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Heat exchanger design for C[O.sub.2] cycle with a linear compressor.

Hermetic linear compressors already have been developed for subcritical cycles, especially for small-capacity refrigeration systems, which have a refrigeration capacity less than 500 W and have demonstrated better performance potential than conventional compressors. In this paper, the performance potential of a transcritical carbon dioxide (C[O.sub.2]) cycle with a linear compressor is explored for medium-temperature refrigeration systems, which have a return air temperature range from -2[degrees]C to 6[degrees]C, through C[O.sub.2] cycle simulation using compressor efficiencies measured. In addition, the design of the heat exchanger for such systems was practiced. The simulated results of the transcritical C[O.sub.2] cycle performance showed that the cooling capacity was 235 W at 40.5[degrees]C ambient temperature and the coefficient of performance was 1.31 at 32.2[degrees]C ambient temperature. The heat exchangers were then designed using a verified heat exchanger design software. As a target application, a 200 liter effective volume refrigerator cabinet was selected, and its refrigeration load was evaluated by using load calculation software. When the C[O.sub.2] linear compressor with suitably designed heat exchangers was applied to the selected refrigeration cabinet operating at medium temperature refrigeration conditions, the simulated results showed that the designed C[O.sub.2] system based on a linear compressor could provide sufficient refrigeration capacity.

INTRODUCTION

Extensive research and development have been performed on carbon dioxide (C[O.sub.2]) technology over the last two decades since the revival of C[O.sub.2] research by Lorentzen and his colleagues in the early 1990s (Lorentzen and Pettersen 1993; Neksa 1994; Pettersen and Skaugen 1994; Pettersen et al. 1995). Among the many investigated applications, there are several successful C[O.sub.2] technology applications: C[O.sub.2] residential heat pump water heaters, C[O.sub.2] environmental control units, C[O.sub.2] automotive climate control systems, and C[O.sub.2] commercial refrigeration systems. C[O.sub.2] heat pump water heaters have been successfully commercialized and introduced exclusively in Japan since 2001 (Endoh et al. 2006). Meanwhile, Cutler et al. (2000) developed a C[O.sub.2] environmental control unit using a 10 kW cooling capacity C[O.sub.2] reciprocating compressor; they reported that a cooling capacity of 10-13 kW and a corresponding coefficient of performance (COP) of 2.3-3.5 could be reached for outdoor ambient temperatures of 27[degrees]C to 41[degrees]C. Additionally, major automobile companies and their suppliers have developed vehicles with C[O.sub.2] automotive climate control systems to provide main cooling and heating or auxiliary heating. Manzione et al. (2006) reported that these C[O.sub.2] automotive climate control systems demonstrated better cooling performance than the current climate control system. Last, C[O.sub.2] commercial refrigeration systems have developed more efficiently. Rohrer (2006) compared the performance of the bottle cooler while using a 600 W input variable-speed two-stage rotary type C[O.sub.2] compressor; he reported that the energy consumption of the C[O.sub.2] system was 30% lower than that of the R-134a baseline system at 32.2[degrees]C ambient temperature. Suss (2005) reported a performance potential of the reciprocating type C[O.sub.2] compressor having 1 kW refrigeration capacity; in his research, 18% energy savings were found as compared to the R-134a system under the European climate.

As summarized, most previous studies have focused on the application of C[O.sub.2] technology to large refrigeration capacities greater than 500 W. Three major international food and beverage companies initiated the Refrigerants, Naturally partnership (RN 2004) to replace hydrofluorocarbon (HFC) refrigerants with natural refrigerants and to reduce the contribution to global warming from the food and drink industry and its supply chain. The partnership's focus is to develop equipment using alternative refrigerants while maintaining high efficiency. Moreover, the number of small-capacity systems that have refrigeration capacities less than 500 W is growing faster than systems of other capacity ranges, especially in developing countries. Therefore, C[O.sub.2] linear compressors with a small capacity range and high efficiency may have great potential in producing energy-efficient refrigeration equipment. The objective of the current study is to provide design examples of the heat exchangers and to investigate the system performance of a medium-temperature C[O.sub.2] refrigeration system of small capacity and with a linear compressor, which has a return air temperature range from -2[degrees]C to 6[degrees]C.

