Boewe, D. E.

Bullard, C. W.

Hrnjak, P. S.

McEnaney, R. P.

Park, Y. C.

Yin, J. M.

English

Purdue E-Pubs

Purdue UniversityPurdue e-PubsInternational Refrigeration and Air ConditioningConference School of Mechanical Engineering1998Experimental Comparison of Mobile A/C SystemsWhen Operated With Transcritical CO2 VersusConventional R134AR. P. McEnaneyUniversity of IllinoisD. E. BoeweUniversity of IllinoisJ. M. YinUniversity of IllinoisY. C. ParkUniversity of IllinoisC. W. BullardUniversity of IllinoisSee next page for additional authorsFollow this and additional works at: http://docs.lib.purdue.edu/iraccThis document has been made available through Purdue e-Pubs, a service of the Purdue University Libraries. Please contact epubs@purdue.edu foradditional information.Complete proceedings may be acquired in print and on CD-ROM directly from the Ray W. Herrick Laboratories at https://engineering.purdue.edu/Herrick/Events/orderlit.htmlMcEnaney, R. P.; Boewe, D. E.; Yin, J. M.; Park, Y. C.; Bullard, C. W.; and Hrnjak, P. S., "Experimental Comparison of Mobile A/CSystems When Operated With Transcritical CO2 Versus Conventional R134A" (1998). International Refrigeration and Air ConditioningConference. Paper 402.http://docs.lib.purdue.edu/iracc/402AuthorsR. P. McEnaney, D. E. Boewe, J. M. Yin, Y. C. Park, C. W. Bullard, and P. S. HrnjakThis article is available at Purdue e-Pubs: http://docs.lib.purdue.edu/iracc/402EXPERIMENTAL COMPARISON OF MOBILE A/C SYSTEMS WHEN OPERATED WITH TRANSCRITICAL C02 VERSUS CONVENTIONAL R134A R.P. McEnaney, D.E. Boewe, J.M.Yin, YC. Park, C.W. Bullard, P.S. Hmjak, Air Conditioning and Refrigeration Center (ACRC) University oflllinois, 1206 W. Green St. Urbana, IL 61801, USA ABSTRACT This paper describes the experimental comparison of performance of prototype transcritical C02 system for mobile ale to it's Rl34a counterpart, unit for Ford Escort that was chosen as a representative for the US, European, and Japanese markets. The paper presents the design of experiments and facilities for such comparisons and preliminary results. INTRODUCTION One objective of this project is to provide an experimentally verified basis for calculation of TEWI (Total Equivalent Warming Impact- see Sand at al. 1997), particularly for R744 that are very sparse like Pettersen et al. (1997a) and (1997b). Test matrices are also defined for the purpose of developing component and system simulation models, as well as supporting data-to-data comparisons at normal, seasonal and extreme operating conditions. The comparison of Rl34a and C02 system is based on equal heat exchanger core volumes because cost data were unavailable. Component volumes were constrained to be equal; the weight was assumed be similar due to same materials used for the design. The face areas are kept equal. More details could be found in Table 1. Volumetric flow rates of air and air side pressure drops are maintained almost the same and are shown in Table 2. EXPERIMENTAL FACILITIES Separate environmental chambers have been constructed for each heat exchanger, each containing wind tunnel with separate piping for the two refrigerants. Three energy balances are obtained for each heat exchanger: air-side, refrigerant-side and room calorimetry. Special care was taken to develop test facilities that will produce accurate data in wide operating ranges, primarily in steady-state but also in transient (mostly cycling) conditions. Three independent methods are used instead of the two required by all applicable standards, not only to facilitate determincrtion of system capacities, but also to have two whenever refrigerant calorimetry is not reliable due to two-phase exit. This occurs mostly during transients, and with constant area expansion devices. Three independent procedures also improved our ability to troubleshoot early tests. The test facility consists of two environmental chambers (see Figure 1) that can maintain outdoor and indoor temperature within ±0.5°C and absolute humidity ±2%. A variable