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Experimental investigation of an ejector-enhanced auto-cascade refrigeration system
Accepted Manuscript Experimental investigation of an ejector-enhanced auto-cascade refrigeration system Tao Bai, Gang Yan, Jianlin Yu PII: DOI: Reference:
S1359-4311(17)34695-1 https://doi.org/10.1016/j.applthermaleng.2017.10.053 ATE 11251
To appear in:
Applied Thermal Engineering
Received Date: Revised Date: Accepted Date:
17 July 2017 17 September 2017 8 October 2017
Please cite this article as: T. Bai, G. Yan, J. Yu, Experimental investigation of an ejector-enhanced auto-cascade refrigeration system, Applied Thermal Engineering (2017), doi: https://doi.org/10.1016/j.applthermaleng. 2017.10.053
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Experimental investigation of an ejector-enhanced auto-cascade refrigeration system Tao Bai, Gang Yan , Jianlin Yu Department of Refrigeration and Cryogenic Engineering，School of Energy and Power Engineering, Xi'an Jiaotong University, Xi'an 710049, China Abstract: This study presents the experimental investigation of an ejector-enhanced auto-cascade refrigeration cycle (EARC) with zeotropic refrigerant R134a/R23. Performance comparisons among the EARC and two conventional cycles were conducted at selected operating conditions. The effects of ambient temperature, charged mass fraction ratio of the mixture, throttle valve opening, and heat load on the performance characteristics of the EARC were investigated. The results indicated that the EARC had more advantages in terms of lower refrigeration temperature and higher energy utilization efficiency over the conventional cycles, and the coefficient of performance (COP) and exergy efficiency improvements of the EARC reached up to 9.6% and 25.1%, respectively. The throttle valve opening was optimal with respect to the maximum system exergy efficiency determination. The refrigerant R134a/R23 with the optimal mass fraction ratio of 0.70/0.30 was proposed. Keywords: experimental investigation; ejector; zeotropic refrigerant; auto-cascade
Corresponding author. Tel:+86-29-82668738. Fax:+86-29-82668725. Email: [email protected]
refrigeration; exergy efficiency Nomenclature COP
coefficient of performance
BV
ball valve
D
diameter (mm)
E
exergy flow rate (W)
EV
expansion valve
MF
mass fraction ratio
m
mass flow rate (kg h-1)
NXP
primary nozzle exit position (mm)
TOP
throttle valve opening (%)
P
pressure (MPa)
Qe
refrigeration capacity (W)
Qheater
heat load of electric heater
QLoss
heat loss
rP
pressure lift ratio of ejector
t
temperature (℃)
T
absolute temperature (K)
Te.av
average thermodynamic temperature (K)
UA
overall heat transfer coefficient
W
power (W)
xi
primary measured variables
Y
calculated parameter
Greek symbols
y
absolute uncertainty
Ex
system exergy efficiency
entrainment ratio
expansion ratio
Subscripts amb
ambient
com
compressor
d
diffuser
f
fan
eje
ejector
e
evaporator
i
inlet
o
outlet
m
mixing chamber
s
secoondary fluid
tot
total
1–14
state point
0
reference condition
1. Introduction At present, low temperature and small-size refrigeration systems were drawing
significant attention for medical storage, infrared sensor cooling, and frozen storage. The conventional cascade cycle and the auto-cascade refrigeration cycle (ARC) are the two common methods of achieving low temperature freezing in vapor compression refrigerators. To produce the cryogenic and ultra low cooling temperature over a range of -180 ℃ to -40 ℃ , the conventional cascade refrigeration system needs multiple compressors, heat exchangers and refrigerants [1]. The conventional cascade refrigeration system generally has a more complex system configuration and higher equipment cost in comparison with the auto-cascade refrigerator [2]. ARC systems have certain low-temperature refrigeration advantages, especially for the small-size low-temperature freezers. The ARC technique has been developed over a long period. Podbielniak first patented the design of an ARC system [3] in 1936. The first application of the ARC system in low-temperature refrigeration was by Kleemenko [4] for the liquefaction of natural gas in 1959. In past decades, energy saving and environmental protection have been the main topics of development in the refrigeration industry. There have been significant research efforts dedicated to performance improvement in AR systems, and many advanced refrigeration technologies have been developed, including system configuration optimization, environment friendly refrigerant replacement, and high-efficiency cycle modification. Bichev and Glukov [5] performed optimization research on a nitrogen–hydrocarbon mixture based ARC system with a thermodynamic method, and the specific refrigeration capacity and optimal discharge pressure were obtained by the theoretical calculation. Missimer [2]
compared the advantages and disadvantages of different refrigerant mixtures in ARC systems, and presented the corresponding replacement of CFC refrigerants mixtures by HFC mixtures. Gong et al [6] conducted performance comparisons among three mixed-refrigerant refrigeration cycles with thermodynamic calculations. The results indicated that both cycles (separation and non-separation) provided an approximately equal performance with an appropriate selection of mixture and operating pressures. Nayak and Venkatarathnam [7] experimentally investigated the phase separator dry-out phenomenon of an ARC refrigerator at different mixtures and operating temperatures and found that the cold-end temperature difference of the first heat exchanger can be used to predict the dry-out of the phase separator. Du et al [8] carried out experimental work on cycle characteristics of an ARC system with a R134a/R23 mixture and discussed the effects of the charged concentration and cooling water temperature on system operation. Wang et al [9] investigated the effect of the mixing position of the low-pressure refrigerant on the performance of an ARC system with a rectifying column. Sivakumar and Somasundaram [10] performed an exergetic analysis on a three-stage auto-cascade refrigeration system, and the environmentally friendly refrigerant zeotropic refrigerant R290/R23/R14 with a mass fraction of 0.218:0.346:0.436 was suggested as an alternative refrigerant for a three-stage ARC. Kim and Kim [11] experimentally investigated the operating characteristics of an ARC system with the zeotropic refrigerant coupling of R134a/CO2 and R290/CO2. Zhang et al [12] utilized a fractionation heat exchanger in an ARC system to improve the separation efficiency of R290 and CO 2; the
experimental results showed that the system a fractionation heat exchanger had a lower evaporation temperature. Wang et al [13] added two vapor-liquid separators in an ARC system to improve the separation efficiency of the zeotropic refrigerants and got good oil management; simulation results showed that the proposed cycle configuration was promising for ARC systems of -60 ℃ using binary refrigerants. Xu et al [14] simulated the effects of the expansion valve openings on the variation of the refrigerant composition and set up an experimental system to verify the variation. In addition, Chen et al [15] proposed a novel ARC with an additional heat exchanger to reduce the temperature before the throttle valve and obtained improvements of 37.5% in system coefficient of performance (COP) and 34.5% in exergy efficiency. Sreenivas et al[16] experimentally studied the relationship between the charged and circulation composition in a J-T refrigerator and an auto-cascade refrigerator with a single separator. From the open literature on ARC techniques, it can be concluded that proper refrigerant selection and cycle modification are both efficient methods of improving system performances. However, conventional ARC systems continue to operate with low system performance because of large irreversibility losses in the expansion process. It is a known fact that the ejector is a promising expansion device to recover the expansion-work and improve system performance. The ejector has a simple structure, low manufacturing cost, good expansion work recovery efficiency and pre-compression effect [17, 18]. In the past decades, considerable research has been devoted to the application of the ejector technique in various refrigeration and heat
pump systems [17]. Boccardi et al [19] carried out an experiment on a multi-ejector air to water heat pump, and tested the system and component performance at full and partial operational loads. Wang et al [20] investigated the performance of a novel ejector-enhanced refrigerator-freezer system two-phase-driven ejector. Liu et al [21] investigated three ejector enhanced household refrigerators and reduced the energy consumption by 7.75%. Yan et al [22] proposed an ejector-enhanced ARC system, and the theoretical simulation results illustrated that the system COP could be improved by 8.42–18.02%. Yu et al [23] designed an efficient ARC system in which the ejector utilized the pressure energy of the liquid from the separator to entrain the vapor from the evaporator, and the theoretical investigation showed that the system COP could be improved by 19.1%. Bai et al [24] further modified the ejector-enhanced ARC system with an additional bypass tube and carried out an advanced exergy analysis on the cycle performance. The results indicated that the compressor had the highest improvement priority. Tan et al [22, 23, 25] combined ejector refrigeration and ARC principle to achieve a temperature of -30 ℃ using zeotropic mixture R32/R236fa. Based on the open literature, it can be found that proper cycle modification with an ejector obviously improves the performance of ARC systems. However, the ejector-enhanced auto-cascade refrigeration cycle (EARC) has mainly focused on theoretical analysis. Few experimental studies have been published on the energy-saving effect of the ejector in the ARC system. This study presents an experimental investigation on an ejector enhanced auto-cascade refrigeration cycle
(EARC). The objectives of this study are to present the operating characteristics of the EARC system and evaluate the energy-saving advantages in comparison with the conventional ARC systems. Therefore, the ARC systems equipped with and without ejector were all established in the experimental setup. The influences of the key operational parameters, including the refrigerant charged concentration, ambient temperature, expansion valve opening, and heat load on the system operating characteristics are all discussed. 2. System description 2.1 System description and experimental setup Fig. 1 displays the schematic diagrams of the experimental setups for the EARC system and conventional ARC systems. The experimental setup consisted of a compressor,
