A new control mechanism for two-phase ejector in vapor compression cycles for automotive applications using adjustable motive nozzle inlet swirl.
Expansion work recovery by two-phase ejector is known to be beneficial to vapor compression cycle performance. However, one of the biggest challenges with ejector vapor compression cycles is that the ejector cycle performance is sensitive to working condition changes which are common in automotive applications. Different working conditions require different ejector geometries to achieve maximum performance. Slightly different geometries may result in substantially different COPs under the same conditions. The ejector motive nozzle throat diameter (motive nozzle restrictiveness) is one of the key parameters that can significantly affect ejector cycle COP. This paper presents a new two-phase nozzle restrictiveness control mechanism which is possibly applicable to two-phase ejectors used in vapor compression cycles. It utilizes an adjustable swirl at the nozzle inlet to control the nozzle restrictiveness on the two-phase flow without changing the physical dimensions of the nozzle geometry. This new control mechanism has the advantages of being simple and potentially less costly. It can possibly also avoid additional frictional losses of previously proposed ejector control mechanisms using an adjustable needle. An adjustable nozzle based on this new control mechanism was designed and manufactured for experiments with R134a. The experimental results showed that, without changing the nozzle geometry, the nozzle restrictiveness on the two-phase flow can be adjusted over a wide range. Under the same inlet and outlet conditions, the mass flow rate through the nozzle can be reduced by 36% of the full load. This feature could be very useful for the future application of ejectors in automotive systems under changing working conditions.
CITATION: Zhu, J. and Elbel, S., "A New Control Mechanism for Two-Phase Ejector in Vapor Compression Cycles for Automotive Applications Using Adjustable Motive Nozzle Inlet Swirl," SAE Int. J. Passeng. Cars - Mech. Syst. 9(1):2016.
Vapor compression cooling cycles deviate from the Carnot refrigeration cycle in several ways, such as isenthalpic expansion of saturated liquid at the condenser outlet and desuperheating of refrigerant vapor at the compressor outlet. Therefore, COPs of vapor compression cooling cycles are always lower than those of a Carnot cycle under the same working conditions. Isenthalpic expansion imposes a two-fold penalty on cycle performance compared with isentropic expansion in the Carnot cycle: the cooling capacity is reduced and the compressor work is increased. Expansion work recovery devices such as ejectors which recover the kinetic energy released during the expansion instead of dissipating it in a throttling process are known to be beneficial to cycle performance. Figure 1 shows the layout and pressure-specific enthalpy diagram of a two-phase ejector cooling cycle first proposed by Gay . In this cycle, high pressure motive flow leaving the condenser enters the ejector through the motive inlet. The motive flow is expanded in the motive nozzle and creates a low pressure zone at the nozzle outlet, which entrains the suction flow from the evaporator. The two streams are mixed in the mixing chamber and kinetic energy is transferred from the motive flow to the suction flow. The mixed fluids leave the ejector through the diffuser. The fluid velocity is reduced in the diffuser which results in recompression of the mixed fluids by converting velocity energy back into pressure energy. Therefore, the ejector diffuser outlet pressure is higher than the suction flow pressure (that is, the evaporator pressure). The two-phase flow then gets separated in the separator. Saturated vapor enters the compressor while saturated liquid gets throttled and is fed into the evaporator via a metering valve. That way, kinetic energy released during expansion is utilized to compress the fluid from the evaporator. As a result, some compressor work is saved while the cooling capacity is increased if the heat rejection capacity remains constant.
Disawas and Wongwises  proposed that, in addition to serving as an expansion device, the ejector can also act as a refrigerant pump for the low-pressure side of the system. The evaporator is therefore fooded with refrigerant and operates as in a liquid recirculation system. Their experimental results showed that the COP of the two-phase ejector refrigeration cycle using R134a was higher than that of the baseline cycle using expansion valve over the whole range of experimental conditions. The maximum improvement achieved was about 13% at low heat sink and heat source temperatures. Liquid recirculation can improve evaporator performance by sending more liquid to the evaporator than is actually evaporated so that dryout in the evaporator can be reduced. It can also improve refrigerant distribution for evaporators with inlet headers by feeding only single-phase liquid to the inlet headers instead of two-phase refrigerant which often results in non-homogeneous distribution of two-phase flow into the parallel channels. Therefore, liquid recirculation can result in higher evaporation pressure, and higher system COP compared to a direct expansion cycle (Lawrence and Elbel ).