CYCLE MODELS

To model the performance of the transcritical C[O.sub.2] cycle, a previously developed computer model of the C[O.sub.2] refrigeration cycle (Hwang and Radermacher 1998) was modified in two steps. First, the compressor efficiencies were measured and correlated. The compressor and volumetric efficiency were measured by using the compressor calorimeter. The compressor efficiency is defined as the ratio of the isentropic compressor work and the actual motor power input. The volumetric efficiency is defined as the ratio of the actual mass flow rate to the ideal mass flow rate (based on displacement and strokes per minute) delivered by the compressor. During the measurements, the gas cooling pressure was varied from 8.5 to 11 MPa while the evaporating pressure and the degree of superheating at the evaporator outlet were kept constant at 2.8 MPa and 12[degrees]C, respectively. The test results were correlated in first-order linear equations as the function of the pressure radio (PR) as shown in Equations 1 and 2. The goodness of the regression, [R.sup.2], was 0.97 for both efficiency models. For the given PR 3.0, the measured compressor and volumetric efficiency were 0.66 and 0.77, respectively.

[[eta].sub.comp] = 0.801 - 0.0477 x PR (1)

[[eta].sub.vol] = 1.140 - 0.1224 x PR (2)

Second, the C[O.sub.2] refrigeration cycle model was modified using those compressor efficiency correlations and the following assumptions for the small capacity refrigeration system:

* Cabinet air temperature is maintained at 3.3[degrees]C.

* Ambient temperature varies at three levels: 24.5[degrees]C, 32.2[degrees]C, and 40.5[degrees]C.

* Evaporating temperature is maintained at -8[degrees]C.

* The degree of superheating at the evaporator outlet is 12[degrees]C to guarantee no frost formation along the suction line (ARI 1997).

* The approach temperature at the gas cooler outlet is 3[degrees]C.

* The compressor displacemental volume is 0.75 cc.

* The compressor speed is 3600 strokes per minute.

* The evaporator and condenser fan powers are 15 W and 30 W, respectively.

[FIGURE 1 OMITTED]

In the evaluation of the performance potential of the cycle, the gas cooling pressure was optimized for the range of operating conditions. The calculated performance of the C[O.sub.2] cycle with the linear compressor is illustrated in Figure 1 for various gas cooling pressures. In Table 1, the performance of the cycles at their optimum gas cooling pressure is summarized for each ambient temperature. Since insufficient cooling capacity becomes an issue at the highest temperature and energy consumption is of concern for the most representative ambient temperature, the capacity at 40.5[degrees]C and the COP at 32.2[degrees]C are of concern. Observations from the results are as follows:

* The linear compressor can provide 235 W cooling capacity at 40.5[degrees]C ambient condition.

* The linear compressor can provide 1.31 COP at 32.2[degrees]C ambient condition.

HEAT EXCHNAGER DESIGN

Once the cycle performance was simulated, the heat exchangers were designed to satisfy the thermal requirements as summarized in Table 1, using a previously developed heat exchanger design software (CEEE 2006). Two types of heat exchangers, fin-and-tube (FT) and minichannel (MC), were investigated. The evaporator and gas cooler were designed to meet burst pressures of 22 and 32 MPa, respectively. In the calculations, it was assumed that the airflow rate was always 0.0531 [m.sup.3]/s for the evaporator and the air velocity was always 1 m/s for the gas cooler. To predict the refrigerant-side heat transfer coefficients, Gnielinski's (1976) single-phase heat transfer correlation and Gungor and Winterton's (1986) two-phase heat transfer correlations were used. To predict the refrigerant-side pressure drops, Churchill's (1977) single-phase pressure drop correlation and Friedel's (1979) two-phase pressure drop correlation were used. Since the air-side heat transfer is determined by the type of fin, two types of fins (corrugated and louvered) were considered. For the air-side heat transfer and pressure drop correlations, the Kim-Youn-Webb correlation (Kim et al. 1997) was used for the corrugated fins and Chang and Wang's (1997) correlation was used for the louvered fin. Since the simulation results show that the louvered fin has approximately 40% better air-side heat transfer coefficients than the corrugated fin, the louvered fin was selected for the final design. For the base tube geometry, 9.5 mm outside diameter tube, which is the most common tube size for FT heat exchangers, was selected. For the base MC tube geometry, a flat tube having circular ports with 0.79 mm hydraulic diameter and 1.5 mm port pitch was selected as a design example. The height of the flat tube was 1.65 mm and its width was determined according to the number of ports as shown in Figure 2.

