EXPERIMENTAL VALIDATION AND DESIGN STUDY OF A TRANSCRITICAL CO 2 PROTOYTPE EJECTOR SYSTEM. S. ELBEL, P. HRNJAK (a) ABSTRACT 1.

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1 EXPERIMENTAL VALIDATION AND DESIGN STUDY OF A TRANSCRITICAL CO 2 PROTOYTPE EJECTOR SYSTEM S. ELBEL, P. HRNJAK (a) (a) Department of Mechanical and Industrial Engineering University of Illinois at Urbana-Champaign 1206 W. Green St., Urbana, IL 61801, USA Fax: +1-(217) , pega@uiuc.edu ABSTRACT This paper presents experimental results obtained from a transcritical CO 2 system using a refrigerant ejector instead of an expansion valve. The cooling capacity and COP simultaneously improved by up to 8% and 7%, respectively. Experiments were analyzed to quantitatively assess the effects on system performance as a result of changes in basic ejector dimensions such as motive nozzle and diffuser sizing. Small diffuser angles (5 o ) yielded best results. It was numerically predicted and experimentally verified that like in a conventional transcritical CO 2 system, the highside pressure control integrated into the ejector can be used to maximize the system performance. Due to difficulties in the ejector throat pressure measurements a more practical performance metric was introduced in order to quantify overall ejector efficiencies. According to this definition, the prototype ejector was able to recover up to 35% of the throttling losses. 1. INTRODUCTION In recent years, the re-discovery of CO 2 as a natural refrigerant with promising thermodynamic properties also increased the general interest in refrigeration and air-conditioning systems with attempted expansion work recovery. Among few predecessors, Lorentzen (1983) pointed out early the large throttling losses encountered in transcritical CO 2 system operation. Thus, CO 2 appears to be the ideal fluid for attempted expansion work recovery, because due to large pressure differentials between the high and low pressure sides, the differences between isenthalpic throttling and isentropic expansion become more pronounced. A device, not as well known as the refrigerant expander, which is theoretically capable of approaching isentropic expansion, is the refrigerant ejector, first patented almost a century ago (Gay, 1931). Its principle is based on the isentropic conversion process of pressure related flow work contained in the driving (motive) fluid stream into kinetic energy. The high velocity associated with low static pressure in the throat of the device is utilized to entrain refrigerant exiting the evaporator by momentum exchange. Before reaching the mixing chamber the entrained flow coming from the evaporator is pre-accelerated in the suction nozzle. After both streams are mixed, the high-speed fluid stream is isentropically slowed down which increases the static pressure at the diffuser exit to a higher value than that at the exit of the evaporator. Due to this pre-compression of the evaporator mass flow rate the compressor net work decreases which eventually increases the system performance in terms of COP and cooling capacity. Despite the ejector s intriguing COP improvement potentials, the open literature available for CFC / HCFC / HFC ejector experiments carried out during the last decades reports COP increases in the range of only a few percent, e.g. by Harrell and Kornhauser (1995) for R134a. Problems mainly arise from difficulties in designing ejectors that perform efficiently throughout the desired range of operating conditions. However, the recent introduction of a heat pump water heater with ejector by a Japanese OEM could well indicate that the increased system complexity can be economically justified by the expected performance improvements when CO 2 is used as the working fluid.

2. ANALYSIS A prototype ejector with a variable motive nozzle throat area was designed and built based on model results obtained from a comprehensive system model originally programmed and validated

2 2. ANALYSIS A prototype ejector with a variable motive nozzle throat area was designed and built based on model results obtained from a comprehensive system model originally programmed and validated for performance predictions of a typical mid-sized car. Elbel and Hrnjak (2004a) give more details about the general model features. The model was adapted to the geometries of the microchannel heat exchangers and the hermetic compressor used in this study. The target system was represented by a CO 2 breadboard version of a 5.3kW Environmental Control Unit as used by the U.S. Army. The ejector calculations were based on an iterative approach first devised by Kornhauser (1990) which is summarized in Table 1. The critical mass flow rates required to calculate the motive nozzle throat area and other important ejector dimensions were obtained from experimental findings for flashing flow nozzles under consideration of thermal non-equilibrium effects (Henry and Fauske, 1971). The design of the ejector used in this study is shown in Figure 1. The needle extending into the throat of the motive nozzle allows for the desired high-side pressure control. Table 1: Iterative ejector calculation approach In order to determine the aforementioned efficiencies from experimental ejector data the knowledge of the mixing pressure is required. Even if the measurement of the static pressure of this high speed two-phase flow would be readily available, it would still be difficult to accurately determine specific enthalpy values due to potentially existing thermal non-equilibrium effects. Thus, as demonstrated by eq. (1), a more applied definition of the ejector efficiency is introduced to assess the level of expansion work recovery achieved. Power saved Power savings potential η EJEC = (1) The denominator in eq. (1) can be calculated according to eq. (2), where the upper limit of the integral represents the specific entropy obtained for isenthalpic throttling between the inlet pressure of the expansion device and the evaporation pressure and the lower limit of the integral represents the specific entropy obtained for isentropic expansion across the same pressure difference. The value of the integral can be easily obtained for constant evaporation temperatures. s 2 Power savings potential = m& T ds (2) GC s 1 EVAP