speed wind tunnel in each chamber simulates the range of operating conditions encountered in real applications, and allows measurement of air-flow rates within ±I%. Coriolis mass flow meters, together with immersion thermocouples and electronic pressure transducers upstream and downstream of every component yield refrigerant-side capacity determinations repeatable within ±I%. Room calorimetry is the most accurate: the walls are made of 30cm thick polyurethane. There are five thermocouples on both sides of each wall, floor, and ceiling of each environmental chamber. Transmission losses are carefully calibrated so that error is within ±0 .1 %, all dry energy inputs (electric) are measured within ±0 .2%. Special care is taken to ensure uniformity of the temperature and velocity profiles at the inlet to the heat exchangers, and representcrtive reading of the exit air temperatures and humidities. Test results show agreement between the independently determined capacities to be within ±5%, primarily due to uncertainties in air-side calorimetry. Figures 2 and 3 show the differences in heat balances determined on air and refrigerant side and for the environmental chamber. 145 RH Condenser Chamber Dpea Evap Evaporator Chamber Hu :::8 RH Gas Cooler Chamber Dpea Evap Evaporator Chamber Hu :::8 Figure Ja. Test facility for R-134a baseline system Figure lb. Test facility for R744 system B - Blower, C - Compressor , CC - Cooling Coil, CH - Glycol Chiller, Cond - Mobile Condenser, Dp -Differential Pressure Transducer, Evap -Mobile Evaporator, F -Fan, FS -Flow Straightener, GC - Gas Cooler, H - Heater, Hu - Humidifier, IC - Residential Indoor Coil, mg - Glycol Mass Flow Meter, ml - Suction Accumulator Liquid Mass Flow Meter, mo - Oil Mass Flow Meter, mr - Refrigerant Mass Flow Meter, Mtr -Motor, N - Nozzle, OC -Residential Outdoor Coil, Or - Orifice Tube, P - Pressure Transducer, RH - Relative Humidity Probe, S - Separator, SA - Suction Accumulator, Sc - Condensate Scale, SLHX - Suction Line Heat Exchanger, Sp - Speed Controller and Tachometer, T - Thermocouple, TC - Temperature Controller, TG -Thermocouple Grid, Tor -Torque Transducer, W - Watt Transducer, XV -Metering Expansion Valve Indices: a - air, c - condenser, cp - compressor, e - evaporator, g - glycol, i - inlet, n - nozzle, o - outlet, r -refrigerant, sh - suction line heat exchanger Compressor is placed in the separate chamber that simulates temperature conditions in the engine compartment. A torque meter is placed between the compressor and the clutch to measure power at the compressor shaft, excluding belt and clutch losses. The suction accumulator is located in the same enclosure, as it would be in the real system. Lines are as short as possible to be representative for real system even in cycling mode. Heat exchangers in the baseline units are high performance types commonly used in conventional systems. Evaporator has brazed aluminum plates and fins while the condenser is an aluminum tube and plate fin type (see Table 1). Due to higher pressure, heat exchangers for C02 (R744) should have either small diameter tubes or thicker walls of the same materiaL We selected micro-channel heat exchangers with specially designed headers to sustain higher operating pressures. Baseline Rl34a system is typical, used in Ford Escorts. It has orifice tube and the suction accumulator. It was not altered in any way except slight modification in piping to fit it into the test facility. The additional charge is added to ensure proper operation following the manufacturer's procedure. 146 12 ~ Q) 0 10 -g~ o 3000rpm, no cycle • 3000rpm, cycle 1800rpm, no cycle 1800rpm, cycle o 950rpm, no cycle • 950rpm, cycle +6% -6% ,_ ('0 ii:.O 8'-------------' ('0 >. C1) E' E ~ £; C1) 6 0 .... ('0 C1) c....c ~ ~ 4 ..c: 0 ..c: ~ 2 2 4 6 8 10 12 14 Capacity measured with air enthalpy method(kw) • 3000rpm, no cycle • 3000 rpm, cycle • 1800rpm, no cycle , • 1800rpm, cycle • 950rpm, no cycle 950rpm, cycle ••• _Q~oo· -:fJ.?