Gas Cooler Design

To design the fin-and-tube type gas cooler (FTGC), three different configurations were examined with typical tube size (9.5 mm outer diameter) as shown in Table 2 and Figure 3. The total height of the evaporator and gas cooler was limited to the height equivalent to 12 rows of tubes by assuming both heat exchangers aligned top to bottom. One six-row height design was selected to investigate the option for the same height for both heat exchangers. Two eight-row height designs were selected to investigate the option for the taller gas cooler and the smaller height evaporator.

Modeling results are summarized in Table 3. The heat transferred by each column is compared in Table 4. From these results the following were observed:

* All three FTGCs meet the capacity requirements for all ambient temperatures.

* The 4 x 8 GC performs better than other two GCs at low ambient temperature (24.5[degrees]C).

* GCs with eight rows of tubes perform better than the GC with six rows due to the higher airflow rate.

* In the 4 x 8 GC, the first column carries only 2% of the total load (Table 4). Therefore, the number of columns should be limited to four.

[FIGURE 2 OMITTED]

[FIGURE 3 OMITTED]

Since the performance of the 3 x 8 GC meets the capacity requirement and its volume is smaller than that of the 4 x 8 GC, the 3 x 8 GC is selected for the gas cooler design.

To further improve the FTGC design, a smaller tube design was investigated. The inner diameter of the smaller tube was set to 3 mm, and the frontal dimension and the fin spacing were kept the same. The vertical spacing of the tube was designed to produce the maximum heat transfer capacity, while the refrigerant-side pressure drop was designed not to exceed 20 kPa. Table 5 shows specifications of the optimized FTGC using tubes with smaller tube diameters and those with the typical tube diameter (9.5 mm). Hereafter, the FTGC with typical tube diameter is referred as the base FTGC.

Table 6 shows the simulated performance of the optimized smaller-tube FTGC and that of the base FTGC. The optimized smaller-tube FTGC delivers performance similar to that of the base FTGC, while its external volume is only 54% of the base FTGC. The reduced tube cross-sectional area contributes to the increased refrigerant-side heat transfer coefficient and pressure drop by 5.3 to 7.8 times and by two orders of magnitude, respectively. This also results in 66% to 131% higher overall heat transfer coefficient. Since the pressure drop of the smaller-tube FTGC is still within an acceptable value of around 100 kPa, it can be concluded that the use of a smaller-diameter tube improves the performance of low refrigerant mass flux applications.

For the design of the minichannel type GC (MCGC), a single column MCGC was examined. Two designs using different numbers of tubes, 20 and 25, were investigated, keeping the total height and frontal area the same as that of the base FTGC. Two fin densities, 1.5 and 3.6 mm, were investigated for both designs. The widths of fin and minichannel tube were determined by adjusting the number of ports to match the delivered capacity to the required capacity. The number of parallel tubes along the path was optimized by changing the location of the internal blocks as illustrated in Figure 4. In Figure 4 and Table 7, the pass configuration sequence number illustrates parallel channels used in sequence as the refrigerant changes its horizontal direction along the heat exchanger pass. For example, a pass of 3-4-5-4-3 means that the number of parallel channels changes as follows: 3 channels for the first pass, 4 channels for the second pass, 5 channels for the third pass, 4 channels for the fourth pass, and 3 channels for the fifth pass.

Modeling results are summarized in Table 8. From these results the following were observed:

* All four MCGCs meet the capacity requirements for all ambient temperatures.