a
condenser
with
an
adjustable
fan,
an
evaporator,
an
evaporator/condenser, an ejector, a separator, two metering valves used as throttling devices and two mass flow meters. The common zeotropic refrigerant R134a/R23 was used as the working fluid. The working principle of the EARC system is as follows: the mixture attains a high temperature and pressure vapour after the compressor and then it becomes partially condensed in the condenser. The mixture exits the condenser as a two-phase fluid and enters the separator. Owing to the partial fractionational characteristics of the zeotropic mixture as a two-phase state, the saturated vapor and liquid leaving the separator has different refrigerant compositions. The saturated vapor is enriched with R23 having a low boiling temperature, and the liquid contains more R134a having a high boiling
temperature. The saturated liquid as the primary fluid enters the ejector nozzle to entrain the vapor or two-phase fluid from the evaporator. Then, the two streams mix in the mixing chamber and a pressure rises in the diffuser. The fluid leaves the ejector and enters the evaporation condenser to liquefy the vapor form the separator. Then the liquid coming from the evaporation condenser enters the evaporator after a pressure drop in expansion valve 1 (EV1). The two-phase fluid leaving the ejector becomes superheat gas after the heat exchange with the R23 enriched fluid from the separator and then re-enters the compressor to complete the cycle. The corresponding schematic and P-h diagrams of the EARC system are shown in Fig. 2. To assess the energy-saving effect of the EARC system, two experimental conventional ARC systems (CARC-a and CARC-b) were built as shown in Fig. 1. The paths of the three cycles are depicted with different colors and line types. 4-9 -10-1 , and the CARC-a The EARC system was achieved with path 1-2-3-5-6-7-8-8' 4-9' -10-1 and and CARC-b systems were realized with the paths 1-2-3-5-6-7-8-8'' 4-9' 1-2-3-5-6-7-8-8''' -10-1 , respectively. Five ball valves (BV1–5) were installed in the
system to control the fluid path, and the two CARC systems could be switched according to the test conditions. Detailed switch operation methods are described in the table embedded in Fig. 1. In comparison with the CARC systems, the main cycle feature of the EARC system is that an ejector is used to replace the conventional throttle valve under the separator, and the expansion work can be partly recovered to gain a pre-compression effect. Thus, the suction pressure is increased and the compression ratio can be
lowered, which is beneficial for compressor power reduction and system COP enhancement. Therefore, the EARC system provides higher performance than the CARC systems. 2.2 Experimental instrumentation and data reduction The specifications of the key components including the compressor, condenser, evaporator, separator and ejector are listed in Table 1. A constant speed hermetic reciprocating compressor with a displacement of 26.3 cm3 was used. A spiral-flow type separator was employed at the outlet of the condenser. A high efficiency helical oil separator was installed between the condenser and the compressor. The electric hotline was twined on the surface of the coil tube of the evaporator to measure the refrigeration capacity. To minimize the system heat losses at low refrigeration temperatures in the operating conditions, the thermal insulation of the evaporator and connecting tubes were well designed. For the evaporator, a plastic rubber foam insulation tube of 30-mm thickness was wrapped around the evaporator tube, and then the evaporator tube was embedded in a box of pearlite, which is a good insulation material for cryogenic applications. In this condition, the total thickness of the two heat insulating materials was 100 mm. Other components and connecting tubes were wrapped by plastic rubber foam insulation material with 60-mm thickness. A brazed plate heat exchanger was used as the evaporation condenser. Two sight glasses were set up to monitor the fluid state at the compressor inlet and the evaporator outlet. The refrigerant temperatures along the cycle loop were measured by T-type thermocouples with an uncertainty of ±0.5 °C. The refrigerant pressures were
measured by the pressure transducers with an uncertainty of ±0.1% at full scales of 0.5 and 5.0 MPa. The mass flow rates of the vapor coming from the compressor and separator were measured by two Coriolis flow meters with a reading uncertainty of ±0.5%. The process was completed and the data were logged based on the parameters of the Agilent 34970A data acquisition instrument. The system was tested in a refrigerator test chamber built according to the technical standards of EN ISO 15502: 2005 and GB/T 8059.1~4-1995. Air with a specified temperature ranging in 10–50 ℃ and a relative humidity of 45–75% was provided. The detailed specifications of the system components and measuring instruments were listed in Tables 1 and 2. The schematic and dimension diagrams of the ejector with a fixed geometry nozzle are shown in the Fig. 3. The ejector consisted of a nozzle, suction chamber, mixing chamber, and diffuser. A convergent-divergent nozzle was designed based on the simulation results. The polytetrafluoroethylene gaskets were adopted to prevent refrigerant leakage and to adjust the nozzle exit position (NXP). The ejector performances were evaluated by pressure lift ratio rP and the entrainment ratio . The pressure lift ratio rP is a typical parameter for representing the pressure recovery effect, and the entrainment ratio indicates the mass lifting capability. It is noted that the heat losses were taken into account and calculated with equation (5). The overall heat transfer coefficient UA of 0.417 was measured with the energy balances method. The energy utilization performance of the system was evaluated by the system COP and exergy
efficiency Ex . The corresponding equations of these parameters are given in Table 3. 2.3. Uncertainty analysis To ensure the measurement accuracy of the experimental results, an uncertainty analysis was carried out for the considered parameters. With the accuracy of different sensors listed in Table 2, the relative uncertainties of the calculated parameters including the pressure lift ratio rP , entrainment ratio , exergy production rate Ee , COP and exergy efficiency Ex were obtained by equation (8) as shown in Table 3. The calculation shows that the maximum relative uncertainties of the pressure lift ratio rP , entrainment ratio , COP and exergy efficiency Ex were estimated to be 0.3%,0.5%, 1.4%, and 2.2%, respectively. 3. Results and discussion 3.1 Effect of ambient temperature The room temperature directly determines the separator inlet temperature and is an important parameter that influences the composition separation behaviors and the system operation performance. The effects of ambient temperature on the performances of different systems were studied. In addition, to compare the performances of the CARC-a, CARC-b, and EARC at different ambient temperatures, critical parameters such as the throttle valve opening (TOP) and the charged concentration of the mixture were optimized to obtain the maximum system exergy efficiency. The mass fraction of refrigerant couple R134a/R23 was selected as 0.7/0.3 at a total charge of 1000 g, and the TOP of the EARC system
was fixed at 15%. The openings of EV1 and EV2 of CARC systems were fixed at 15% and 70%, respectively, to obtain the maximum system exergy efficiency. The heat load on the evaporator Qheater was maintained at 100 W because the system was designed for low-temperature freezer applications with the same level of refrigeration capacity. The air temperature in the test room was varied in the range of 15–27 ℃. The condenser fan operated with a fixed power of 50 W to produce a fixed speed. Fig. 4 shows the effect of ambient temperature on the system COP and exergy efficiency. When tamb increased from 15 to 27 ℃，the COP of the EARC and the CARC-b decreased from 0.149 to 0.125 and from 0.139 to 0.119, respectively, and the COP of the CARC-a reached the peak value of 0.1412 at a tamb of 18 ℃. The largest system COP improvements of EARC over CARC-a and CARC-b reached up to 6.0% and 7.9%, respectively. The mechanism of the peak value of the CARC-a at a tamb of 18 ℃is as follows. The circulation concentration of R23 reduced as the increasing ambient temperature. In addition, the suction specific volume of the CARC-a increased with the ambient temperature because of the increase in the suction temperature. This resulted in a reduction in the mass flow rate of the compressor; i.e., mtot decreased from 8.176 kg.h-1 to 7.837 kg.h-1. Consequently, the compressor power of the CARC-a system reduced and the COP increased. When the tamb increased further, the increased suction pressure played the dominate role in the reduction of the specific volume at the compressor inlet. The total mass flow rate increased, the compressor power
increased, and the system COP decreased. Therefore, the COP of the CARC-a increased first and then decreased as the ambient temperature increased. When the ambient temperature increased from 15 to 18 ℃, the total mass flow rate reductions of 0.20 kg.h-1 and 0.31 kg.h-1 occurred in the EARC and the CARC-b, respectively, which were smaller than that of the CARC-a. Thus, the compressor power of the EARC and the CARC-b both increased as the ambient temperature increased. In this condition, only the CARC-a presented the maximum value at an ambient temperature of 18 ℃, and the EARC and the CARC-b showed a linear reduction in COP as the increasing ambient temperature. In addition, the system exergy production rate of the three cycles decreased as ambient temperature increased, and this eventually led to a decreased system exergy efficiency Ex . At the selected operating conditions, the exergy efficiency Ex of the EARC, CARC-a, and CARC-b varied in the ranges of 3.02–4.24%， 3.00–3.87% and 2.60–3.59%, respectively. The system maximum improvements of the exergy efficiency for the EARC over both the CARC-a and CARC-b were 9.8% and 19.3%, respectively. This means that the ejector efficiently utilized part of the expansion work and achieved a relatively high improvement in system performance. However, the energy-saving advantage of the EARC over the CARC-a gradually reduced as the ambient temperature increased. This can be explained as follows. The suction pressure of the EARC was larger at higher ambient temperatures because of the stronger pressure lift effect in this condition.
