. The experiments were conducted in psychrometric rooms with fixed humidity and temperature values. We measured four different temperatures in the compressor shell and pressures in the cylinder during the cold and warm startup processes. The pressures and temperatures in the R290 heat pump system during the startup processes were also measured. The following conclusions may be drawn: 1) Under the low ambient temperature heating condition, the minimum suction pressure (-48 9 Page 9 of 13
kPa) during the cold startup was lower than that during the warm startup (155 kPa). Air may have been sucked into the system as a result of the sub-atmospheric pressure under the cold startup process, which could easily lead to security risks because of the flammability of R290. The equilibrium times for the suction side (5 min) and discharge side (8 min) of the R290 heat pump during the warm startup in this study were much longer than those of an R410A heat pump system (low-pressure side: 2 min; high-pressure side: 3 min)(Kim & Bullard, 2001). This resulted in a greater electric power consumption for the R290 heat pump to build up these pressures. 2) Under the low ambient temperature heating condition, the equilibrium time for the four discussed temperatures were longer than those during the warm startup. The minimum suction temperature (-23 °C) during the cold startup was lower than that during the warm startup (-7 °C). The very low suction temperature during the cold startup had adverse effects on the solubility of R290 in oil, and thus on the flow characteristic of the R290/oil mixture. 3) No liquid slugging occurred in the cylinder during the initial compression process for both startup processes, whereas it did occur in the late stage of the discharge process. The maximum forces in the cylinder during the cold and warm startup processes under the heating condition were lower than those under the cooling condition (Wu et al., 2016), which indicated that the heating condition is beneficial for improving the reliability and prolonging the life of an R290 compressor. Acknowledgement This work was supported by the HCFC phase out management plane in room heat pump sector technical assistance fund (China). References Cai, D., He, G., Yokoyama, T., Tian, Q., Yang, X., & Pan, J. (2015). Simulation and comparison of leakage characteristics of R290 in rolling piston type rotary compressor. International Journal of Refrigeration, 53, 42-54. doi:10.1016/j.ijrefrig.2015.02.001 Fukuda;, T., & Hayano;, M. (1996). HFC/POE Lubricity Evaluation on the Rotary Compressor in System Operation. International Compressor Engineering Conference at Purdue, 121-126. Haberschill, P., Gay, L., Aubouin, P., & Lallemand, M. (2003). Dynamic model of a vapor-compression refrigerating machine using R-407C. HVAC&R Research, 9(4), 451-466. Hrnjak, P., & Hoehne, M. (2004). Charge minimization in systems and components using hydrocarbons as a refrigerant. Urbana, 51, 61801. Judge, J., Hwang, Y., & Radermacher, R. (1996). A transient and steady state study of pure and mixed refrigerants in a residential heat pump (Doctoral dissertation, research directed by Dept. of Mechanical Engineering.University of Maryland at College Park). Kim, M.-H., & Bullard, C. W. (2001). Dynamic characteristics of a R-410A split air-conditioning system.
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Lect, K. A. A. J. A., & Al-Amir, Q. R. A. (2014). Performance evaluation of small scale air-conditioning system using R22 and alternative refrigerants. Journal of Engineering, 20(1). Murphy, W. E., & Goldschmidt, V. W. (1984). Transient response of air conditioners―A qualitative interpretation through a sample case. ASHRAE transactions, 90(1B), 997-1008. Murphy, W. E., & Goldschmidt, V. W. (1985). Cyclic characteristics of a typical residential air conditioner-modeling of start-up transients. ASHRAE transactions, 91(2A), 427-437. Murphy, W. E., & Goldschmidt, V. W. (1986). Cycling characteristics of a residential air conditioner-modeling of shutdown transients. ASHRAE Transactions, 92(1A), 186-202. Protocol, M. (1987). Montreal protocol on substances that deplete the ozone layer. Sept, 16(1987), 26. Palm, B. (2008). Hydrocarbons as refrigerants in small heat pump and refrigeration systems–a review. International Journal of Refrigeration, 31(4), 552-563. doi:10.1016/j.ijrefrig.2007.11.016 Sami, S. M., & Dahmani, A. (1996). Numerical prediction of dynamic performance of vapour-compression heat pump using new HFC alternatives to HCFC-22. Applied thermal engineering, 16(8), 691-705. Tanaka, N., Ikeuchi, M., & Yamanaka, G. (1982). Experimental study on the dynamic characteristics of a heat pump. ASHRAE transactions, 88, 323-331. Wu, J., Lin, J., Zhang, Z., Chen, Z., Xie, J., & Lu, J. (2016). Experimental investigation on cold startup characteristics of a rotary compressor in the R290 air-conditioning system under cooling condition. International Journal of Refrigeration, 65, 209-217. Zhou, G., & Zhang, Y. (2010). Performance of a split-type air conditioner matched with coiled adiabatic capillary tubes using HCFC22 and HC290. Applied Energy, 87(5), 1522-1528. doi:10.1016/j.apenergy.2009.10.005
Fig.1 Schematic of experimental apparatus
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Fig.2 Pressure sensors installed in sub-bearing
（a）cold startup （b）warm startup Fig.3 Variation of gauge pressures in heat pump system during startup
（b）warm startup （a）cold startup Fig.4 Variation of temperatures in heat pump system
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（a）cold startup （b）warm startup Fig.5 Temperatures change in compressor during startup
（a）cold startup （b）warm startup Fig.6 Variation of pressures at position #1
（b）warm startup （a）cold startup Fig.7 Variation of pressure at position #2 during startup
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