Many research efforts have been devoted to R744 transcritical ejector cycles. R744 transcritical cycles usually have larger expansion losses caused by throttling process than subcritical cycles under common working conditions. It is very beneficial to apply ejector to R744 transcritical cycles due to the large recovery potential.
Ozaki et al.  carried out an experiment on an automotive transcritical R744 air conditioning system using an ejector to improve system COP. The experiment showed COP improvements of 20% over a baseline cycle using a conventional expansion valve.
Banasiak et al.  reported a maximum increase in COP of 8% over a baseline cycle with a conventional expansion valve.
Elbel and Hrnjak  and Elbel  experimentally investigated a transcritical R744 system using a refrigerant ejector. They reported that for the test conditions considered the cooling capacity and COP can be simultaneously improved by up to 8% and 7%, respectively. Extrapolation was used to determine that the COP could have been improved by as much as 18% at matched cooling capacities.
Less attention has been given to low-pressure working fluids in the literature for ejector cooling cycles compared with R744 due to their lower work recovery potential. However, ejector cooling cycles using low-pressure refrigerants, such as R134a or R1234yf, can still have noticeable performance improvements.
Early investigation of a two-phase ejector cycle using R134a by Harrell and Kornhauser  predicted a cooling COP improvement of approximately 23% for a typical refrigerating cycle and an ideal ejector. An improvement of 12% could be achieved if the ejector performed as well as typical single-phase ejectors. Ejector performance achieved from later ejector tests corresponded to refrigeration cycle COP improvements ranging from 3.9% to 7.6%.
Lawrence and Elbel  experimentally investigated the performance of an alternate two-phase ejector cycle in which the pressure lift provided by the ejector was utilized in order to provide multiple evaporation temperatures. Low-pressure fluids R134a and R1234yf were used. The ejector cycle showed maximum COP improvements of 12% with R1234yf and 8% with R134a when compared to a two evaporation temperature expansion valve cycle. When compared to a single evaporation temperature expansion valve cycle, the ejector cycle showed maximum COP improvements of 6% with R1234yf and 5% with R134a.
Ejector cycle performance is usually sensitive to working condition changes which are common in automotive systems. Different working conditions require different ejector geometries to achieve maximum performance. Slightly different geometries may result in substantially different COPs under the same conditions. Therefore, it is desirable to introduce an adjustable feature to the ejector so that ejector cycle performance can be optimized under different working conditions, which could make ejector technology more suitable for real world applications (Sumeru et al. ).
The ejector motive nozzle throat diameter is one of the key dimensions that affect ejector cycle COP. It has a direct impact on motive mass flow rate. Other important ejector geometric parameters that affect ejector efficiency and ejector cycle COP include motive nozzle position, constant area diameter of the mixing chamber and suction chamber converging angle. Additional information can be found in Sarkar . One way to adjust the motive nozzle throat diameter in order to optimize ejector cycle performance according to the working conditions is by using a needle which moves back and forth so that the nozzle throat diameter can be varied, as illustrated in Figure 2.
Elbel and Hrnjak  were the first researchers to publish experimental results of introducing a variable two-phase ejector to a transcritical R744 system by installing a needle in the motive nozzle to control the motive nozzle throat diameter. The needle mechanism allowed control of gas cooler high-side pressure, which is an important task for a transcritical cycle to get optimum performance. However, nozzle and ejector efficiencies were impaired because of the additional frictional losses introduced by the needle. It was found that the benefits of high-side pressure control offset the losses in nozzle and ejector efficiencies.
Hu et al.  experimentally investigated the R410A ejector cooling cycle under different conditions with four different ejector motive nozzle throat diameters and a variable ejector with adjustable needle in the motive nozzle. It was shown that the ejector motive nozzle throat diameter has a significant impact on cycle performance and different conditions require different motive nozzle throat diameters. Optimal cycle performance achieved by the variable ejector under different conditions was close to the ejector cycle performance with the most suitable ejector motive nozzle throat diameter among the four. However, compared with the baseline cycle using an electronic expansion valve, the COP of the ejector cooling cycle was only increased slightly. In one condition the baseline was even better than the ejector cycle. This may be because the investigated ejector motive nozzle throat diameters did not include the diameter that would have yielded maximum COP and the variable ejector with adjustable needle has lower nozzle and ejector efficiencies because of the additional frictional losses incurred by the needle.
