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4.1 Cycle Assumptions Chapter 4: Real Cycle Considerations The ideal cycle analysis in the previous chapter neglects practical considerations such as compressor losses and finite area heat exchangers. Based on several assumptions these effects can be included for a more realistic comparison of R410A and R744. The goal is to make assumptions that place practical limits on the cycle, but do not impose restrictions that would favor one refrigerant or the other. In this chapter the effect of using a fixed vs. variable capacity compressor with regard to cycle performance is briefly explored, then the trade-off between comfort, efficiency and heat exchanger sizing are more fully developed for the air conditioning and heat pump cycles independently. As in the ideal cycle analysis the modeled heat exchangers are of a counterflow configuration. It is assumed that the heat exchange areas of the R410A and R744 heat exchangers are identical and in both cases microchannel heat exchangers are used. Except as noted the air and refrigerant side pressure drops are neglected since the optimum heat exchanger design for these parameters would be different depending on the refrigerant. In Table 4.1 the assumptions used in the following analysis are compared with parameters from a commercially available R410A system and a proposed prototype R744 system. Since many of the values given in Table 4.1 depend on operating conditions (pressure drop, area/kW capacity, etc.), the values given are based on the heat pump rating condition of an outdoor temperature of 8.3oC. The air side heat transfer coefficients are based on values typical for microchannel heat exchangers, 90 W/m2 K at 0.052 kg/s airflow rate per kW capacity (Yin, 2000), and the assumption that the coefficient varies as Re0.8 . In the following analysis the R744 cycle includes an internal heat exchanger where the R410A cycle does not. Compared to R410A, R744 has much higher evaporative heat transfer coefficients (Kirkwood et al., 1999 and Hihara, 2000), and, as a result, the temperature difference between the heat exchanger wall and the refrigerant would be smaller for R744 compared to R410A. This means that the evaporating temperature could be higher for R744 as compared to R410A. Based on the ratio of areas and heat transfer coefficients for the calculated R410A and R744 cycles with comfort constraints shown in Table 4.1, a 1.5oC refrigerant-wall temperature difference may be expected for R410A, while R744 could operate at an evaporating temperature more than 1oC higher. In practice, however, about 0.5 oC of this difference would be dissipated due to pressure drop in a suction line heat exchanger, which would likely be present in an R744 system. Due to the slope of the vapor pressure curve for R410A, the pressure drop penalty would probably negate any COP advantage, so none is assumed to be present in the R410A analysis presented here. 22PDF Image | Comparison of R744 and R410A
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