Compact Counterflow Gas Cooler for R-744

C.W. Bullard, Ph.D. J.M. Yin, Ph.D. P.S. Hrnjak, Ph.D.Fellow ASHRAE Member ASHRAE

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ABSTRACT

The exit temperature from the gas cooler is an extremelyimportant parameter affecting the efficiency of the transcrit-ical air-conditioning and heat pump cycle. Interest in carbondioxide as a refrigerant has recently led to the development andtesting of several prototypes, all of cross-flow design. This arti-cle employs a simulation model to analyze trade-offs (pressuredrop vs. heat transfer) involved in approaching counter-flowconfiguration through multi-slab configurations. It alsopresents the first experimental confirmation of the model.

INTRODUCTION

As a natural and environmentally benign refrigerant, CO2(R-744) is attracting significant attention. Systems with R-744are more common in laboratories and research and develop-ment centers than in real life, at least at this time. Applicationscan range from automotive air conditioning, residential airconditioning (a/c) and heat pumps, and water heaters to somespecific applications, such as dry cleaning.

In most air-conditioning operating ranges, R-744 systemsoperate in transcritical mode. The major difference betweentranscritical and conventional operation is that heat rejectionis in the supercritical region because the critical temperaturefor R-744 is 31°C. Consequently, pressure and temperature arenot related and this opens additional control and optimizationissues.

Transcritical systems with R-744, in particular, werenotorious for their poor performance (Bhatti 1997). Newlydeveloped systems described in Pettersen et al. (1997), as wellin publications by a group such as Yin et al. (1998), demon-strated very good performance. That affects expectations andshows increased strength of the research activity.

Compressor suction and discharge pressures (3-12 MPa)are far higher than for conventional hydrofluorocarbon (HFC)refrigerants. Thanks to the p-t relationship at these pressures,higher pressure drops may be tolerated in the heat exchangerswithout significantly impairing cycle performance. Moreover,higher compressor efficiencies can be achieved because of thelower pressure ratio. Heat transfer is better also because offavorable thermophysical properties.

The logical question is how to improve performancefurther or is there a limit to the COP. In this paper, we will tryto address the first question, focusing on gas cooler issues.

This paper will discuss the importance of the refrigeranttemperature at the gas cooler exit and then investigate quan-titatively two distinct gas cooler design concepts: the multi-pass single-slab designs used in existing prototypes and thesingle-pass multi-slab counterflow configurations proposedhere. This idea is of utmost importance in the counterflowarrangement for a transcritical R-744 system conceived in1998 during our earlier experiments.

IMPORTANCE OF THE GAS COOLER EXIT TEMPERATURE

Figure 1 shows the predicted relationship between COPand discharge pressure for different gas cooler exit tempera-tures and shows the obtainable benefits from a closer approachto ambient air temperature. The evaporation temperature inthis example is set at 3.9°C (corresponding to evaporationpressure 3.85 MPa), and the suction line internal heatexchanger effectiveness is 0.8. Pressure drop is neglected tosimplify the presentation. The compressor efficiency used inthe calculation was fitted from a previous experiment (Boeweet al. 1999) at 2000 rpm.

THIS PREPRINT IS FOR DISCUSSION PURPOSES ONLY, FOR INCLUSION IN ASHRAE TRANSACTIONS 2002, V. 108, Pt. 1. Not to be reprinted in whole or inpart without written permission of the American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc., 1791 Tullie Circle, NE, Atlanta, GA 30329.Opinions, findings, conclusions, or recommendations expressed in this paper are those of the author(s) and do not necessarily reflect the views of ASHRAE. Writtenquestions and comments regarding this paper should be received at ASHRAE no later than January 25, 2002.

C.W. Bullard and P.S. Hrnjak are professors at the University of Illinois, Urabana, Ill. J.M. Yin is a technical advisor at Modine Manufac-turing Company, Racine, Wisc.

Despite the assumptions underlying Figure 1, severalconclusions can be drawn from it. First, when the air inlettemperature entering the gas cooler increases, the R-744 exittemperature also increases so the operating pressure needs tobe increased in order to maximize cycle COP in transcriticaloperation. This COP decreases as the gas cooler exit temper-ature increases. Second, the COP curve tends to be flatterwhen the temperature at the gas cooler exit is higher. Third, theCOP curve is not symmetric along the pressure axis. At lowerexit temperatures, the COP drops more steeply on the low-pressure side than on the high-pressure side. Also, accuratehigh-side pressure control is more important for maximizingCOP on cool days than for the case of higher inlet air temper-ature. And fourth, when the refrigerant temperature at the gascooler exit is lower than the critical temperature (31°C), thereis no local optimum COP.