A variable geometry ejector with adjustable needle in the motive nozzle can optimize ejector cycle performance under different conditions, but this design is complicated and costly, and more frictional losses are incurred because of the additional surface area introduced which results in lower nozzle and ejector efficiencies. This provides motivation to develop a new technology to control the motive nozzle restrictiveness.
In this paper, a novel two-phase nozzle restrictiveness control mechanism, which is called swirl control, is presented. This control mechanism is possibly applicable to the control of ejector cooling cycles. It utilizes an adjustable swirl at the nozzle inlet to control the nozzle restrictiveness on the low vapor quality flow expanded in the nozzle without changing the physical dimensions of the nozzle geometry. The ejector and adjustable nozzle which employ this swirl control mechanism are called swirl ejector and swirl nozzle in this paper. At least one control valve is needed for the implementation of this two-phase nozzle restrictiveness control mechanism. This design has the advantages of being simple, possibly less expensive, and can potentially avoid the additional frictional losses in previously proposed motive nozzle restrictiveness control mechanisms. Controlling the ejector motive nozzle restrictiveness by adjusting the motive nozzle inlet swirl strength can be very useful for future application of ejector cooling cycles in automotive systems that experience a wide range of operating conditions.
In the following sections, the new ejector design and control mechanism will first be introduced in detail. A hypothesis for the influence of nozzle inlet swirl on the nozzle restrictiveness on the two-phase flow will also be provided. An experimental facility for the investigation of the inlet swirl influence on the nozzle restrictiveness will be described and the experimental results as well as preliminary visualization results will be presented and discussed.
SWIRL EJECTOR AND SWIRL CONTROL
A swirl ejector which employs the swirl control to adjust motive nozzle restrictiveness differs from a conventional ejector in that an adjustable swirl is generated at the ejector motive inlet, as is shown in Figure 3. The motive inlet swirl can be created by injecting part of the motive flow tangentially. After injection the tangential flow will be mixed with the axial motive flow. The total mass flow rate passing through the swirl motive nozzle is equal to the sum of mass flow rates entering through the motive nozzle's axial and tangential flow inlets. The ejector cooling cycle shown in Figure 4 that uses a swirl ejector is almost the same as the conventional ejector cooling cycle of Figure 1. The only difference is that the flow at the condenser outlet of the swirl ejector cooling cycle is separated into two streams. One stream enters the swirl ejector through the motive flow tangential inlet and another enters through the motive flow axial inlet. In such a way, a swirl is created at the ejector motive inlet. The ratio of mass flow rates through the two inlets can be adjusted by a valve installed at the motive flow tangential inlet, thereby changing the swirl strength. The pressure drop across the control valve is usually small. It can be assumed that the thermodynamic state at the motive nozzle inlet after the swirl is introduced (downstream of the tangential inlet valve) is the same as the refrigerant state at the condenser outlet.
A hypothesis for the influence of nozzle inlet swirl on the restrictiveness of the two-phase nozzle has been proposed. Rapid expansion of two-phase flow is usually not in thermodynamic equilibrium because of the short time the flow remains in the nozzle and the limited rates of bubble nucleation as well as bubble growth (Henry and Fauske ; Schrock et al. ). Introduction of the swirl can possibly enhance vapor bubble growth during the expansion. The swirl will create a relative motion in radial direction between the vapor bubbles and the surrounding liquid, as shown in Figure 5(b). This relative motion occurs because of the much lower density of the vapor phase compared with that of the liquid phase. In the nozzle's centrifugal acceleration field, the lighter bubbles move towards the nozzle center as a result of being displaced by the denser liquid. The liquid is directed towards the nozzle wall. As the bubbles grow, the latent heat of vaporization is supplied by conduction or convection from the liquid into the bubbles. Therefore, the surrounding liquid experiences a cooling effect (Plesset and Zwick ). When there is relative motion between the liquid and vapor, the surrounding liquid keeps being replaced by fresh, warmer liquid which is easier to be evaporated. Therefore, bubbles can grow faster during the expansion in the nozzle when there is more relative motion (Florschuetz et al. ; Ruckenstein and Davis ). As a result of the increased bubble growth rates, the average density of the two-phase fluids expanded in the nozzle is reduced and the two-phase nozzle restrictiveness can be adjusted. This is achieved without any changes in nozzle geometry, as illustrated in Figure 5.