* Using high fin density (fin pitch 1.54 mm) for the MCGCs reduces the volume of the MCGC by 44% rather than using low fin density (fin pitch 3.63 mm).

When the same low fin density is used for both FTGCs and MCGCs, the volume of the MCGCs is 13% to 27% smaller than that of the FTGC with 7.92 mm inner diameter tube. However, the volume of the MCGCs is 36% to 61% greater than that of the FTGC with 3.0 mm inner diameter tube. The MCGC with 25 tubes and higher fin density has 11% smaller volume than that of the smaller-tube FTGC.

[FIGURE 4 OMITTED]

Evaporator Design

Two different configurations of FT heat exchangers were investigated for the evaporator with the most common tube size (9.52 mm outer diameter) as shown in Table 9 and Figure 5. A six-row height design was selected to investigate the option for the same height (6 rows) heat exchangers for the evaporator and gas cooler. Another four-row height design was selected to investigate the option to keep the total number of rows for the evaporator and gas cooler at 12 rows.

[FIGURE 5 OMITTED]

Modeling results are summarized in Table 10. The heat transfer rate and air-side outlet temperature distribution are summarized in Table 11 and Figure 6, respectively. From these results the following were observed:

* Both designs can meet the capacity requirements for all ambient temperatures.

* Both designs perform similarly.

* In the 3 x 4 EV, the first column carries only 4% of total load (Table 11). Therefore, the number of columns should be limited to three.

* Since three-column EV has less contribution from the superheated vapor area on the air cooling than the two-column EV, the 3 x 4 EV has wider area of low temperature (Figure 6).

SYSTEM PERFORMANCE

To evaluate the applicability of the cycle components designed for medium-temperature refrigeration systems, the cooling load of the reach-in unit with 200-liter effective internal volume was evaluated using refrigeration system transient simulation software (CEEE 2005). For the cabinet load evaluation, the cabin average temperature was assumed to be 3.3[degrees]C. Results from the simulation are shown in Figure 7 and Table 12. The calculated cabinet loads were 68.0, 92.7, and 119.3 W at ambient temperatures 24.5[degrees]C, 32.2[degrees]C, and 40.5[degrees]C, respectively. When the calculated cabinet loads were compared to the cooling capacity delivered by the C[O.sub.2] refrigeration system, the cooling capacity of the designed system provided 75% or more cooling capacity than required for all ambient temperatures investigated. Therefore, it can be concluded that the C[O.sub.2] refrigeration cycle designed with a linear compressor, a 3 x 8 FTGC, and a 3 x 4 FTEV with typical tube size (9.52 mm) can provide adequate cooling capacity for a 200 liter cabinet.

CONCLUSIONS

The performance potential of a C[O.sub.2] transcritical cycle with a hermetic linear compressor was simulated for medium-temperature refrigeration systems by using measured compressor efficiencies. The simulated results show that the cooling capacity was 235 W at 40.5[degrees]C ambient temperature and the COP was 1.31 at 32.2[degrees]C ambient temperature. Two types of heat exchangers, fin-and-tube and minichannel, were designed and optimized using heat exchanger design software. The comparison of the design parameters found that the use of smaller-diameter tubes improves the performance of the fin-and-tube heat exchanger, especially for low refrigerant mass flux applications. When the selected minichannel gas cooler designs were compared to the selected fin-and-tube gas cooler designs, the volume of the minichannel gas coolers was 17%-23% smaller than that of the fin-and-tube gas cooler. However, the minichannel gas cooler requires a higher fin density to compete with the fin-and-tube gas cooler having smaller diameter tubes. Therefore, when air-side contamination is not a serious issue for the system operating environment, the minichannel is an excellent option to save heat exchanger volume. The performance of the C[O.sub.2] linear compressor together with heat exchangers suitably designed can be successfully applied to medium-temperature refrigeration systems having 200 liter effective volume.

[FIGURE 6 OMITTED]

ACKNOWLEDGMENTS

The support of this research through the Alternative Cooling Technologies and Applications Consortium of the Center for Environmental Energy Engineering is gratefully acknowledged. Furthermore, the authors acknowledge LG Electronics for providing the performance data of the C[O.sub.2] linear compressor.