~c-· +10"·'·<·6· • --6% -1"0•,4 · .. ··· 0 .. -~> 2 4 6 8 10 12 14 Capcacity mesured with air enthalpy method(kW) Figure 2. Comparison of air-chamber heat balances Table 1: Comparison of components for mobile system Figure 3. Comparison of air-chamber heat balances System Refrigerant RJ34a R744 Type Ford Escort PrototvDe Compressor: 'JYPe Reciprocating Reciprocating Displacement 155 cm3 20.7 cm3 Expansion device Orifice tube Manual valve/back pressure valve Description Wavy AI fins, round AI tubes, Microchannel, brnzed AI tubes, 3-pass, 21 pass, OD= 6mm parallel flow Mass 2.0kg 2.3 kg Condenser/ Face area 36.1 x 54.4 = 1964 cnl 36.8 x 53.0 = 1950 cm2 Core depth 2.2cm l.65cm Gas cooler Core volume 4320 cm3 3320cm:~ Air side surface 1.2nr 5.2m2-Refrigerant side 0.40m2 0.49 m2 surface area Description Brazed AI plate (drawn cup, Microchannel, brazed, AI, laminated), 4-pass 17 plates 7-nass, oornnel flow Mass l.8kg 2.2kg Evaporator Face area 18.4 x 22.0 = 405 cm2 18.2 x 22.4 = 408 cm2 Core depth 9.2cm 9.1 em Core volume 3720 err? 3710 cm3 Air side surface 3.5nr 4.2m2 Refrigerant side 0.55 m2 0.66m2 surface area Suction line Description Two aluminum coaxial tubes, suction vaoor in annulus, counter-flow heat Mass 0.7kg exchanger Length 1.5m Table 2 Pressure drops on the mr and refrigerant szdes Heat exchanger Test conditions Rl34a R744 Air side Ap [Pa] Condenser~s cooler 1250cfm 74 61 Evaporator 250cfm drY coil 114 118 Refrigerant side dp Condenser/gas cooler H3 87 545 (kPa] Eva.pQ_rator H3 65 273 147 R744 system is shown in Figure lb. It has suction accumulator (with sight glasses) and suction line heat exchanger. Expansion device was a manually set back pressure valve. Oil separator was used to reduce influence of increased circulation of lubricant required for the early prototype of R 7 44 compressor. TEST MATRIX The test matrix was designed to serve our major objectives: to provide data for a TEWI analysis, to analyze system and component perfonnance in order to improve component design and system operation; and to verify the models. The first objective requires identification of representative operating conditions, and yields a matrix that is based on empirical assumptions of realistic operating conditions. The second group of objectives (design analysis and model verification) requires a larger (more data points and test regimes), meticulously designed test matrix. A summary of the test conditions is presented in the Table 3. There are three compressor speeds in the Table 3 (I - idle, M - intermediate and H -highway speeds) and three outdoor temperatures (21.1, 32.2, and 43.3 °C). The relation between compressor speed and condenser air flow rate is given in the footer of the Table 3. It is assumed that the inlet air temperature to the condenser is 11.1 °C (20~ higher than the ambient temperature due to recirculation of air and the influence ofthe hot pavement (Bhatti, 1997). Air temperature and flow rate in the passenger compartment are a function of the system capacity and passenger comfort: at higher outdoor temperatures the indoor temperature are generally higher and the passenger increases the flow rate of air. Table 3. Test matrix of representative operating conditions for mobile ale Test no IM26 128 123 M23 ll3 M12 I6 M3 I19jM10 117 jM8IIll I M5 Condenser inlet air 15.5 21.1 32.2 21.1 43.3 32.2 54.4 43.3 32.2 43.3 temperature, rc] Passenger compartment 21.1 26.7 32.2 26.6 21.1 126.6 air temptaauan:;, rq Evaporator air flow rate, [m3/min] 2.832 4.956 7.08 Relative humidity, [%] 40 (I) .. - - .