EXPERIMENTAL FACILITY AND METHODS
In order to investigate the influence of the inlet swirl on the nozzle restrictiveness and visualize the two-phase flow expanded in the convergent-divergent nozzle, a transparent nozzle with controllable swirl at the nozzle inlet has been designed and manufactured, as shown in Figure 6. Important dimensions of the swirl nozzle have been summarized in Table 1 and these dimensions are shown with corresponding letters in Figure 6. It should be noted that the nozzle is only part of a swirl ejector. Experimental investigation of swirl ejectors in ejector cooling systems will be conducted in the future. The swirl nozzle is composed of three components: a tee-shaped part made of brass, a sleeve and a convergent-divergent nozzle, as shown in Figure 7, both made of an optically clear resin called Waterclear Ultra 10122 from SOMOS and manufactured with a Stereo Lithography Apparatus (SLA) from 3D SYSTEMS.
The tee-shaped part serves as the swirl generator. The tangential inlet on the tee allows flow to be injected tangentially and mix with the axial flow, thus creating a swirl. The tee and the nozzle are joined by an NPT thread and sealed by epoxy adhesive. The other NPT thread on the nozzle is for connection with a visualization chamber. The sleeve is designed to provide a smooth transition between the tee part and the nozzle. There is no gap between the sleeve and the nozzle so that no additional disturbance is introduced to the flow. The inner diameter of the sleeve is the same as that of the nozzle entrance. There is a tangential inlet on the sleeve. The tangential inlet on the tee and the tangential inlet on the sleeve are coaxial and have the same inner diameter. For visualization purpose, the flow needs to travel a long distance from the tangential inlet to the starting point of the convergent part of the nozzle. This distance is called swirl decay distance, as shown in Figure 6 with the letter 'g'. Because of the fluid viscosity and turbulence, swirl strength will decay over this distance. For actual applications, such a long distance between the tangential inlet and nozzle convergent part may not be necessary. Therefore, it is desirable to find out the swirl strength at the starting point of the nozzle convergent part so that the true relation between the nozzle inlet swirl strength (at the starting point of the convergent part) and the nozzle restrictiveness can be determined. Future work will be performed to correct for the swirl decay over the swirl decay distance. The sleeve also ensures that the swirling flow travels and decays in a pipe with constant surface properties before it reaches the convergent part of the nozzle, which simplifies future calculation of swirl decay over the swirl decay distance.
The layout of the experimental facility for the swirl nozzle tests is shown in Figure 8. A pumped-refrigerant-loop was used for adjustment of test conditions to investigate the influence of inlet swirl on the two-phase flow expanded in the convergent-divergent nozzle. The working fluid was R134a. A visualization chamber was built from clear PVC pipe. The temperature readings were all obtained from ungrounded Type-T immersion thermocouples. The measured temperatures are regarded as total temperatures. Absolute pressures were read by piezo-electric pressure transducers. Pressures and temperatures at the axial and tangential inlets of the nozzle were measured. The differences between the swirl nozzle axial inlet pressures and tangential inlet pressures were generally within 10 kPa. The axial inlet pressure is assumed to be the nozzle inlet pressure [P.sub.in]. The pressure at the nozzle outlet [P.sub.out] was measured as well. The total mass flow rate [m.sub.total] and the nozzle axial inlet mass flow rate [m.sub.axial] were measured by Coriolis-type mass flow meters. The nozzle's tangential inlet mass flow rate [m.sub.tangential] can be calculated by subtracting the nozzle axial inlet mass flow rate from the total mass flow rate. The ratio of the nozzle tangential inlet mass flow rate to the total mass flow rate was adjusted by two valves. The larger the ratio is, the large the swirl strength is for the same total mass flow rate. In this paper, the swirl strength is defined as the ratio of the nozzle tangential inlet mass flow rate to the total mass flow rate, which can be expressed as shown in Equation 1:
It should be noted that in order to use the full range of swirl control from zero swirl to maximum swirl, two valves were installed in the test rig at both the nozzle's axial and tangential inlets. However, in actual applications, one valve should be sufficient to achieve nozzle restrictiveness control over a suitable range.