Thus, the total mass flow rate increased rapidly and led to an increase in the compressor power and a reduction in system COP. Furthermore, the CARC-a outperformed the CARC-b in system COP and exergy efficiency at the given operating conditions. Compared with CARC-a, CARC-b operated with a higher compression ratio. This resulted in the higher compressor power and lower system COP in CARC-b. The ejector performance was closely related to system performance in the EARC. Fig. 5 shows the effect of the ambient temperature tamb on the pressure lift ratio and the entrainment ratio of the ejector as well as the refrigerant mass flow rates. When the ambient temperature increased, the pressure lift ratio rP increased and the entrainment ratio first increased and then decreased. When the ambient temperature increased from 15 ℃ to 27 ℃, the pressure lift ratio rP and the entrainment ratio varied in the ranges of 1.198–1.227 and 0.574–0.624, respectively. These findings are understandable because the discharge pressure increased with an increase in the ambient temperature. This finally resulted in an increase in the pressure drop in the ejector nozzle, and more expansion work could be recovered and a larger pressure lifting effect could be offered by the ejector. On the other hand, the entrainment ratio first increased and then decreased as the tamb increased, and the maximum of 0.624 was obtained at a tamb of 18 ℃. The main reason for this was that the mass flow rate of the secondary fluid increased as the tamb , and the mass flow rate of the compressor and the ejector nozzle first decreased and then increased linearly as shown in this figure, eventually
contributing the peak value of the entrainment ratio at an ambient temperature of 18 ℃ eventually. Fig. 6 shows the effect of ambient temperature on the evaporation temperature of different cycles. The evaporation temperature te , i.e., evaporator inlet temperature, increased as the ambient temperature tamb . The mechanism for this can be explained as follows. When the ambient temperature increases, less high-boiling-temperature refrigerant, i.e., R134a, is cooled as liquid, and the concentration of R23 in the fluid entering the evaporator decreases. In this condition, the evaporation temperature increased. In addition, the EARC exhibited a lower evaporation temperature than that of the CARC-a and CARC-b. When the ambient temperature was increased from 15 ℃ to 27 ℃, the evaporation temperature of the EARC varied from -50.2 ℃ to -40.4 ℃. Compared with the CARC-a and CARC-b, the EARC operated with 2.7 ℃ and 4.1 ℃ lower evaporation temperatures, respectively. This indicates that the use of an ejector in ARC provides a lower evaporation temperature, which is favored for low temperature freezer applications. In addition, the evaporation temperature in the CARC-b was higher than that of the CARC-a. This is explained by the fact that the cooling effect of the fluid leaving the evaporator was not utilized in the evaporation condenser for CARC-b, which resulted in increased liquid temperature before EV1 and higher evaporation temperature. Therefore, it can be concluded that the EARC has more advantages over the CARC-a and CARC-b in terms of lower temperature and higher energy efficiency. 3.2 Effect of charged mass fraction ratio
The component composition of the zeotropic mixture is an important parameter in ARC systems. Circulation concentration is closely related to the charged concentration [16]. To investigate the effect of the charged concentration MFR23 on the performance of EARC, experiments research were carried out at a total charge of 1000 g with five mass fraction ratios (R134a/R23) of 0.75/0.25, 0.70/0.30, 0.65/0.35, 0.60/0.40 and 0.55/0.45. The TOP of EV1 was fixed at 15%. The ambient temperature was kept at 17 ℃. Heat loads of 50 W and 100 W were selected. The experimental data acquisition was carried out when the system operated in a steady state condition. Fig. 7 shows the variation of system COP and compressor power with the mass fraction ratio of R23 MFR23. It can be observed that the system COP decreased as the MFR23 increased and the compressor power yielded an opposite variation tendency. It is possible that when the concentration of R23 increased, both the discharge pressure and the suction pressure of the compressor rose, and then the total mass flow rate increased. This contributed to an increase in compressor power and a reduction in the system COP. When the MFR23 increased from 0.25 to 0.45 at a heat load of 100 W, the system COP decreased from 0.149 to 0.114 because of the increasing compressor power. These results indicated that charging component with low boiling temperature leads to a decrease in the system COP. Furthermore, the system presented higher COP and the compressor power at higher heat load. The compressor power at the heat load of 100 W on average is 33.4 W higher than that at the heat load of 50 W.
Fig. 8 shows the variation of the refrigeration exergy production and system exergy efficiency with the MFR23. The exergy production rate Ee increased as the MFR23. When the MFR23 varied from 0.25 to 0.45, the Ee increased from 31.33 W to 35.85 W. This can be explained as follows. The circulation concentration of R23 increased as the charged concentration increased. More R23 enriched vapor was obtained at the separator and finally entered into the evaporator, which reduced the evaporator outlet temperature at a constant heat load. In this condition, the average refrigeration temperature declined and the exergy production rate increased. Furthermore, an optimal charged concentration of R23 MFR23 existed with respect to the maximum system exergy efficiency, and the system exergy efficiency reached the maximum value of 3.99% at the optimum MFR23 of 0.30 when the heat load was maintained at 100 W. This suggests that the proper selection of the charged concentration of R23 is meaningful to obtain higher system performance. Ejector performance directly influences the energy saving effect of the EARC system. Fig. 9 shows the effect of MFR23 on the pressure lift ratio and entrainment ratio at the heat loads of 50 W and 100 W. The pressure lift ratio rP first decreased and then increased as the MFR23 increased, and the minimum value was obtained an the MFR23 of 0.30, which mainly resulted from the opposite variation tendency of the entrainment ratio . When the MFR23 varied from 0.25 to 0.45 at the Qheater of 100 W, the rP and varied in the ranges of 1.144–1.233 and 0.590–0.989, respectively. The maximum entrainment ratio and the minimum pressure lift ratio occurred at an MFR23 of 0.30. Furthermore, the entrainment ratio at a Qheater of 50 W was larger
than that at the Qheater of 100 W. A possible reason for this is that the circulation concentration of R23 decreased as the heat load increased. Similar composition shift behaviors are be found in the literature [29]. Consequently, the mass fraction ratio of R23 at the condenser outlet decreased as the heat load increased, which contributed to the reduction in the mass flow rate of the secondary fluid ms and entrainment ratio .