Figure 1: Cut-away view of the modular prototype ejector with high-side pressure control The numerator of eq. (1) could be calculated according to eq. (3).

3 Figure 1: Cut-away view of the modular prototype ejector with high-side pressure control The numerator of eq. (1) could be calculated according to eq. (3). However, this approach proves to be not very practical, because deviations from isentropic processes inside the ejector such as frictional pressure drop or shearing between the streams to be mixed complicate the determination of the integrand as a function of the pressure. PDIFF ( P) Power saved = m& v dp (3) EVAP PEVAP Instead of using eq. (3) the power saved is calculated by noticing that the evaporator mass flow rate undergoes a pressure lift equal to the pressure difference between diffuser and evaporator exit which does not have to be provided by the compressor when the ejector is used instead of an expansion valve. Thus, a theoretical compressor suction state at a lower pressure is computed based on the actual compressor suction condition of the ejector system and the measured pressure lift across the ejector. It is also assumed that the isentropic efficiency of the compressor would not change in comparison to the values determined experimentally. Hence, the compressor power required for the theoretical compression of the measured evaporator flow rate across the measured ejector pressure lift is used as the numerator in eq. (1). The main advantage of the so-defined efficiency metric is that its calculation requires only inputs which can be obtained from standard measurements having high accuracy. 3. EXPERIMENTAL FACILITY The experimental setup is shown in Figure 2. The facility consisted of two closed-loop wind tunnels designed to accommodate the microchannel gas cooler and evaporator. The remainder of the refrigeration system was installed in between the ducts to enable realistic piping lengths. By replacing the ejector with an expansion valve and by re-routing a limited number of refrigeration lines, a comparison system could be constructed. Thus, a comparable baseline system was readily available acting as a benchmark against which the ejector system performance could be measured in an accurate manner. Air entering the heat exchangers was pre-conditioned by a combination of electric heaters, coolers and steam humidifiers enabling a wide range of possible test conditions. Two independent energy balances were available, namely for the air and the refrigerant sides of the system. On the air-side, the cooling capacity was determined from flow nozzle measurements in combination with humidity and Type-T thermocouple readings. The evaporator capacity was also determined from a refrigerant mass flow meter measurement and the specific enthalpies across the evaporator.

4 [Abbreviations] ABL : Air blender, ACU : Air-cooled condensing unit, AIR : Air, BL : Blower, BV : Ball valve, BVE : Building ventilation exhaust, CC : Cooling coil, CP : Charge port, CPR : Compressor, CR : Condensate removal, CRH : Capacitive relative humidity sensor, CWC : Chilled water coil, DG : Damper gate, DP : Differential pressure transducer, EJC : Ejector, ERH : Electric resistance heater, EVP : Evaporator, FAI : Fresh air intake, FN : Flow nozzle, FS : Flow straightener, GC : Gas cooler, HCO : Humidity controller (PID), HU : Humidifier, IHX : Internal heat exchanger, LFV : Liquid flow valve, MFM : Mass flow meter, MV : Metering valve, OFV : Oil flow valve, OS : Oil separator, P : Absolute pressure transducer, PSV : Pressure transducer switching valve, PU : Pump, SG : Sight glass, SV : Solenoid valve, T : Type-T thermo-couple, TCG : Thermo-couple grid, TCO : Temperature controller (PID), VFD : Variable frequency drive, VFV : Vapor flow valve, VLS : Vapor-liquid separator, WR : Water reservoir, WT : Watt transducer Figure 2: Experimental test facility for CO 2 ejector testing Further, the refrigerant based energy balance for the expansion valve system assumed purely adiabatic heat exchange in the internal heat exchanger. In case of the ejector two independent energy balances could only be obtained for superheat at the evaporator exit. For this situation, the independently determined evaporator capacities typically agreed within 5% of each other. The electric power supplied to the hermetic compressor was measured with a watt transducer. Thus, the calculated COPs were not solely based on compressor shaft power, but rather included the