Experiments conducted earlier with a prototype R-744system (Boewe et al. 1999) showed that for a single slab, threepass gas cooler, the approach temperature difference betweenrefrigerant exit and ambient air ranged from 2°C to 9°C. If thatdifference could be reduced by redesigning the gas cooler(without violating packaging constraints), system perfor-mance could be significantly improved, as shown in Figure 1.For example, suppose the air inlet temperature is 34°C. If therefrigerant exit temperature from the gas cooler is 38°C, fromFigure 1 the highest COP is about 2.54, and the correspondingdischarge pressure is 9.28 MPa. If the discharge temperaturecan be reduced to 36°C, the corresponding COP and pressurewill be 2.79 MPa and 8.83 MPa, respectively. Namely, a 2°Creduction in refrigerant exit temperature could increase COPby as much as 10%.

Heat transfer close to the critical point is a strong functionof specific heat (Cp). Figure 2 shows how the specific heatchanges with temperature at different pressures. The specific

heat of R-744 in the gas cooler will gradually increase down-stream, then go through a peak region when close to the criticalpoint and then decrease again as it approaches the ambient airtemperature. As the Cp changes along the gas cooler, so alsodoes the temperature difference between refrigerant and airalong the heat exchanger tube. Due to high refrigerant-airtemperature difference in the inlet region and lower Cp, therefrigerant temperature falls quickly. Then as the fluid entersthe peak Cp region, even at the same heat transfer rate, therefrigerant temperature tends to be flatter for two reasons:higher Cp and lower refrigerant-air temperature difference.

GAS COOLER DESIGNS TO BE COMPARED

This paper will focus on two designs: multi-pass(currently used) vs. newly proposed multi-slab (counterflow)arrangement.

To replace the condenser in a typical compact car, the gascooler should fit in a rectangular box of length = 523 mm,height = 355 mm, and depth = 16.51 mm. The baseline heatexchanger (multi-pass) and its dimensions are shown in Table1 and Figure 3.

MULTIPASS ARRANGEMENT

Figure 4 shows a schematic (tube arrangement andnumbers in each pass) of the multi-pass heat exchangerpresented in Figure 3.

Experiments have been done with such a gas cooler andcompared to the model evaluation. Figure 5 shows the predic-tion results for the operating conditions listed in Table 2.

As shown in Figure 5, most of the heat transfer occurs inthe first pass, and the last pass has the smallest heat transferrate because of the small temperature difference. The refrig-erant exit temperature from the model was about 49.8°C at thiscondition, which means the approaching temperature differ-

Figure 1 Effect of refrigerant exit temperature on COP forrealistic high-side pressures.

Figure 2 Specific heat of R-744 in the transcritical region.

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ence was about 6.9°C, and the corresponding refrigerantenthalpy at the gas cooler exit was about 357.5 kJ/kg. Thisresult is for a medium air flow rate of 0.499 kg/s. When the airflow rate is reduced, the approaching temperature differenceincreases.

MULTI-SLAB COUNTERFLOW ARRANGEMENT

The objective is to stay within the same volume and useavailable fins and tubes as before but demonstrate that coun-terflow arrangement will provide lower approaching temper-ature—cooler refrigerant temperature at the exit. Thisarrangement is more suited for a fluid that rejects heat in singlephase. Counterflow is more important for the gas cooler thanfor the evaporator or condenser because the refrigerant sidehas a large temperature gradient. One effective way to realizea counterflow arrangement is to use a multi-slab arrangement.Figure 6 shows our approach in this case. The advantage ofthis arrangement is that the third slab will always be exposedto the cold entering air. There are numerous practical ways torealize such a concept.

TABLE 1 MAC1 Gas Cooler Dimensions

Finned Length (m) 0.523 Height (m) 0.355 Depth (m) 0.0165

Free flow area (m2) 0.167 Number of tubes 34 Louver angle (o) 23

Volume (m3) 0.0035 Fin thickness (mm) 0.102 Louver pitch (mm) 0.99

Airside area (m2) 5.312 Fin pitch (mm) 1.155 Fin height (mm) 8.45

Ref side area (m2) 0.49 Number of fins per meter 866 Louver length (mm) 6.8

TABLE 2 Measured Operating Conditions and Exit Temperature

Air inlet temperature (°C) 42.8

Air flow rate (kg/s) 0.499

Refrigerant inlet temperature (°C) 124.3

Refrigerant inlet pressure (MPa) 10.95

Refrigerant mass flow rate (g/s) 38

Refrigerant exit temperature (°C) 49.04

Figure 3 Photo of the multi-pass conventional gas cooler.

Figure 4 Multi-pass gas cooler.

Figure 5 Temperature distribution along the length ofeach pass.

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To stay within the required depth, we have opted to use amicrochannel tube with a smaller major axis. One possiblearrangement is shown in Figure 7. With this size we could havethree microchannel tubes in the depth of one shown in Table 1.

Since R-744 is less sensitive to pressure drop than otherconventional refrigerants, we have the freedom to choosesmaller port diameters. As shown in Figure 7, the tube thick-ness is only about 1.095 mm. It is even smaller than the finpitch with fin density of 23 fpi. If the fin density is less than23 fpi, an increase in the microchannel tube number (or reduc-tion of fin height) will counterintuitively increase the air-sideheat transfer area and reduce the refrigerant-side pressuredrop. Practically, the number of microchannel tubes or slabs isalso affected by the cost of the tube and bending.