[FIGURE 7 OMITTED]

NOMENCLATURE

AFR = airflow rate

COP = coefficient of performance

EV = evaporator

FT = fin-and-tube

GC = gas cooler

HFC = hydrofluorocarbon

HTC = heat transfer coefficient

HX = heat exchanger

MC = mini-channel

P = pressure

PD = pressure drop

PR = pressure ratio

T = temperature

Greek Letters

[eta] = efficiency

Subscripts

amb = ambient

comp = compressor

dis = discharge

ev = evaporator

gc = gas cooler

opt = optimum

ref = refrigerant

vol = volumetric

REFERENCES

ARI. 1997. ANSI/ARI Standard 520?aPositive Displacement Condensing Units. Arlington, VA: Air-Conditioning and Refrigeration Institute.

CEEE. 2005. TransRef: Transient Simulation Program of Vapor Compression Systems, Version 3.1. College Park, MD: University of Maryland.

CEEE. 2006. CoilDesigner: Heat Exchanger Design Software, Version 2.5. College Park, MD: University of Maryland.

Chang, Y.J., and C.C. Wang. 1997. A generalized heat transfer correlation for louver fin geometry. Int. J. Heat Mass Transfer. 40(3):553-44.

Churchill, S.W. 1977. Friction factor equation spans all fluid flow regimes. Chem. Eng. 84(24):91-92.

Cutler, B., Y. Hwang, and R. Radermacher. 2000. Development of carbon dioxide environmental control unit. Proceedings of the 4th IIR-Gustav Lorentzen Conference, Purdue, IN, pp. 91-98.

Endoh, K., T. Kouno, M. Gommori, T. Tanami, K. Mizutani, and M. Miyata. 2006. Instant hot-water supply heat-pump water heater using CO2 refrigerant for home use. Proceedings of the 7th IIR Gustav Lorentzen Conference on Natural Working Fluids, Trondheim, Norway, pp. 27-30.

Friedel, L. 1979. Improved friction pressure drop correlations for horizontal and vertical two phase pipe flow. Proceedings of the European Two Phase Flow Group Meeting, Ispra, Italy, June, paper E2.

Gnielinski, V. 1976. New equations for heat and mass transfer in turbulent pipe and channel flow. International Chemical Engineering 16(2):359-68.

Gungor, K.E., and R.H.S. Winterton. 1986. A general correlation for flow boiling in tubes and annuli. Int. J. Heat Mass Transfer 29(3):351-58.

Hwang, Y., and R. Radermacher. 1998. Theoretical evaluation of carbon dioxide refrigeration cycle. HVAC & R Research 4(3):245-63.

Kim, N.H., J.H. Yun, and R.L. Webb. 1997. Heat transfer and friction correlations for wavy plate fin-and-tube heat exchangers. Transactions of the ASME 119:560-67.

Lorentzen, G., and J. Pettersen. 1993. A new, efficient and environmentally benign system for car air-conditioning. International Journal of Refrigeration 16(1):4-12.

Manzione, J., S. Collier, S. Memory, and P. Harnak. 2006. An improved CO2 cooling system for the up-armored HMMWV. Proceedings of the 7th IIR Gustav Lorentzen Conference on Natural Working Fluids, Trondheim, Norway, pp. 53-56.

Neksa, P. 1994. Trans-critical vapour compression heat pumps. Proceedings of I.I.R. Conference for New Applications of Natural Working Fluids in Refrigeration and Air Conditioning, Hanover, Germany, pp. 395-404.

Petterson, J., and G. Skaugen. 1994. Operation of trans-critical C[O.sub.2] vapor compression circuits in vehicle air conditioning. Proceedings of I.I.R. Conference for New Applications of Natural Working Fluids in Refrigeration and Air Conditioning, Hanover, Germany, pp. 495-504.