Different nozzle inlet pressures were achieved by adjusting the heating water temperature and pump speed which determine the saturation pressure of the refrigerant in the heater. The nozzle outlet pressure can be adjusted by a valve installed downstream of the nozzle. For all experimental results shown in this paper, the liquid flow at the nozzle inlet was subcooled by approximately 0.5 [degrees]C. Admittedly, the calculated subcooling is close to the uncertainty of thermocouple reading and it is imprudent to claim the inlet is subcooled solely based on the thermocouple readings. The sight glass installed at the nozzle inlet allows for visual confirmation that no bubbles are present at the nozzle inlet, which provides a double check for inlet subcooling. Different nozzle inlet states with different levels of subcooling or vapor quality will be the subject of future work.
SWIRL NOZZLE TESTS WITH REFRIGERANT (R134A)
Figure 9 shows the effect of the swirl nozzle outlet pressure on the swirl nozzle total mass flow rate at constant inlet pressures. The nozzle was tested under two inlet conditions: 826 kPa and 32 [degrees]C; 925 kPa and 36 [degrees]C. The inlet swirl strength was varied between 0 and 1. The maximum inlet swirl can be achieved by fully closing the nozzle's axial inlet valve and fully opening the nozzle's tangential inlet valve. The Bernoulli equation has been used to theoretically calculate the incompressible single-phase liquid mass flow rate for the tested nozzle under certain inlet and outlet conditions. Nozzle inlet and outlet diameters in the calculation were 15 mm and 1.7 mm, respectively. The density of the incompressible single-phase liquid in the calculation has been assumed to be the density of the subcooled liquid at the nozzle inlet. The theoretically calculated mass flow rates for the two inlet conditions are shown in Figure 9 with solid/dashed lines. They represent the upper limits for the refrigerant mass flow rate passing through the nozzle for certain inlet and outlet conditions.
It can be observed that the two-phase flow expanded in the nozzle is choked when the nozzle outlet pressure is lower than 550 kPa. When the flow is choked, decreasing the nozzle outlet pressure does not further increase the mass flow rate. Flow choking has been observed for both inlet conditions with or without swirl when the nozzle outlet pressure is sufficiently low. Secondly, the nozzle inlet swirl reduces the total mass flow rate under the same inlet and outlet conditions, which implies larger nozzle restrictiveness. Moreover, as the nozzle outlet pressure approaches the inlet pressure, the curves of total mass flow rate with or without swirl appear to converge to the theoretically calculated limit for incompressible single-phase liquid flow. This indicates that when the nozzle outlet pressure is close to the nozzle inlet pressure, the flow expanded in the nozzle behaves like single-phase liquid. This is possibly because of the small potential to superheat liquid and generate vapor by depressurization. It can also be concluded that the influence of the swirl on the flow diminishes when the fluid behaves more like incompressible single-phase liquid, since the mass flow rate curves with or without swirl under the same inlet and outlet conditions converge to each other. This can be explained by the earlier hypothesis that relates stronger swirls to increased vapor generation during the expansion process. When the flow expanded in the nozzle behaves like single-phase liquid, there is little vapor generation and therefore the proposed swirl control mechanism fails to work as it relies on controlling rate of vapor bubble generation.
Figure 10 shows visualization of the flow expanded in the convergent-divergent nozzle. Figure 10(a) shows that the choked flow in the convergent part of the nozzle is still clear and no bubbles can be observed. It is possible that there are very small bubbles existing in the convergent part of the nozzle which are not visible to the observer without using more advanced visualization techniques. A high speed camera with high resolution and adequate lighting will be used in the future for visualization to capture more intricate flow features. The choked flow creates visible bubbles immediately after passing through the nozzle throat, which indicates that much more vapor generation takes place in the divergent part than in the convergent part of the nozzle. When the nozzle outlet pressure is close to the inlet pressure, for both the convergent and divergent part of the nozzle the flow is clear and no bubbles can be observed, as is shown in Figure 10(b). For the case shown in Figure 10(b), the inlet and outlet conditions were [P.sub.in] = 930 kPa, [P.sub.out] = 885 kPa, [T.sub.in] = 35.9 [degrees]C. This observation supports the previous argument that the flow behaves more like single-phase liquid when the pressure difference between the nozzle inlet and outlet is small.