3.3 Effect of expansion valve opening In the EARC, one metering valve was installed as the throttling valve before the evaporator inlet. The TOP plays a significant role in refrigerant allocation, evaporation pressure and the overall system performance. In this experiment, heat loads of 50 W and 100 W were maintained. The charged mass fraction of the R23 MFR23 was kept at 30% with a total charge of 1000 g. The ambient temperature was fixed at 17 ℃. The TOP was varied from 10% to 20% to investigate the effect of the TOP on system performance. Fig. 10 shows the variations of the system COP and exergy efficiency with the throttle valve opening. It can be seen that the system COP and the exergy efficiency first increased and then decreased as the increasing TOP. The system COP and the exergy efficiency reached the peak values of 0.144 and 4.05%, respectively, at the TOP of 15%. The underlying cause can be explained as follows: When the TOP increased, the evaporator outlet temperature increased, and the average evaporation temperature first decreased and then increased. In this condition, a maximum refrigeration capacity and refrigeration exergy rate was
obtained at a TOP of 15%. Therefore, the maximum system COP and exergy efficiency existed at the optimal throttle opening. The variations of the TOP on the pressure-lift and entrainment ratios are presented in Fig. 11. It was observed that the ejector performance varied markedly as the TOP increased. When the TOP was increased from 10% to 20%, the pressure lift ratio showed a remarkable decrease from 1.222 to 1.097, the expansion ratio
Pn.i Ps.i
decreased from 5.76 to 4.35, and the entrainment ratio increased from 0.546 to 1.324 at the selected operating conditions. These test results confirm that the pressure lift ratio was closely related to the expansion ratio and the entrainment ratio, and larger pressure lift ratio was obtained at larger expansion ratios and lower entrainment ratios. In addition, from the obtained rP and values, it can be seen that the ejector worked well in the operating conditions, and an acceptable pressure lift ratio value of rP =1.097 could be obtained even at large entrainment ratio of 1.324. 3.4 Effect of heat load The system refrigeration capacity was measured by summing the heat load of electric power Qheater and heat loss occurring in the evaporator box QLoss . The heat load accounts for a major portion of the refrigeration capacity. In this experiment, the variation characteristics of the system performance with refrigeration capacity were investigated by adjusting the heat load from 50 W to 100 W. The range of the heat load was selected according the refrigeration capacity of a low temperature freezer. According to the results discussed already, other operating conditions
were as follows. The mass fraction ratio MFR23 was fixed at the optimum of 30%, and a TOP of the EARC 15% was used. The TOP of EV1 and EV2 in the CARC-a and the CARC-b were fixed at 15% and 70%, respectively, and the ambient temperature was maintained at 17 ℃. Fig. 12 shows the effect of heat load on the system COP and exergy efficiency of the EARC、CARC-a and CARC-b. It can be observed that the COP and exergy efficiency Ex both increased as the heat load Qheater . The increased COP as the
Qheater
resulted from the increased average evaporation temperature. The
refrigeration exergy output increased as the Qheater , which eventually led to an increase in the system exergy efficiency. When the heat load varied from 50 W to 100 W, the COP of the EARC, CARC-a, and CARC-b varied in the ranges of 0.090–0.142, 0.089–0.136 and 0.080–0.135, respectively. The corresponding system exergy efficiency varied in the ranges of 2.63–3.98、2.58–3.65% and 2.05–3.39%, respectively. The EARC exhibited higher system COP and exergy efficiency at the selected heat loads. Compared with the CARC-a and CARC-b at the Qheater ranging from 50 W to 100 W, the EARC could obtain 4.5% and 11.7% higher COP, and 9.6% and 25.1% higher exergy efficiency, respectively. Fig. 13 shows the effect of heat load on the system pressure and the ejector performance. The system pressure includes the discharge pressure Pdis , suction pressure Psuc and ejector outlet pressure Pd.o . It can be seen that the system pressures all increased as the heat load increased. The suction pressure was on average 9.1 kPa lower than the ejector outlet pressure because of the pressure drop in the
evaporation condenser. Furthermore, the entrainment ratio of the ejector reached a maximum value at a heat load of 60 W and then decreased as the heat load increased. A possible reason for the decrease in the can be explained as follows. The evaporator temperature increased as the heat load increased. Consequently, more liquid R134a in the cold section of the evaporator vaporized as gas and the liquid holdup in the evaporator decreased. Thus, the system circulation fraction of R134a increased, and the density of the mixture at the compressor inlet and the total mass flow rate increased. However, the secondary fluid varied slightly as the heat load increased eventually resulted in a decrease in the entrainment ratio. The rP first decreased and then increased when the
Qheater was over 90 W. This is because the discharge pressure, i.e., the inlet pressure of the primary fluid, rose rapidly because of an increase in the total mass flow rate when the Qheater was increased from 90 W to 100 W, as shown in Fig. 13. However, from the changes of the rP and obtained in the operating conditions considered, it can be concluded that the fluctuation of the ejector performance with the variation of the heat load was slight, and the ejector operated steadily at the selected heat loads. 4. Conclusions An experimental investigation indicated that the EARC’s performance was superior to those of the CARCs. Lower evaporation temperatures and higher system COP and exergy efficiency were obtained with the EARC. Highest improvements in the COP and exergy efficiency of the ejector cycle over the two
conventional cycles reached up to 9.6% and 25.1%, respectively. The mass fraction ratio of R23 significantly influences the system performance, and an optimal mass fraction ratio MFR23 of 0.30 was proposed to get the maximum system exergy efficiency. A TOP of 0.15% was achieved in the EARC. The ejector worked well in the selected operating conditions, and the entrainment ratio could reach up to 1.233 at a low R23 charged mass fraction ratio of R23, while the heat load slightly influenced the ejector performance. The EARC has large potential in low-temperature freezing applications. Acknowledgements
This study is financially supported by the National Natural Science Foundation of China (NSFC) under the grant No. 51776147. The authors would like to thank NSFC for the sponsorship.
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Captions Figures Fig. 1 Schematic diagram of the experimental setup Fig. 2 Schematic and P-h diagrams of the EARC Fig. 3 Schematic and dimension diagrams of the ejector Fig. 4 Effect of ambient temperature on the system COP and exergy efficiency Fig. 5 Effect of ambient temperature on the ejector performance and refrigerant mass flow rate Fig. 6 Effect of ambient temperature on the evaporation temperature Fig. 7 Effect of mass fraction MFR23 on the system COP and compressor power Fig. 8 Effect of mass fraction MFR23 on the system exergy production and exergy efficiency Fig. 9 Effect of MFR23 on the ejector performance
Fig. 10 Effect of throttle valve opening on the system COP and exergy efficiency Fig. 11 Effect of throttle valve opening on the ejector performance Fig. 12 Effect of heat load on the system COP and exergy efficiency Fig. 13 Effect of heat load on the system pressures and ejector performance Tables Table 1 Specifications on system components Table 2 Instrumentation and propagated uncertainties in the experimental study Table 3 Equations of the performance indicator and relative uncertainties
Table 1 Specifications on system components Equipment name
Specification
Compressor
Type: Hermetic reciprocating compressor by Secop Displacement: 26.3 cm3, Frequency: 50 Hz
Condenser
Type: Fin tube heat exchanger (aluminum/copper) Dimension: 553 mm (height), 250 mm(weight), 93 mm (thickness)
Evaporator
Type: coil tube with twined with electric hotline Dimension: Φ3/8 inch(diameter),10.6 m (length) Thermal insulation: material: plastic rubber foam insulation tube and pearlite total thickness: 1000 mm
Evaporation condenser
Type: Brazed plate heat exchanger by Kaori Dimension: 311 mm (height), 74 mm(weight), 78 mm (thickness) Heat exchanger area: 0.546 m2
Ejector
Dn=0.8 mm, Dm=3mm, NXP=6 mm
Expansion valve
Type: Needle valve by Hoke Valve port diameter: 1.19 mm
Oil separator
Type: Helical oil separator
Refrigerant
Binary component and purity: R134a, 99.9% R23, 99.9%
Lubricant oil
Type: POE lubricant RL 32H by Emkarate
Table 2 Instrumentation and propagated uncertainties in the experimental study Parameter
Instrument
Accuracy
Full scale
Temperature
T-type thermocouple
±0.5 ℃
-250-350 ℃
Pressure
Pressure transducer
±0.1%
0-0.5 MPa, 0-5 MPa
Mass flow rate
±0.5%
Coriolis flow meter
0-30 kg/h, 0-60 kg/h
Refrigerant weight
Electronic balance
0.1 g
0-35000 g
Compressor and electric
Power meter
±0.5%
0-20 kW
hotline power
Table 3 Equations of the performance indicator and relative uncertainties Parameters
Equations
Pressure lift ratio rP
rP
Pd.o Ps.i
（1）
Entrainment ratio
ms mp
（2）
COP
COP
Refrigeration capacity Qe
Qe Qheater QLoss
Qe Wcom Wf
（3）
（4）
QLoss =UA (Tamb -Te.av）, Te.av
Exergy production rate of refrigeration
Ee [26-27]
System exergy efficiency Ex
Ee Qe (1
Te.o Te.i T ln( e.o ) Te.i
（5）
（6）
T0 ) Te.av
T0 is fixed at 298.15 K
Ex
Ee Wcom Wf
（7）
（8）
Relative uncertainties [28]
Y Y
n
Y i 2 ) Y f ( x1 , x2 , x3 ...xn ) i y
( x 1
,
Highlights >Experimental research was conducted on an ejector enhanced ARC system. >Comparison among ejector-enhanced and conventional ARC cycles was performed. >The improvements in system COP and exergy efficiency reached up to 9.6% and 25.1%. >Optimal charged mass fraction of R134a/R23 and throttle valve opening were obtained.