~ I . Idle condition. compressor speed - 950 rpm, condenser au flow rate - 22.67 m /mm (z) M: Intermediate speed: compressor speed= 1800 rpm, condenser air flow xate = 26.90 m3/min (3) H : Highway condition: compressor speed= 3000 rpm, condenser air flow rate= 35.40 m3/min The conditions in the Table 3 do not prescribe either the evaporation or condensation temperatures. They are established by the load/capacity equilibrium. The purpose of the test matrix in Table 3 is to reduce number of data points to representative operating conditions for typical mobile ale system. Another, more complete set of test conditions is designed to obtain data for model verification and better understanding of component performance (Yin at al. 1998). Operating parameters are varied over a wide range: compressor speed 950-3000rpm, air flow through condenser/gas cooler 22-36m3/min at temperatures 15-55°C, air flow through evaporator 2.8-7m3/min at temperatures 21-43°C, and relative humidity 25-70%. The purpose of these tests is to verify the models of heat exchangers and the systems over the wide range. SOME EXPERIMENTAL RESULTS The experimental data are presented in Table 4 following the test matrix shown in Table 3. Table lists test designation (see Table 4 and 4), refrigeration capacity, coefficient of performance, condensate collected, evaporation temperature, temperature of air leaving evaporator, isentropic and volumetric efficiencies of the compressors, discharge/suction pressure ratio, and on time ratio for the cycling data points. All data are either measured or determined directly from our measurements. Data shown are those that were available at the time of printing. Capacities shown are heat exchanged on the evaporator surface. Coefficient of performance is determined as the ratio of refrigeration capacity and power on the compressor shaft. Compressor torque was measured between compressor and the clutch. All data are either measured directly or determined from our measurements. 148 All data shown are for the same speed of the Rl34a and R744 compressors. Please note that condensate removal (latent load) is very similar for both refrigerants. It is related to very close evaporating temperatures. R744 compressor operates at much lower pressure ratios and consequently both isentropic and volumetric efficiences are comsiderably higher. Table 4. Test results of Rl34a and R7 44 mobile ale system at representative operating conditions Test Refri- Q COP Con d. Tevap Teao Isentr. Vo1um. Pd/Ps Ontime gerant [kW] [-] (g/s] [OC] rq eff [-] eff. [-] [-] fraction M3 Rl34a 4.194 1.58 0.631 4.8 14.9 0.59 0.56 7.1 1.0 R744 4.672 1.59 0.666 8.3 12.8 0.76 0.79 2.7 1.0 M12 Rl34a 2.548 1.82 0.318 1.5 9.7 0.64 R744 2.532 1.93 0.253 4.5 8.4 0.57 M23 Rl34a 1.107 2.38 0.079 2.7 6.8 0.24 R744 1.149 3.17 0.071 2.2 5.8 0.16 M26 Rl34a 1.104 2.51 0.077 2.7 7.2 0.09 R744 1.187 4.00 0.087 2.5 5.9 0.15 !6 R134a 2.939 1.75 0.307 ll.5 18.1 0.58 0.54 5.7 1.0 R744 2.795 1.57 0.229 13.7 17.5 0.70 0.72 2.6 1.0 113 Rl34a 2.475 1.90 0.280 3.1 9.4 0.60 0.59 5.5 1.0 R744 2.624 1.84 0.299 4.8 8.6 0.79 0.79 2.7 1.0 123 Rl34a 1.219 1.91 0.114 1.1 6.3 0.58 R744 1.237 2.35 0.092 0.9 5.4 0.50 !28 Rl34a 1.141 2.70 0.092 2.1 7.0 0.45 R744 1.184 3.85 0.085 2.2 5.8 0.30 130 Rl34a 1.132 3.12 0.082 1.9 6.3 0.41 R744 1.182 4.49 0.079 1.9 5.7 0.29 119 Rl34a 3.142 2.58 0.346 3.4 11.1 0.66 0.67 4.5 1.0 R744 3.405 2.76 0.034 5.3 9.2 0.83 0.84 2.4 1.0 M10 R134a 3.754 1.68 0.473 -1.0 9.4 0.56 0.54 6.8 1.0 R744 4.587 1.86 0.570 0.8 5.3 0.80 0.84 2.7 1.0 117 R134a 2.208 1.75 0.149 1.7 8.7 0.60 0.58 5.6 1.0 R744 2.290 1.76 0.090 4.1 7.5 0.79 0.79 2.5 1.0 M8 R134a 2.817 1.27 0.261 -2.5 6.5 0.54 0.50 7.8 1.0 R744 3.466 1.31 0.310 -1.9 2.9 0.76 0.77 3.1 1.0 Ill R134a 2.773 2.03 0.273 5.4 12.3 0.62 0.60 5.2 1.0 R744 2.780 2.01 0.214 7.5 11.2 0.78 0.79 2.5 1.0 M5 Rl34a 3.480 1.44 0.401 0.8 9.4 0.56 0.52 7.5 1.0 R744 4.722 1.47 0.502 2.1 6.5 0.75 0.77 3.1 1.0 Figure 4 shows trends in COP -capacity relation for transcritical R744 and Rl34a systems. Generally higher compressor speeds produce higher capacities at expense of low COP. Due to greater capacity system cycles earlier at higher compressor speeds. Reduced speed (idle) increase COP but capacity is compromised. The highest COPs are achieved at cycling conditions. Cycling frequency is high (approx. 