Figure 11 shows the influence of the nozzle inlet swirl strength on the total mass flow rate through the nozzle at constant inlet conditions. The nozzle outlet pressures were kept below 500 kPa so that the two-phase flow was choked. When the inlet pressure and temperature were 1034 kPa and 40 [degrees]C, respectively, which corresponds to 0.6 [degrees]C subcooling, by changing the inlet swirl strength the total mass flow rate varied from 20.2 g [s.sup.-1] (when there was zero swirl) to 12.9 g [s.sup.-1] (when the swirl strength was maximized). These results again show that nozzle restrictiveness can be changed by nozzle inlet swirl. The stronger the swirl is the larger the nozzle restrictiveness is, since for the same inlet conditions less mass flow rate can be driven through the nozzle. In this case, the mass flow rate can be reduced by 36% with swirl control under the same inlet and outlet conditions, which indicates substantial capacity modulation large enough to be considered for real world applications.
Figure 12 displays that the nozzle inlet pressure can vary in a wide range with different inlet swirl strengths at constant total mass flow rates when the flow is choked. Similarly, the nozzle outlet pressures were kept below 500 kPa to ensure choking of the two-phase flow.
When the total mass flow rate was kept constant at 15 g [s.sup.-1], the nozzle inlet pressure varied from 795 kPa to 1039 kPa when the swirl strength was adjusted from 0.22 to 0.46 which again shows a large range of controllability that can be achieved with the proposed approach. Judging by the almost linear dependence of nozzle inlet pressure on swirl strength, it is reasonable to expect that the control range of the nozzle inlet pressure can be further broadened if the swirl strength is increased. At higher total mass flow rates, the required nozzle inlet pressure to achieve the same mass flow rate increases faster with the swirl strength. Data points for swirl strength between 0.65 and 1 are missing, because when the nozzle axial inlet mass flow rate is small heat losses in the tubes have significant cooling effect on the axial inlet flow even with insulation. In that case it is difficult to keep the same conditions at both nozzle inlets. When the nozzle axial inlet is fully closed, there is no such problem since only the inlet conditions at the nozzle tangential inlet need to be controlled.
In this study, a new nozzle restrictiveness control mechanism called swirl control has been proposed and verified. The approach seems suitable to be applied for the purpose of capacity modulation of two-phase ejector vapor compression cycles, which is very relevant for automotive applications. This control mechanism is simple, possibly inexpensive and can potentially yield better efficiencies than other approaches that attempt to control nozzle flow rates. A hypothesis for the influence of nozzle inlet swirl on the restrictiveness of the two-phase nozzle has been provided. A convergent-divergent nozzle utilizing swirl control was designed and manufactured for experiments with R134a. According to the experimental results, it has been shown that the strength of the nozzle inlet swirl can change the restrictiveness of the two-phase nozzle without changing the nozzle geometry. The nozzle becomes more restrictive as the strength of the swirl increases. The mass flow rate can be reduced by 36% with swirl control under the same inlet and outlet conditions. The control range of inlet pressures and mass flow rates that can be achieved by swirl control appears to be large enough to be applicable for real world applications.
[1.] Gay, N. H., "Refrigerating System," U.S. Patent 1,836,318, 1931.
[2.] Disawas, S. and Wongwises, S., "Experimental investigation on the performance of the refrigeration cycle using a two-phase ejector as an expansion device," International Journal of Refrigeration, 27(6): 587-594, 2004.
[3.] Lawrence, N., and Elbel, S., "Experimental and Numerical Study on the Performance of R410A Liquid Recirculation Cycles with and without Ejectors," 15th International Refrigeration and Air Conditioning Conference at Purdue, West Lafayette, IN, USA", Paper 2187, 2014.