Pettersen, J., P. Neksa, O.M. Nesje, P.A. Schiefloe, and H. Rekstad. 1995. Recent advances in CO2 refrigeration. Proceedings of the 19th International Congress of Refrigeration, The Hague, The Netherlands, pp. 961-68.

RN. 2004. Refrigerants, Naturally. www.refrigerantsnaturally.com.

Rohrer, C. 2006. Transcritical C[O.sub.2] bottle cooler development. Proceedings of the 7th IIR Gustav Lorentzen Conference on Natural Working Fluids, Trondheim, Norway, pp.179-81.

Suss, J. 2005. Low capacity CO2 systems. ASHRAE Winter Meeting, Orlando, FL, February 5-9.

Yunho Hwang, PhD

Member ASHRAE

Varun Singh

Reinhard Radermacher, PhD

Member ASHRAE

Received October 24, 2006; accepted February 16, 2007

Yunho Hwang is a research associate professor, Varun Singh is a graduate research assistant, and Reinhard Radermacher is a professor in the Department of Mechanical Engineering at the University of Maryland, College Park.
Table 1. Performance of C[O.sub.2] Cycle at Optimum Gas Cooling Pressure

[T.sub.amb], [P.sub.gc,opt], Capacity, Power, Mass Flow
[degrees]C MPa W COP W Rate, g/s

24.5 7.50 394 1.85 213 2.3
32.2 8.75 305 1.31 233 2.2
40.5 11.00 235 0.92 257 1.9

[[eta].sub.comp] [[eta].sub.vol] PR [T.sub.dis], [degrees]C

0.67 0.80 2.82 92
0.64 0.74 3.28 109
0.60 0.64 4.11 136

Table 2. Detailed Specifications of Three Fin-and-Tube Type Gas Coolers

Column x Row 4 x 6 3 x 8 4 x 8
Frontal Area, [m.sup.2] 0.06 0.08 0.08
Heat Transfer Area, [m.sup.2] 2.7 2.7 3.6

Tube Material Copper Copper Copper
 Outside Diameter, mm 9.52 9.52 9.52
 Thickness, mm 0.8 0.8 0.8
 Vertical Pitch, mm 25 25 25
 Horizontal Pitch, mm 21.65 21.65 21.65
 Length, mm 400 400 400

Fin Material Aluminum Aluminum Aluminum
 Type Louvered Louvered Louvered
 Pitch, mm 3.63 3.63 3.63
 Thickness, mm 0.152 0.152 0.152
 Height, mm 150 200 200
 Depth, mm 86.6 65.0 86.6

Table 3. Modeling Results for the Fin-and-Tube Gas Cooler

 [T.sub.amb], Capacity, P[D.sub.gc], [T.sub.gc,out],
Column x Row [degrees]C W kPa [degrees]C

4 x 6 24.5 560 0.1 26.0
Heat transfer 32.2 515 0.1 33.2
 area:
2.7 [m.sup.2] 40.5 452 0.1 40.8
3 x 8 24.5 565 0.1 25.5
Heat transfer 32.2 517 0.1 33.0
 area:
2.7 [m.sup.2] 40.5 453 0.1 40.7
4 x 8 24.5 572 0.1 24.7
Heat transfer 32.2 524 0.1 32.4
 area:
3.6 [m.sup.2] 40.5 454 0.1 40.5

 HT[C.sub.air]/
 AFR, P[D.sub.gc,air], HT[C.sub.ref],
Column x Row [m.sup.3]/s Pa W/[m.sup.2] x K

4 x 6 0.0610 16 81/838
Heat transfer 0.0610 16 81/500
 area:
2.7 [m.sup.2] 0.0610 16 81/345
3 x 8 0.0813 12 89/789
Heat transfer 0.0813 12 89/491
 area:
2.7 [m.sup.2] 0.0813 12 89/343
4 x 8 0.0813 16 81/701
Heat transfer 0.0813 16 81/470
 area:
3.6 [m.sup.2] 0.0813 16 80/340

Table 4. Heat Transfer Rate per Column ([T.sub.amb] = 32.2[degrees]C)