5 inefficiencies of the motor and the variable frequency drive. The power required to drive the blowers was not taken into account in the COP calculations. 4. RESULTS AND DISCUSSION 4.1 Improved Cooling Capacity and COP The model results shown in Figure 3 were obtained for an outdoor temperature of 45 o C and an indoor temperature of 27 o C at 30% relative humidity. The model predicted that like in a conventional transcritical expansion valve system (Inokuty, 1928 or more recently Park et al., 1999) the high-side pressure of the ejector system can also be used to maximize the system performance. Furthermore, it can be seen from the model predictions that for identical IHX effectiveness, the values of the COP maximizing high-side pressures of the ejector system are similar to those obtained with an expansion valve system. A standard definition of the IHX effectiveness was used according to Incropera and DeWitt (1996). It also appears that the ejector system can have a less effective IHX to achieve a COP similar to that of expansion valve system having a more efficient IHX. Figure 3: Predicted COP-maximization for expansion valve and ejector systems having different IHX effectiveness The same test condition was investigated experimentally and the corresponding results are shown in Figure 4. The air flow rates were 2040m 3 /h and 680m 3 /h for the outdoor and indoor sides, respectively. The 26cm 3 displacement compressor was used at a constant (nominal) speed of 1800rpm throughout all tests. The diffuser angle was 5 o. The oil separator upstream of the evaporator (outlined in Figure 2) was used to reduce the amount of oil circulating through the evaporator. With the help of the metering valve, the evaporator exit quality was adjusted to values between 0.9 and 1.0. The shim thickness between the motive nozzle and mixing section modules determined size of the suction nozzle and it was kept at 5.35mm for all tests reported in this paper. Variations of the evaporator oil circulation rate, the evaporator exit condition as well as the suction nozzle size were all experimentally found to have significant impact on the system performance; however, a detailed description of the study regarding these parameters is outside the scope of this work. The experimental results confirmed the basic trends predicted by the model, i.e. both the cooling capacity and COP could be maximized by variation of the high-side pressure. Furthermore, the ejector high-side pressure for maximum COP were lower than that for maximum capacity and for a more effective IHX these pressures shifted to lower values. For each of the curves, the maximum

6 possible high-side pressure was limited by the maximum compressor discharge temperature (150 o C). Especially for the IHX having 60% effectiveness it can be seen that it was possible to operate the ejector system under a wider range of test conditions than the expansion valve baseline system, because the reduced compressor ratio of the ejector system also reduced the discharge temperature at the compressor exhaust. Figure 4: Cooling capacity and COP of expansion valve and ejector systems with varying IHX effectiveness (experimental data) The COP maximizing high-side pressure of the ejector system ( 11000kPa) was similar to that of the expansion valve system, which was also well predicted by the model. Also, the experiments confirmed that the highest COP was achieved with ejector and a high IHX effectiveness. For the test condition considered, the ejector improved the cooling capacity and the COP by up to 8% and 7%, respectively. It should be noticed that the system capacity and COP increased at the same time, because the hermetic compressor used showed a significant variation of isentropic efficiency as a function of the motor speed. It is believed that this resulted from variations of the electric motor efficiency when used at off-design speeds. For reasons of keeping the compressor motor efficiency constant, it was not attempted to run tests with matched cooling capacities. Instead all tests were carried out with the same motor speed. Also, the improvements observed in terms of capacity and COP are believed to only partly arise from the primary ejector benefits which result from the reduced compressor pressure ratio and its related increase in isentropic efficiency. Further, the ejector setup improves the evaporator performance based on a number of secondary benefits which result from providing low quality CO 2 to the coil. Elbel and Hrnjak (2004b) experimentally found these improvements in forms of a more homogenous refrigerant distribution in the microchannel evaporator, an increase of the average evaporative CO 2 heat transfer coefficient as well as a reduction of the refrigerant-side pressure drop. 4.2 Ejector Efficiency The ejector pressure lift is shown as function of the power savings potential in Figure 5. The ejector efficiencies given were calculated according to eq. (1). The graph on the left was plotted for the aforementioned test conditions. As expected, the IHX with the lowest effectiveness yielded the highest expansion work recovery potential, because it resulted in the highest values of specific enthalpy at the ejector s motive nozzle inlet. The highest ejector efficiencies resulted in a recovery of more than 30% of the available potential. At these conditions, the motive nozzle area was largest, resulting in a large motive flow rate at a relatively low high-side pressure. For a constant temperature at the inlet of the motive nozzle, test conditions with lower high-side pressures resulted