[4.] Ozaki, Y., Takeuchi, H., and Hirata, T., "Regeneration of expansion energy by ejector in C[O.sub.2] cycle," 6th IIR Gustav Lorentzen Conference on Natural Working Fluid, Glasgow, UK, 11-20, 2004.
[5.] Banasiak, K., Hafiner, A., and Andresen, T., "Experimental and numerical investigation of the influence of the two-phase ejector geometry on the performance of the R744 heat pump," International Journal of Refrigeration, 35(6): 1617-1625, 2012.
[6.] Elbel, S., and Hrnjak, P., "Experimental validation of a prototype ejector designed to reduce throttling losses encountered in transcritical R744 system operation," International Journal of Refrigeration, 31(3): 411-422, 2008.
[7.] Elbel, S., "Historical and present developments of ejector refrigeration systems with emphasis on transcritical carbon dioxide air-conditioning applications," International Journal of Refrigeration, 34(7): 1545-1561, 2011.
[8.] Harrell, G. S., and Kornhauser, A. A., "Performance tests of a two-phase ejector," American Society of Mechanical Engineers, New York, NY, United States, 1995.
[9.] Lawrence, N. and Elbel S., "Experimental investigation of a two-phase ejector cycle suitable for use with low-pressure refrigerants R134a and R1234yf," International Journal of Refrigeration, 38: 310-322, 2014.
[10.] Sumeru, K., Nasution, H., and Ani, F. N., "A review on two-phase ejector as an expansion device in vapor compression refrigeration cycle," Renewable and Sustainable Energy Reviews, 16(7): 4927-4937, 2012.
[11.] Sarkar, J., "Ejector enhanced vapor compression refrigeration and heat pump systems - A review," Renewable and Sustainable Energy Reviews, 16(9): 6647-6659, 2012.
[12.] Hu, J., Shi, J., Liang, Y. , Yang, Z., and Chen, J., "Numerical and experimental investigation on nozzle parameters for R410A ejector air conditioning system," International Journal of Refrigeration, 40: 338-346, 2014.
[13.] Henry, R. E., and Fauske, H. K., "The two-phase critical flow of one-component mixtures in nozzles, orifices, and short tubes," Journal of Heat Transfer, 93(2): 179-187, 1971.
[14.] Schrock, V. E., Starkman, E. S., and Brown, R. A., "Flashing flow of initially subcooled water in convergent-divergent nozzles," Journal of Heat Transfer, 99(2): 263-268, 1977.
[15.] Plesset, M. S., and Zwick, S. A., "The growth of vapor bubbles in superheated liquids," Journal of Applied Physics, 25(4): 493-500, 1954.
[16.] Florschuetz, L. W., Henry, C. L., and Khan, A. R., "Growth rates of free vapor bubbles in liquids at uniform superheats under normal and zero gravity conditions," International Journal of Heat and Mass Transfer, 12(11): 1465-1489, 1969.
[17.] Ruckenstein, E., and Davis, E. J., "The effects of bubble translation on vapor bubble growth in a superheated liquid," International Journal of Heat and Mass Transfer, 14(7): 939-952, 1971.
University of Illinois
Creative Thermal Solutions Inc.
The authors would like to thank the member companies of the Air Conditioning and Refrigeration Center at the University of Illinois at Urbana-Champaign for their support.
COP - coefficient of performance [-]
h - specific enthalpy [kJ [kg.sup.-1]]
m - mass flow rate [g [s.sup.-1]]
P - pressure [kPa]
T - temperature [[degrees]C]
axial - axial inlet
in - inlet
out - outlet
tangential - tangential inlet
total - total mass flow rate through the nozzle
Table 1. Swirl nozzle geometric parameters. (a) Nozzle inlet diameter (mm) 15.0 (b) Nozzle throat diameter (mm) 1.0 (c) Nozzle outlet diameter (mm) 1.7 (d) Nozzle convergent part length (mm) 9.9 (e) Nozzle divergent part length(mm) 40.0 (f) Tangential inlet inner diameter (mm) 2.0 (g) Swirl decay distance (mm) 138.0
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|Author:||Zhu, Jingwei; Elbel, Stefan|
|Publication:||SAE International Journal of Passenger Cars - Mechanical Systems|
|Date:||Apr 1, 2016|
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