S1359-4311(17)34695-1 https://doi.org/10.1016/j.applthermaleng.2017.10.053 ATE 11251
To appear in:
Applied Thermal Engineering
Received Date: Revised Date: Accepted Date:
17 July 2017 17 September 2017 8 October 2017
Please cite this article as: T. Bai, G. Yan, J. Yu, Experimental investigation of an ejector-enhanced auto-cascade refrigeration system, Applied Thermal Engineering (2017), doi: https://doi.org/10.1016/j.applthermaleng. 2017.10.053
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Experimental investigation of an ejector-enhanced auto-cascade refrigeration system Tao Bai, Gang Yan , Jianlin Yu Department of Refrigeration and Cryogenic Engineering，School of Energy and Power Engineering, Xi'an Jiaotong University, Xi'an 710049, China Abstract: This study presents the experimental investigation of an ejector-enhanced auto-cascade refrigeration cycle (EARC) with zeotropic refrigerant R134a/R23. Performance comparisons among the EARC and two conventional cycles were conducted at selected operating conditions. The effects of ambient temperature, charged mass fraction ratio of the mixture, throttle valve opening, and heat load on the performance characteristics of the EARC were investigated. The results indicated that the EARC had more advantages in terms of lower refrigeration temperature and higher energy utilization efficiency over the conventional cycles, and the coefficient of performance (COP) and exergy efficiency improvements of the EARC reached up to 9.6% and 25.1%, respectively. The throttle valve opening was optimal with respect to the maximum system exergy efficiency determination. The refrigerant R134a/R23 with the optimal mass fraction ratio of 0.70/0.30 was proposed. Keywords: experimental investigation; ejector; zeotropic refrigerant; auto-cascade
Corresponding author. Tel:+86-29-82668738. Fax:+86-29-82668725. Email: [email protected]
refrigeration; exergy efficiency Nomenclature COP
coefficient of performance
BV
ball valve
D
diameter (mm)
E
exergy flow rate (W)
EV
expansion valve
MF
mass fraction ratio
m
mass flow rate (kg h-1)
NXP
primary nozzle exit position (mm)
TOP
throttle valve opening (%)
P
pressure (MPa)
Qe
refrigeration capacity (W)
Qheater
heat load of electric heater
QLoss
heat loss
rP
pressure lift ratio of ejector
t
temperature (℃)
T
absolute temperature (K)
Te.av
average thermodynamic temperature (K)
UA
overall heat transfer coefficient
W
power (W)
xi
primary measured variables
Y
calculated parameter
Greek symbols
y
absolute uncertainty
Ex
system exergy efficiency
entrainment ratio
expansion ratio
Subscripts amb
ambient
com
compressor
d
diffuser
f
fan
eje
ejector
e
evaporator
i
inlet
o
outlet
m
mixing chamber
s
secoondary fluid
tot
total
1–14
state point
0
reference condition
1. Introduction At present, low temperature and small-size refrigeration systems were drawing
significant attention for medical storage, infrared sensor cooling, and frozen storage. The conventional cascade cycle and the auto-cascade refrigeration cycle (ARC) are the two common methods of achieving low temperature freezing in vapor compression refrigerators. To produce the cryogenic and ultra low cooling temperature over a range of -180 ℃ to -40 ℃ , the conventional cascade refrigeration system needs multiple compressors, heat exchangers and refrigerants [1]. The conventional cascade refrigeration system generally has a more complex system configuration and higher equipment cost in comparison with the auto-cascade refrigerator [2]. ARC systems have certain low-temperature refrigeration advantages, especially for the small-size low-temperature freezers. The ARC technique has been developed over a long period. Podbielniak first patented the design of an ARC system [3] in 1936. The first application of the ARC system in low-temperature refrigeration was by Kleemenko [4] for the liquefaction of natural gas in 1959. In past decades, energy saving and environmental protection have been the main topics of development in the refrigeration industry. There have been significant research efforts dedicated to performance improvement in AR systems, and many advanced refrigeration technologies have been developed, including system configuration optimization, environment friendly refrigerant replacement, and high-efficiency cycle modification. Bichev and Glukov [5] performed optimization research on a nitrogen–hydrocarbon mixture based ARC system with a thermodynamic method, and the specific refrigeration capacity and optimal discharge pressure were obtained by the theoretical calculation. Missimer [2]
compared the advantages and disadvantages of different refrigerant mixtures in ARC systems, and presented the corresponding replacement of CFC refrigerants mixtures by HFC mixtures. Gong et al [6] conducted performance comparisons among three mixed-refrigerant refrigeration cycles with thermodynamic calculations. The results indicated that both cycles (separation and non-separation) provided an approximately equal performance with an appropriate selection of mixture and operating pressures. Nayak and Venkatarathnam [7] experimentally investigated the phase separator dry-out phenomenon of an ARC refrigerator at different mixtures and operating temperatures and found that the cold-end temperature difference of the first heat exchanger can be used to predict the dry-out of the phase separator. Du et al [8] carried out experimental work on cycle characteristics of an ARC system with a R134a/R23 mixture and discussed the effects of the charged concentration and cooling water temperature on system operation. Wang et al [9] investigated the effect of the mixing position of the low-pressure refrigerant on the performance of an ARC system with a rectifying column. Sivakumar and Somasundaram [10] performed an exergetic analysis on a three-stage auto-cascade refrigeration system, and the environmentally friendly refrigerant zeotropic refrigerant R290/R23/R14 with a mass fraction of 0.218:0.346:0.436 was suggested as an alternative refrigerant for a three-stage ARC. Kim and Kim [11] experimentally investigated the operating characteristics of an ARC system with the zeotropic refrigerant coupling of R134a/CO2 and R290/CO2. Zhang et al [12] utilized a fractionation heat exchanger in an ARC system to improve the separation efficiency of R290 and CO 2; the
experimental results showed that the system a fractionation heat exchanger had a lower evaporation temperature. Wang et al [13] added two vapor-liquid separators in an ARC system to improve the separation efficiency of the zeotropic refrigerants and got good oil management; simulation results showed that the proposed cycle configuration was promising for ARC systems of -60 ℃ using binary refrigerants. Xu et al [14] simulated the effects of the expansion valve openings on the variation of the refrigerant composition and set up an experimental system to verify the variation. In addition, Chen et al [15] proposed a novel ARC with an additional heat exchanger to reduce the temperature before the throttle valve and obtained improvements of 37.5% in system coefficient of performance (COP) and 34.5% in exergy efficiency. Sreenivas et al[16] experimentally studied the relationship between the charged and circulation composition in a J-T refrigerator and an auto-cascade refrigerator with a single separator. From the open literature on ARC techniques, it can be concluded that proper refrigerant selection and cycle modification are both efficient methods of improving system performances. However, conventional ARC systems continue to operate with low system performance because of large irreversibility losses in the expansion process. It is a known fact that the ejector is a promising expansion device to recover the expansion-work and improve system performance. The ejector has a simple structure, low manufacturing cost, good expansion work recovery efficiency and pre-compression effect [17, 18]. In the past decades, considerable research has been devoted to the application of the ejector technique in various refrigeration and heat
pump systems [17]. Boccardi et al [19] carried out an experiment on a multi-ejector air to water heat pump, and tested the system and component performance at full and partial operational loads. Wang et al [20] investigated the performance of a novel ejector-enhanced refrigerator-freezer system two-phase-driven ejector. Liu et al [21] investigated three ejector enhanced household refrigerators and reduced the energy consumption by 7.75%. Yan et al [22] proposed an ejector-enhanced ARC system, and the theoretical simulation results illustrated that the system COP could be improved by 8.42–18.02%. Yu et al [23] designed an efficient ARC system in which the ejector utilized the pressure energy of the liquid from the separator to entrain the vapor from the evaporator, and the theoretical investigation showed that the system COP could be improved by 19.1%. Bai et al [24] further modified the ejector-enhanced ARC system with an additional bypass tube and carried out an advanced exergy analysis on the cycle performance. The results indicated that the compressor had the highest improvement priority. Tan et al [22, 23, 25] combined ejector refrigeration and ARC principle to achieve a temperature of -30 ℃ using zeotropic mixture R32/R236fa. Based on the open literature, it can be found that proper cycle modification with an ejector obviously improves the performance of ARC systems. However, the ejector-enhanced auto-cascade refrigeration cycle (EARC) has mainly focused on theoretical analysis. Few experimental studies have been published on the energy-saving effect of the ejector in the ARC system. This study presents an experimental investigation on an ejector enhanced auto-cascade refrigeration cycle
(EARC). The objectives of this study are to present the operating characteristics of the EARC system and evaluate the energy-saving advantages in comparison with the conventional ARC systems. Therefore, the ARC systems equipped with and without ejector were all established in the experimental setup. The influences of the key operational parameters, including the refrigerant charged concentration, ambient temperature, expansion valve opening, and heat load on the system operating characteristics are all discussed. 2. System description 2.1 System description and experimental setup Fig. 1 displays the schematic diagrams of the experimental setups for the EARC system and conventional ARC systems. The experimental setup consisted of a compressor,
a
condenser
with
an
adjustable
fan,