Column Number 1 2 3 4
4 x 6 6% 15% 24% 55%
3 x 8 8% 27% 65% --
4 x 8 2% 8% 26% 64%

Table 5. Detailed Specifications of Two Fin-and-Tube Type Gas Coolers

Column x Row 3 x 8 2 x 20
Frontal Area, [m.sup.2] 0.08 0.08
Heat Transfer Area, [m.sup.2] 2.7 1.6
Heat Exchanger Volume, [m.sup.3] 0.0052 0.0028

Tube Material Copper Copper
 Outside Diameter, mm 9.52 3.5
 Thickness, mm 0.8 0.25
 Vertical Pitch, mm 25 10
 Horizontal Pitch, mm 21.65 17.6
 Length, mm 400 400
Fin Material Aluminum Aluminum
 Type Louvered Louvered
 Pitch, mm 3.63 3.63
 Thickness, mm 0.152 0.152
 Height, mm 200 200
 Depth, mm 65.0 35.2

Table 6. Modeling Results for the Fin-and-Tube Gas Cooler

 [T.sub.amb], Capacity, P[D.sub.gc], [T.sub.gc,out],
Column x Row [degrees]C W kPa [degrees]C

3 x 8 24.5 565 0.1 25.5
Tube ID: 7.92 mm 32.2 517 0.1 33.0
 40.5 462 0.1 40.7
2 x 20 24.5 567 18.3 25.3
Tube ID: 3.00 mm 32.2 522 15.5 32.6
 40.5 463 10.5 40.5

 HT[C.sub.air]/
 AFR, P[D.sub.gc,air], HT[C.sub.ref],
Column x Row [m.sup.3]/s Pa W/[m.sup.2] x K

3 x 8 0.0813 12 89/789
Tube ID: 7.92 mm 0.0813 12 89/491
 0.0813 12 89/332
2 x 20 0.0813 13 81/6,949
Tube ID: 3.00 mm 0.0813 13 81/3,170
 0.0813 13 81/2,081

Table 7. Detailed Specifications of Four Minichannel Gas Coolers

Column x Row 1 x 20 1 x 25
Pass Configuration 4-4-4-4-4 5-4-4-4-4-4
Frontal Area, [m.sup.2] 0.08 0.08
Heat Transfer Area, [m.sup.2] 2.9 2.5
Heat Exchanger Volume, [m.sup.3] 0.0045 0.0038

Minichannel Tube Material Aluminum Aluminum
 Width, mm 56.4 47.1
 Height, mm 1.65 1.65
 Length, mm 400 400
 Vertical Pitch, mm 10.6 8.3
 Port I.D., mm/No. [ea] 0.79/38 0.79/31
Fin Material Aluminum Aluminum
 Type Louvered Louvered
 Pitch, mm 3.63 3.63
 Thickness, mm 0.152 0.152
 Height, mm 98.35 98.35
 Depth, mm 56.4 47.1

Column x Row 1 x 20 1 x 25
Pass Configuration 4-4-4-4-4 5-4-4-4-4-4
Frontal Area, [m.sup.2] 0.08 0.08
Heat Transfer Area, [m.sup.2] 3.5 2.9
Heat Exchanger Volume, [m.sup.3] 0.0030 0.0025

Minichannel Tube Material Aluminum Aluminum
 Width, mm 37.0 31.0
 Height, mm 1.65 1.65
 Length, mm 400 400
 Vertical Pitch, mm 10.6 8.3
 Port I.D., mm/No. [ea] 0.79/25 0.79/20

Fin Material Aluminum Aluminum
 Type Louvered Louvered
 Pitch, mm 1.54 1.54
 Thickness, mm 0.152 0.152
 Height, mm 98.35 98.35
 Depth, mm 37.0 31.0

Table 8. Modeling Results for the Minichannel Gas Cooler

Tube no./ [T.sub.amb], Capacity, P[D.sub.gc], [T.sub.gc,out],
Fin Pitch [degrees]C W kPa [degrees]C]