7 in higher specific enthalpies at the motive nozzle inlet from which higher kinetic energies could be extracted within the ejector. However, for a given power savings potential, the system with the most effective IHX achieved the highest pressure lifts resulting in maximum COPs. Figure 5: Ejector pressure lift and efficiency as a function of the power savings potential for different IHX effectiveness, ambient conditions and diffuser angles The graph depicted on the right of Figure 5 shows the influences of a variation in outdoor temperature and different diffuser angles. When the temperature was reduced from 45 o C to 35 o C, the power savings potential generally decreased as well, which again could be explained by the reduced specific enthalpy of the fluid entering the motive nozzle. The data also indicated that the pressure lift capabilities of the ejector increased with smaller diffuser angles for the design explored at that time. For an identical power savings potential, the longest diffuser having the smallest angle produced the highest pressure lift. It is believed that boundary separation along the diffuser wall occurred earlier for larger angles, creating vortices which eventually decreased the diffuser efficiency. A design trade-off should exist between boundary separation induced efficiency reduction for large diffuser angles and increased frictional pressure drop for very long diffusers having very small angles. 5. CONCLUSIONS Experimental data were presented for a transcritical CO 2 ejector system and compared to a conventional expansion valve system. Due to the use of the ejector the COP and cooling capacity were simultaneously improved by up to 7% and 8%, respectively. Up to 35% of the available power savings potential were recovered by the ejector in accordance with a newly defined efficiency metric. The data confirmed that the ejector principle works well with CO 2. It further validated the prototype design obtained from a simulation model. The strong influence of the effectiveness of the IHX on the ejector performance was studied and it was found that the maximum system performance was achieved with a highly effective IHX integrated into the ejector system. However, from a practical standpoint it was observed that an ejector system could have a reduced IHX effectiveness (shorter, easier to integrate and less expensive) and still achieve capacities and COPs that were comparable to that of the expansion valve baseline system. Furthermore, this study revealed that high ambient temperatures and high refrigerant mass flow rates as well as small diffuser angles were beneficial to an efficient ejector system operation.

8 ACKNOWLEDGEMENTS The authors thankfully acknowledge the support provided by the Modine Manufacturing Company, U.S. Army RDECOM and the DaimlerChrysler AG. NOMENCLATURE COP coefficient of performance (-) Subscripts h specific enthalpy (kj/kg) DIFF diffuser IHX internal heat exchanger EJEC ejector m& mass flow rate (kg/s) EVAP evaporator OEM original equipment manufacturer GC gas cooler P pressure (kpa) isen isentropic r ratio (-) s specific entropy (kj/kg-k) Greek T temperature ( o C) α angle u velocity (m/s) η efficiency v specific volume (kg/m 3 ) x vapor quality (-) REFERENCES 1. Elbel S.W., Hrnjak, P.S., 2004, Effect of Internal Heat Exchanger on Performance of Transcritical CO 2 Systems with Ejector, Proc Int. Refrig. Conf. Purdue, Paper R Elbel S., Hrnjak, P., 2004, Flash Gas Bypass for Improving the Performance of Transcritical R744 Systems that use Microchannel Evaporators, Int. J. Refrig., 27(7): Gay, N.H., 1931, Refrigerating System, U.S. Patent No. 1,836, Harrell, G.S., Kornhauser, A.A., 1995, Performance Tests of a Two-Phase Ejector, Proc. 30 th IECEC, ASME: Henry R., Fauske, H., 1971, The Two-Phase Critical Flow of One-Component Mixtures in Nozzles, Orifices, and Short Tubes, J. Heat Transf., 93(2): Incropera, F.P., DeWitt, D.P., 1996, Fundamentals of Heat and Mass Transfer, 4 th ed., John Wiley & Sons, New York, 886 p. 7. Inokuty, H., 1928, Graphical Method of Finding Compression Pressure of CO 2 Refrigerating Machine for Maximum Coefficient of Performance, Proc. 5 th Int. Congr. Refrig., IIR: Kornhauser, A.A., 1990, The use of an Ejector as a Refrigerant Expander, Proc USNCR/IIR-Purdue Refrigeration Conference, USNCR/IIR: Lorentzen G., 1983, Throttling the Internal Haemorrhage of the Refrigeration Process, Proc. Inst. Refrig., Vol. 80: Park, Y.C., Yin, J.M., Bullard, C.W., Hrnjak, P.S., 1999, Experimental and Model Analysis of Control and Operating Parameters of Transcritical CO 2 Mobile Air Conditioning System, Vehicle Thermal Management Systems (VTMS4) Conf. Proc., Professional Engineering Publishing Ltd:

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