an
evaporator,
an
evaporator/condenser, an ejector, a separator, two metering valves used as throttling devices and two mass flow meters. The common zeotropic refrigerant R134a/R23 was used as the working fluid. The working principle of the EARC system is as follows: the mixture attains a high temperature and pressure vapour after the compressor and then it becomes partially condensed in the condenser. The mixture exits the condenser as a two-phase fluid and enters the separator. Owing to the partial fractionational characteristics of the zeotropic mixture as a two-phase state, the saturated vapor and liquid leaving the separator has different refrigerant compositions. The saturated vapor is enriched with R23 having a low boiling temperature, and the liquid contains more R134a having a high boiling
temperature. The saturated liquid as the primary fluid enters the ejector nozzle to entrain the vapor or two-phase fluid from the evaporator. Then, the two streams mix in the mixing chamber and a pressure rises in the diffuser. The fluid leaves the ejector and enters the evaporation condenser to liquefy the vapor form the separator. Then the liquid coming from the evaporation condenser enters the evaporator after a pressure drop in expansion valve 1 (EV1). The two-phase fluid leaving the ejector becomes superheat gas after the heat exchange with the R23 enriched fluid from the separator and then re-enters the compressor to complete the cycle. The corresponding schematic and P-h diagrams of the EARC system are shown in Fig. 2. To assess the energy-saving effect of the EARC system, two experimental conventional ARC systems (CARC-a and CARC-b) were built as shown in Fig. 1. The paths of the three cycles are depicted with different colors and line types. 4-9 -10-1 , and the CARC-a The EARC system was achieved with path 1-2-3-5-6-7-8-8' 4-9' -10-1 and and CARC-b systems were realized with the paths 1-2-3-5-6-7-8-8'' 4-9' 1-2-3-5-6-7-8-8''' -10-1 , respectively. Five ball valves (BV1–5) were installed in the
system to control the fluid path, and the two CARC systems could be switched according to the test conditions. Detailed switch operation methods are described in the table embedded in Fig. 1. In comparison with the CARC systems, the main cycle feature of the EARC system is that an ejector is used to replace the conventional throttle valve under the separator, and the expansion work can be partly recovered to gain a pre-compression effect. Thus, the suction pressure is increased and the compression ratio can be
lowered, which is beneficial for compressor power reduction and system COP enhancement. Therefore, the EARC system provides higher performance than the CARC systems. 2.2 Experimental instrumentation and data reduction The specifications of the key components including the compressor, condenser, evaporator, separator and ejector are listed in Table 1. A constant speed hermetic reciprocating compressor with a displacement of 26.3 cm3 was used. A spiral-flow type separator was employed at the outlet of the condenser. A high efficiency helical oil separator was installed between the condenser and the compressor. The electric hotline was twined on the surface of the coil tube of the evaporator to measure the refrigeration capacity. To minimize the system heat losses at low refrigeration temperatures in the operating conditions, the thermal insulation of the evaporator and connecting tubes were well designed. For the evaporator, a plastic rubber foam insulation tube of 30-mm thickness was wrapped around the evaporator tube, and then the evaporator tube was embedded in a box of pearlite, which is a good insulation material for cryogenic applications. In this condition, the total thickness of the two heat insulating materials was 100 mm. Other components and connecting tubes were wrapped by plastic rubber foam insulation material with 60-mm thickness. A brazed plate heat exchanger was used as the evaporation condenser. Two sight glasses were set up to monitor the fluid state at the compressor inlet and the evaporator outlet. The refrigerant temperatures along the cycle loop were measured by T-type thermocouples with an uncertainty of ±0.5 °C. The refrigerant pressures were
measured by the pressure transducers with an uncertainty of ±0.1% at full scales of 0.5 and 5.0 MPa. The mass flow rates of the vapor coming from the compressor and separator were measured by two Coriolis flow meters with a reading uncertainty of ±0.5%. The process was completed and the data were logged based on the parameters of the Agilent 34970A data acquisition instrument. The system was tested in a refrigerator test chamber built according to the technical standards of EN ISO 15502: 2005 and GB/T 8059.1~4-1995. Air with a specified temperature ranging in 10–50 ℃ and a relative humidity of 45–75% was provided. The detailed specifications of the system components and measuring instruments were listed in Tables 1 and 2. The schematic and dimension diagrams of the ejector with a fixed geometry nozzle are shown in the Fig. 3. The ejector consisted of a nozzle, suction chamber, mixing chamber, and diffuser. A convergent-divergent nozzle was designed based on the simulation results. The polytetrafluoroethylene gaskets were adopted to prevent refrigerant leakage and to adjust the nozzle exit position (NXP). The ejector performances were evaluated by pressure lift ratio rP and the entrainment ratio . The pressure lift ratio rP is a typical parameter for representing the pressure recovery effect, and the entrainment ratio indicates the mass lifting capability. It is noted that the heat losses were taken into account and calculated with equation (5). The overall heat transfer coefficient UA of 0.417 was measured with the energy balances method. The energy utilization performance of the system was evaluated by the system COP and exergy
efficiency Ex . The corresponding equations of these parameters are given in Table 3. 2.3. Uncertainty analysis To ensure the measurement accuracy of the experimental results, an uncertainty analysis was carried out for the considered parameters. With the accuracy of different sensors listed in Table 2, the relative uncertainties of the calculated parameters including the pressure lift ratio rP , entrainment ratio , exergy production rate Ee , COP and exergy efficiency Ex were obtained by equation (8) as shown in Table 3. The calculation shows that the maximum relative uncertainties of the pressure lift ratio rP , entrainment ratio , COP and exergy efficiency Ex were estimated to be 0.3%,0.5%, 1.4%, and 2.2%, respectively. 3. Results and discussion 3.1 Effect of ambient temperature The room temperature directly determines the separator inlet temperature and is an important parameter that influences the composition separation behaviors and the system operation performance. The effects of ambient temperature on the performances of different systems were studied. In addition, to compare the performances of the CARC-a, CARC-b, and EARC at different ambient temperatures, critical parameters such as the throttle valve opening (TOP) and the charged concentration of the mixture were optimized to obtain the maximum system exergy efficiency. The mass fraction of refrigerant couple R134a/R23 was selected as 0.7/0.3 at a total charge of 1000 g, and the TOP of the EARC system
was fixed at 15%. The openings of EV1 and EV2 of CARC systems were fixed at 15% and 70%, respectively, to obtain the maximum system exergy efficiency. The heat load on the evaporator Qheater was maintained at 100 W because the system was designed for low-temperature freezer applications with the same level of refrigeration capacity. The air temperature in the test room was varied in the range of 15–27 ℃. The condenser fan operated with a fixed power of 50 W to produce a fixed speed. Fig. 4 shows the effect of ambient temperature on the system COP and exergy efficiency. When tamb increased from 15 to 27 ℃，the COP of the EARC and the CARC-b decreased from 0.149 to 0.125 and from 0.139 to 0.119, respectively, and the COP of the CARC-a reached the peak value of 0.1412 at a tamb of 18 ℃. The largest system COP improvements of EARC over CARC-a and CARC-b reached up to 6.0% and 7.9%, respectively. The mechanism of the peak value of the CARC-a at a tamb of 18 ℃is as follows. The circulation concentration of R23 reduced as the increasing ambient temperature. In addition, the suction specific volume of the CARC-a increased with the ambient temperature because of the increase in the suction temperature. This resulted in a reduction in the mass flow rate of the compressor; i.e., mtot decreased from 8.176 kg.h-1 to 7.837 kg.h-1. Consequently, the compressor power of the CARC-a system reduced and the COP increased. When the tamb increased further, the increased suction pressure played the dominate role in the reduction of the specific volume at the compressor inlet. The total mass flow rate increased, the compressor power
increased, and the system COP decreased. Therefore, the COP of the CARC-a increased first and then decreased as the ambient temperature increased. When the ambient temperature increased from 15 to 18 ℃, the total mass flow rate reductions of 0.20 kg.h-1 and 0.31 kg.h-1 occurred in the EARC and the CARC-b, respectively, which were smaller than that of the CARC-a. Thus, the compressor power of the EARC and the CARC-b both increased as the ambient temperature increased. In this condition, only the CARC-a presented the maximum value at an ambient temperature of 18 ℃, and the EARC and the CARC-b showed a linear reduction in COP as the increasing ambient temperature. In addition, the system exergy production rate of the three cycles decreased as ambient temperature increased, and this eventually led to a decreased system exergy efficiency Ex . At the selected operating conditions, the exergy efficiency Ex of the EARC, CARC-a, and CARC-b varied in the ranges of 3.02–4.24%， 3.00–3.87% and 2.60–3.59%, respectively. The system maximum improvements of the exergy efficiency for the EARC over both the CARC-a and CARC-b were 9.8% and 19.3%, respectively. This means that the ejector efficiently utilized part of the expansion work and achieved a relatively high improvement in system performance. However, the energy-saving advantage of the EARC over the CARC-a gradually reduced as the ambient temperature increased. This can be explained as follows. The suction pressure of the EARC was larger at higher ambient temperatures because of the stronger pressure lift effect in this condition.