20 ea/3.63 mm 24.5 566 0.3 24.5
 32.2 519 0.3 32.8
 40.5 463 0.2 40.5
25 ea/3.63 mm 24.5 566 0.3 24.5
 32.2 520 0.3 32.8
 40.5 463 0.3 40.5
20 ea/1.54 mm 24.5 566 0.4 25.4
 32.2 520 0.4 32.8
 40.5 463 0.3 40.5
25 ea/1.54 mm 24.5 568 0.5 25.1
 32.2 520 0.5 32.8
 40.5 463 0.5 40.5

 HT[C.sub.air]/
Tube no./ AFR, P[D.sub.gc,air], HT[C.sub.ref],
Fin Pitch [m.sup.3]/s Pa W/[m.sup.2] x K

20 ea/3.63 mm 0.0813 1.8 59/362
 0.0813 1.9 59/296
 0.0813 2.0 59/300
25 ea/3.63 mm 0.0813 1.6 68/363
 0.0813 1.7 67/297
 0.0813 1.8 67/301
20 ea/1.54 mm 0.0813 6.5 75/354
 0.0813 6.6 75/293
 0.0813 6.8 74/299
25 ea/1.54 mm 0.0813 6.6 85/386
 0.0813 6.7 85/294
 0.0813 6.9 84/300

Table 9. Detailed Specifications of Two Evaporators

Column x Row 2 x 6 3 x 4
Heat transfer area, [m.sup.2] 1.3 1.4

Tube Material Copper Copper
 Outside Diameter, mm 9.52 9.52
 Thickness, mm 0.6 0.6
 Vertical Pitch, mm 25 25
 Horizontal Pitch, mm 21.65 21.65
 Length, mm 400 400
Fin Material Aluminum Aluminum
 Type Louvered Louvered
 Pitch, mm 3.63 3.63
 Thickness, mm 0.152 0.152
 Height, mm 150 100
 Depth, mm 43.3 65.0

Table 10. Modeling Results for the Evaporator

 [T.sub.amb], Capacity, P[D.sub.ev], [T.sub.ev,out],
Column x Row [degrees]C W kPa [degrees]C

2 x 6 24.5 398 0.3 1.7
Heat transfer 32.2 327 0.3 2.5
 area:
1.4 [m.sup.2] 40.5 249 0.2 2.9
3 x 4 24.5 397 0.3 -2.2
Heat transfer 32.2 327 0.3 0.2
 area:
1.4 [m.sup.2] 40.5 249 0.2 1.8

 HT[C.sub.air]/
 AFR, P[D.sub.ev,air], HT[C.sub.ref],
Column x Row [m.sup.3]/s Pa W/[m.sup.2] x K

2 x 6 0.053 7 92/816
Heat transfer 0.053 7 92/649
 area:
1.4 [m.sup.2] 0.053 7 92/475
3 x 4 0.053 19 98/829
Heat transfer 0.053 19 98/664
 area:
1.4 [m.sup.2] 0.053 19 98/475

Table 11. Heat Transfer Rate per Column ([T.sub.amb] = 32.2[degrees]C)

Row 1 2 3

2 x 6, Louvered fin 9% 91% --
3 x 4, Louvered fin 4% 40% 56%

Table 12. Comparison of Cabinet Load and Cooling Capacity of Condensing
Unit

Ambient Temperature, [degrees]C 24.5 32.2 40.5
Cabin Temperature, [degrees]C 3.3 3.3 3.3
Light, W 15 15 15
Cabinet Heat Transfer Load Through Wall, W 68.0 92.7 119.3
Total Cabinet Thermal Load, W 83.0 107.7 134.3
Cooling Capacity, W 394 305 235
Capacity/Load Ratio 4.75 2.83 1.75
COPYRIGHT 2007 American Society of Heating, Refrigerating, and Air-Conditioning Engineers, Inc.
No portion of this article can be reproduced without the express written permission from the copyright holder.
Copyright 2007 Gale, Cengage Learning. All rights reserved.

Article Details
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Author:Hwang, Yunho; Singh, Varun; Radermacher, Reinhard
Publication:HVAC & R Research
Geographic Code:1USA
Date:May 1, 2007
Words:5212
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