2 per min) so heat exchangers operate over the entire period (both in off and on time). Combined COP-capacity performance ofR744 system is better over the most of the operating range except at the center of the diagram - at almost same capacity R134a has slightly higher COPs. This occurs at extremely high outside temperature (55°C) and at idling (950rpm). Some details are given in Table 4. ht order to prevent evaporator frosting, systems of the size tested usually limit the evaporation temperature by a low-pressure switch that actuates the clutch. This typical control scheme, which was used in our tests, caused cycling at low load. The low-pressure switch is not adjustable. Due to very short cycling time (on time approx. 3-12s) air inlet temperatures to evaporator vary only 0.2 oc while air temperatures at the exit vary a bit more, 1-3°C. Evaporation pressure varies 5-7kPa (for 149 Rl34a). Initial control scheme for R744 tests to mimic Rl34a saturation temperatures when system cycles on/off was changed due to practical problems. Results shown in Table 4 are obtained by turning the system on/off at the same air temperatures that were measured in test runs with Rl34a system cycling. Small change in air temperatures in cycling indicate why heat balance on air side and calorimetry are in very good agreement. Nevertheless, heat balance on refrigerant side can not be used to determine capacity due to two-phase flow at the evaporator exit. 5 4 z. "(j 3 ro a. ro u 2 1 r-0 I I ~1 0 • R134a 1800rpm D R744 equal rpm steady state ~/ data MlO 950rpm 1:1~ : steady state data and one 1800rpm cycling data r ' . . 16 119 M8 · ~ 01111 _j_ M1~113 i 117 .• •' ~M26128 130 -~3·•. Iii 2 3 COP •' .. .. Cycling data ~ ... o _I 4 5 Figure 4. Difference in Rl34a and R744 performance CONCLUSION 6 This paper presents ongoing project in residential and mobile ale and heat pwnp systems that use HFC and R744 as a refrigerant along with some preliminary comparisons for mobile a/c. One objective of the project is to help understand the performance of the systems that use these refrigerants at similar operating conditions and provide valid data for TEWI analysis. The initial data-to-data comparisons presented here showed that the R-744 system had generally higher capacity than the R-134a baseline system. The R-744 system was sized to provide approximately equal capacity at the extreme high-temperature (54.4 °C) idling condition, but its COP felllO% short ofthe baseline system at that point. At outdoor ambient temperatures below 40°C where most a/c operation occurs, the R-744 system COP exceeded that of the baseline R-134a system by 40% and more. 8 REFERENCES Sand, J.R, S.K. Fischer, V.D. Baxter, 1997. Energy and global warming impact ofHFC refrigerants and emerging technologies. A report of study sponsored by the Alternative Fluorocarbons Environmental Acceptability Study (AFEAS) and the U.S. Department of Energy Bhatti, M.S, 1997. A critical look at R-744 and R-134a mobile air conditioning systems, SAE paper 970527, SAE Congress Proceeding, pp. 117-141 Pettersen, J., R Aarlien, P. Neksaa, G. Skaugen and K. Aflekt, 1997a A comparative evaluation of C02 and R22 residential air-conditioning systems in a Japanese climate, 1IRIIEA Workshop in C02 technology in refrigeration, heat pumps and air conditioning systems, Tronheim, Norway Pettersen, J. and Hafner, A 1997b. Energetischer Wirkungsgrad und 1EWI von CO:z-Fahrzeug-Klimaanlagen. Seminar: Fahrzeugklimatisierung mit Naturlichen Kaetlemitteln, Karlsruhe March 8, C.F. Mueller Verlag J. Yin, Y. C. PaJk, D. Boewe, R McEnaney, A Beaver, C. W. Bullard, P. S. Hmjak 1998. Experimental and model comparison oftranscritical CO:z versus R134a and R410 system performance, Natural Working Fluids '98, IIR-Gustav Lorentzen conference, Oslo 150 

Experimental Comparison of Mobile A/C Systems When Operated With Transcritical CO2 Versus Conventional R134A

https://docs.lib.purdue.edu/cgi/viewcontent.cgi?article=1401&amp;context=iracc

Experimental Comparison of Mobile A/C Systems When Operated With Transcritical CO2 Versus Conventional R134A

Abstract

Similar works

Full text

Available Versions

Purdue E-Pubs