Thus, the total mass flow rate increased rapidly and led to an increase in the compressor power and a reduction in system COP. Furthermore, the CARC-a outperformed the CARC-b in system COP and exergy efficiency at the given operating conditions. Compared with CARC-a, CARC-b operated with a higher compression ratio. This resulted in the higher compressor power and lower system COP in CARC-b. The ejector performance was closely related to system performance in the EARC. Fig. 5 shows the effect of the ambient temperature tamb on the pressure lift ratio and the entrainment ratio of the ejector as well as the refrigerant mass flow rates. When the ambient temperature increased, the pressure lift ratio rP increased and the entrainment ratio first increased and then decreased. When the ambient temperature increased from 15 ℃ to 27 ℃, the pressure lift ratio rP and the entrainment ratio varied in the ranges of 1.198–1.227 and 0.574–0.624, respectively. These findings are understandable because the discharge pressure increased with an increase in the ambient temperature. This finally resulted in an increase in the pressure drop in the ejector nozzle, and more expansion work could be recovered and a larger pressure lifting effect could be offered by the ejector. On the other hand, the entrainment ratio first increased and then decreased as the tamb increased, and the maximum of 0.624 was obtained at a tamb of 18 ℃. The main reason for this was that the mass flow rate of the secondary fluid increased as the tamb , and the mass flow rate of the compressor and the ejector nozzle first decreased and then increased linearly as shown in this figure, eventually
contributing the peak value of the entrainment ratio at an ambient temperature of 18 ℃ eventually. Fig. 6 shows the effect of ambient temperature on the evaporation temperature of different cycles. The evaporation temperature te , i.e., evaporator inlet temperature, increased as the ambient temperature tamb . The mechanism for this can be explained as follows. When the ambient temperature increases, less high-boiling-temperature refrigerant, i.e., R134a, is cooled as liquid, and the concentration of R23 in the fluid entering the evaporator decreases. In this condition, the evaporation temperature increased. In addition, the EARC exhibited a lower evaporation temperature than that of the CARC-a and CARC-b. When the ambient temperature was increased from 15 ℃ to 27 ℃, the evaporation temperature of the EARC varied from -50.2 ℃ to -40.4 ℃. Compared with the CARC-a and CARC-b, the EARC operated with 2.7 ℃ and 4.1 ℃ lower evaporation temperatures, respectively. This indicates that the use of an ejector in ARC provides a lower evaporation temperature, which is favored for low temperature freezer applications. In addition, the evaporation temperature in the CARC-b was higher than that of the CARC-a. This is explained by the fact that the cooling effect of the fluid leaving the evaporator was not utilized in the evaporation condenser for CARC-b, which resulted in increased liquid temperature before EV1 and higher evaporation temperature. Therefore, it can be concluded that the EARC has more advantages over the CARC-a and CARC-b in terms of lower temperature and higher energy efficiency. 3.2 Effect of charged mass fraction ratio
The component composition of the zeotropic mixture is an important parameter in ARC systems. Circulation concentration is closely related to the charged concentration [16]. To investigate the effect of the charged concentration MFR23 on the performance of EARC, experiments research were carried out at a total charge of 1000 g with five mass fraction ratios (R134a/R23) of 0.75/0.25, 0.70/0.30, 0.65/0.35, 0.60/0.40 and 0.55/0.45. The TOP of EV1 was fixed at 15%. The ambient temperature was kept at 17 ℃. Heat loads of 50 W and 100 W were selected. The experimental data acquisition was carried out when the system operated in a steady state condition. Fig. 7 shows the variation of system COP and compressor power with the mass fraction ratio of R23 MFR23. It can be observed that the system COP decreased as the MFR23 increased and the compressor power yielded an opposite variation tendency. It is possible that when the concentration of R23 increased, both the discharge pressure and the suction pressure of the compressor rose, and then the total mass flow rate increased. This contributed to an increase in compressor power and a reduction in the system COP. When the MFR23 increased from 0.25 to 0.45 at a heat load of 100 W, the system COP decreased from 0.149 to 0.114 because of the increasing compressor power. These results indicated that charging component with low boiling temperature leads to a decrease in the system COP. Furthermore, the system presented higher COP and the compressor power at higher heat load. The compressor power at the heat load of 100 W on average is 33.4 W higher than that at the heat load of 50 W.
Fig. 8 shows the variation of the refrigeration exergy production and system exergy efficiency with the MFR23. The exergy production rate Ee increased as the MFR23. When the MFR23 varied from 0.25 to 0.45, the Ee increased from 31.33 W to 35.85 W. This can be explained as follows. The circulation concentration of R23 increased as the charged concentration increased. More R23 enriched vapor was obtained at the separator and finally entered into the evaporator, which reduced the evaporator outlet temperature at a constant heat load. In this condition, the average refrigeration temperature declined and the exergy production rate increased. Furthermore, an optimal charged concentration of R23 MFR23 existed with respect to the maximum system exergy efficiency, and the system exergy efficiency reached the maximum value of 3.99% at the optimum MFR23 of 0.30 when the heat load was maintained at 100 W. This suggests that the proper selection of the charged concentration of R23 is meaningful to obtain higher system performance. Ejector performance directly influences the energy saving effect of the EARC system. Fig. 9 shows the effect of MFR23 on the pressure lift ratio and entrainment ratio at the heat loads of 50 W and 100 W. The pressure lift ratio rP first decreased and then increased as the MFR23 increased, and the minimum value was obtained an the MFR23 of 0.30, which mainly resulted from the opposite variation tendency of the entrainment ratio . When the MFR23 varied from 0.25 to 0.45 at the Qheater of 100 W, the rP and varied in the ranges of 1.144–1.233 and 0.590–0.989, respectively. The maximum entrainment ratio and the minimum pressure lift ratio occurred at an MFR23 of 0.30. Furthermore, the entrainment ratio at a Qheater of 50 W was larger
than that at the Qheater of 100 W. A possible reason for this is that the circulation concentration of R23 decreased as the heat load increased. Similar composition shift behaviors are be found in the literature [29]. Consequently, the mass fraction ratio of R23 at the condenser outlet decreased as the heat load increased, which contributed to the reduction in the mass flow rate of the secondary fluid ms and entrainment ratio .