Figure 8 shows the temperature distribution along therefrigerant flow path in each slab. Again, the slab that is closestto the refrigerant inlet has the highest heat transfer rate, but theslab closest to the air inlet is more effective than the third passin the multi-pass case because the whole slab is exposed to thecoldest inlet air temperature.

The refrigerant exit temperature is about 46.4°C at thesame condition as shown in Table 2, based on model predic-tion. The approaching temperature difference is about 3.6°C,and the corresponding refrigerant enthalpy at gas cooler exit isabout 335.4 kJ/kg. Compared to the three-pass arrangement,the approaching temperature is reduced by about 49% (from6.9°C to 3.6°C), and the refrigerant exit enthalpy is reduced byabout 6% (from 357.5 kJ/kg to 335.4 kJ/kg). Refrigerant-side

TABLE 3 MACX Gas Cooler Dimensions

Finned Length (m) 0.523 Height (m) 0.355 Depth (m) 0.0165

Free flow area (m2) 0.150 Number of tubes 54 Louver angle (o) 23

Volume (m3) 0.0035 Fin thickness (mm) 0.102 Louver pitch (mm) 0.99

Airside area (m2) 5.312 Fin pitch (mm) 1.159 Fin height (mm) 8.45

Ref side area (m2) 0.69 Number of fins per meter 863 Louver length (mm) 6.8

Figure 6 Multi-slab arrangement—side view.

Figure 7 Microchannel tube dimensions used for(a)MAC1 and (b) MACX.

Figure 8 Predicted temperature distribution along thelength of each slab for the multi-slabarrangement.

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pressure drops at conditions presented are 32 kPa for MAC1and 65 kPa for MACX.

CONCLUSIONS

A multi-slab, counterflow arrangement, gas cooler designwas proposed as conceived in 1998 during earlier experi-ments. To our best knowledge this is the first proposal of thiskind in the open literature.

This new design offers better performance than thecommonly used multi-pass design. For the given gas coolervolume, the new design can have 11% higher gas coolercapacity than the multi-pass. Predicted approach temperaturedifference is reduced from 6.9°C to 3.6°C. Points A and B inFigure 1 represent the exit conditions from two heat exchangerdesigns. The new design is more effective at lower air-flowrate and high capacity conditions.

We have built the heat exchanger based on this concept.It is in our laboratories under testing.

ACKNOWLEDGMENT

We are grateful to Hydro Aluminum A.S. for funding theresearch that generated this idea.

REFERENCES

Bhatti, M.S. 1997. A critical look at R-744 and R-134amobile air conditioning systems. SAE paper 970527,SAE Congress Proceedings, pp. 117-141.

Boewe, D., C.W. Bullard, and P.S. Hrnjak. 1999. An experi-mental investigation of transcritical carbon dioxide sys-tems for residential air conditioning. University ofIllinois at Urbana-Champaign, ACRC CR-18.

McEnaney, R., Y.C. Park, J.M. Yin, C.W. Bullard, and P.S.Hrnjak. 1999. Performance of the prototype of a tran-

scritical R-744 mobile air-conditioning system. SAEpaper 1999-01-0872.

McEnaney, R.P., D.E. Boewe, J.M. Yin, Y.C. Park, C.W.Bullard, and P.S. Hrnjak. 1998. Experimental compari-son of mobile air-conditioning systems when operatedwith transcritical CO2 versus conventional R-134a.International Refrigeration Conference at Purdue, pp.145-150.

Park, Y.C., J.M. Yin, C.W. Bullard, and P.S. Hrnjak. 1999.Experimental and model analysis of control and operat-ing parameters of transcritical CO2 mobile air-condi-tioning system. VTMS, London, May.

Pettersen, J., R. Aarlien, P. Neksaa, G. Skaugen, and K.Aflekt. 1997. A comparative evaluation of CO2 and R-22 residential air-conditioning systems in a Japanese cli-mate. IIR/IEA Workshop in CO2 technology in refriger-ation, heat pumps and air conditioning systems,Trondheim, Norway.

Yin, J.M., Y.C. Park, R.P. McEnaney,D.E. Boewe, A. Bea-ver, C.W. Bullard, and P.S. Hrnjak. 1998. Experimentaland model comparison of transcritical CO2 versus R-134a and R-410A system performance. IIR conferenceGustav Lorentzen, Oslo, Proceedings, Preprints, pp.331-340.

Pitla, S.S., S. Ramadhyani, and E.A. Groll. 2000. Convectiveheat transfer from in-tube flow of turbulent supercriticalcarbon dioxide, Part 1: Numerical analysis. Acceptedfor publication in October 2000 in International Journalof HVAC&R Research.

Pitla, S.S., E.A. Groll, and S. Ramadhyani. 2000. Convectiveheat transfer from in-tube cooling of turbulent supercrit-ical carbon dioxide, Part 2: Experimental data andnumerical predictions. Accepted for publication inOctober 2000 in International Journal of HVAC&RResearch.

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