3.3 Effect of expansion valve opening In the EARC, one metering valve was installed as the throttling valve before the evaporator inlet. The TOP plays a significant role in refrigerant allocation, evaporation pressure and the overall system performance. In this experiment, heat loads of 50 W and 100 W were maintained. The charged mass fraction of the R23 MFR23 was kept at 30% with a total charge of 1000 g. The ambient temperature was fixed at 17 ℃. The TOP was varied from 10% to 20% to investigate the effect of the TOP on system performance. Fig. 10 shows the variations of the system COP and exergy efficiency with the throttle valve opening. It can be seen that the system COP and the exergy efficiency first increased and then decreased as the increasing TOP. The system COP and the exergy efficiency reached the peak values of 0.144 and 4.05%, respectively, at the TOP of 15%. The underlying cause can be explained as follows: When the TOP increased, the evaporator outlet temperature increased, and the average evaporation temperature first decreased and then increased. In this condition, a maximum refrigeration capacity and refrigeration exergy rate was
obtained at a TOP of 15%. Therefore, the maximum system COP and exergy efficiency existed at the optimal throttle opening. The variations of the TOP on the pressure-lift and entrainment ratios are presented in Fig. 11. It was observed that the ejector performance varied markedly as the TOP increased. When the TOP was increased from 10% to 20%, the pressure lift ratio showed a remarkable decrease from 1.222 to 1.097, the expansion ratio
Pn.i Ps.i
decreased from 5.76 to 4.35, and the entrainment ratio increased from 0.546 to 1.324 at the selected operating conditions. These test results confirm that the pressure lift ratio was closely related to the expansion ratio and the entrainment ratio, and larger pressure lift ratio was obtained at larger expansion ratios and lower entrainment ratios. In addition, from the obtained rP and values, it can be seen that the ejector worked well in the operating conditions, and an acceptable pressure lift ratio value of rP =1.097 could be obtained even at large entrainment ratio of 1.324. 3.4 Effect of heat load The system refrigeration capacity was measured by summing the heat load of electric power Qheater and heat loss occurring in the evaporator box QLoss . The heat load accounts for a major portion of the refrigeration capacity. In this experiment, the variation characteristics of the system performance with refrigeration capacity were investigated by adjusting the heat load from 50 W to 100 W. The range of the heat load was selected according the refrigeration capacity of a low temperature freezer. According to the results discussed already, other operating conditions
were as follows. The mass fraction ratio MFR23 was fixed at the optimum of 30%, and a TOP of the EARC 15% was used. The TOP of EV1 and EV2 in the CARC-a and the CARC-b were fixed at 15% and 70%, respectively, and the ambient temperature was maintained at 17 ℃. Fig. 12 shows the effect of heat load on the system COP and exergy efficiency of the EARC、CARC-a and CARC-b. It can be observed that the COP and exergy efficiency Ex both increased as the heat load Qheater . The increased COP as the
Qheater
resulted from the increased average evaporation temperature. The
refrigeration exergy output increased as the Qheater , which eventually led to an increase in the system exergy efficiency. When the heat load varied from 50 W to 100 W, the COP of the EARC, CARC-a, and CARC-b varied in the ranges of 0.090–0.142, 0.089–0.136 and 0.080–0.135, respectively. The corresponding system exergy efficiency varied in the ranges of 2.63–3.98、2.58–3.65% and 2.05–3.39%, respectively. The EARC exhibited higher system COP and exergy efficiency at the selected heat loads. Compared with the CARC-a and CARC-b at the Qheater ranging from 50 W to 100 W, the EARC could obtain 4.5% and 11.7% higher COP, and 9.6% and 25.1% higher exergy efficiency, respectively. Fig. 13 shows the effect of heat load on the system pressure and the ejector performance. The system pressure includes the discharge pressure Pdis , suction pressure Psuc and ejector outlet pressure Pd.o . It can be seen that the system pressures all increased as the heat load increased. The suction pressure was on average 9.1 kPa lower than the ejector outlet pressure because of the pressure drop in the
evaporation condenser. Furthermore, the entrainment ratio of the ejector reached a maximum value at a heat load of 60 W and then decreased as the heat load increased. A possible reason for the decrease in the can be explained as follows. The evaporator temperature increased as the heat load increased. Consequently, more liquid R134a in the cold section of the evaporator vaporized as gas and the liquid holdup in the evaporator decreased. Thus, the system circulation fraction of R134a increased, and the density of the mixture at the compressor inlet and the total mass flow rate increased. However, the secondary fluid varied slightly as the heat load increased eventually resulted in a decrease in the entrainment ratio. The rP first decreased and then increased when the
Qheater was over 90 W. This is because the discharge pressure, i.e., the inlet pressure of the primary fluid, rose rapidly because of an increase in the total mass flow rate when the Qheater was increased from 90 W to 100 W, as shown in Fig. 13. However, from the changes of the rP and obtained in the operating conditions considered, it can be concluded that the fluctuation of the ejector performance with the variation of the heat load was slight, and the ejector operated steadily at the selected heat loads. 4. Conclusions An experimental investigation indicated that the EARC’s performance was superior to those of the CARCs. Lower evaporation temperatures and higher system COP and exergy efficiency were obtained with the EARC. Highest improvements in the COP and exergy efficiency of the ejector cycle over the two
conventional cycles reached up to 9.6% and 25.1%, respectively. The mass fraction ratio of R23 significantly influences the system performance, and an optimal mass fraction ratio MFR23 of 0.30 was proposed to get the maximum system exergy efficiency. A TOP of 0.15% was achieved in the EARC. The ejector worked well in the selected operating conditions, and the entrainment ratio could reach up to 1.233 at a low R23 charged mass fraction ratio of R23, while the heat load slightly influenced the ejector performance. The EARC has large potential in low-temperature freezing applications. Acknowledgements
This study is financially supported by the National Natural Science Foundation of China (NSFC) under the grant No. 51776147. The authors would like to thank NSFC for the sponsorship.
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Captions Figures Fig. 1 Schematic diagram of the experimental setup Fig. 2 Schematic and P-h diagrams of the EARC Fig. 3 Schematic and dimension diagrams of the ejector Fig. 4 Effect of ambient temperature on the system COP and exergy efficiency Fig. 5 Effect of ambient temperature on the ejector performance and refrigerant mass flow rate Fig. 6 Effect of ambient temperature on the evaporation temperature Fig. 7 Effect of mass fraction MFR23 on the system COP and compressor power Fig. 8 Effect of mass fraction MFR23 on the system exergy production and exergy efficiency Fig. 9 Effect of MFR23 on the ejector performance
Fig. 10 Effect of throttle valve opening on the system COP and exergy efficiency Fig. 11 Effect of throttle valve opening on the ejector performance Fig. 12 Effect of heat load on the system COP and exergy efficiency Fig. 13 Effect of heat load on the system pressures and ejector performance Tables Table 1 Specifications on system components Table 2 Instrumentation and propagated uncertainties in the experimental study Table 3 Equations of the performance indicator and relative uncertainties
Table 1 Specifications on system components Equipment name
Specification
Compressor
Type: Hermetic reciprocating compressor by Secop Displacement: 26.3 cm3, Frequency: 50 Hz
Condenser
Type: Fin tube heat exchanger (aluminum/copper) Dimension: 553 mm (height), 250 mm(weight), 93 mm (thickness)
Evaporator
Type: coil tube with twined with electric hotline Dimension: Φ3/8 inch(diameter),10.6 m (length) Thermal insulation: material: plastic rubber foam insulation tube and pearlite total thickness: 1000 mm
Evaporation condenser
Type: Brazed plate heat exchanger by Kaori Dimension: 311 mm (height), 74 mm(weight), 78 mm (thickness) Heat exchanger area: 0.546 m2
Ejector
Dn=0.8 mm, Dm=3mm, NXP=6 mm
Expansion valve
Type: Needle valve by Hoke Valve port diameter: 1.19 mm
Oil separator
Type: Helical oil separator
Refrigerant
Binary component and purity: R134a, 99.9% R23, 99.9%
Lubricant oil
Type: POE lubricant RL 32H by Emkarate
Table 2 Instrumentation and propagated uncertainties in the experimental study Parameter
Instrument
Accuracy
Full scale
Temperature
T-type thermocouple
±0.5 ℃
-250-350 ℃
Pressure
Pressure transducer
±0.1%
0-0.5 MPa, 0-5 MPa
Mass flow rate
±0.5%
Coriolis flow meter
0-30 kg/h, 0-60 kg/h
Refrigerant weight
Electronic balance
0.1 g
0-35000 g
Compressor and electric
Power meter
±0.5%
0-20 kW
hotline power
Table 3 Equations of the performance indicator and relative uncertainties Parameters
Equations
Pressure lift ratio rP
rP
Pd.o Ps.i
（1）
Entrainment ratio
ms mp
（2）
COP
COP
Refrigeration capacity Qe
Qe Qheater QLoss
Qe Wcom Wf
（3）
（4）
QLoss =UA (Tamb -Te.av）, Te.av
Exergy production rate of refrigeration
Ee [26-27]
System exergy efficiency Ex
Ee Qe (1
Te.o Te.i T ln( e.o ) Te.i
（5）
（6）
T0 ) Te.av
T0 is fixed at 298.15 K
Ex
Ee Wcom Wf
（7）
（8）
Relative uncertainties [28]
Y Y
n
Y i 2 ) Y f ( x1 , x2 , x3 ...xn ) i y
( x 1
,
Highlights >Experimental research was conducted on an ejector enhanced ARC system. >Comparison among ejector-enhanced and conventional ARC cycles was performed. >The improvements in system COP and exergy efficiency reached up to 9.6% and 25.1%. >Optimal charged mass fraction of R134a/R23 and throttle valve opening were obtained.