Document XR895geaKooB88VX3v0v08oNB

548 CHAPTER 38 ,1959 Guide. Sat. Temp. F Afas.Prvss.Lb per Sq to. Table 3 .... Properties of TridUoromoqofhforomemane (CGjF)1 Volume Enthalpy rfha(py and Entropy Total from --40 F 6dropy 25 F Superheat liquid Vapor Liquid Vapor Liquid Vapor Enthalpy Entropy 0 5 10 15 20 . 25 2.59 2.96 3.38 3.85 4.36 .4.94 0.01020 0.01024 0.01028 0.01032 0.01036 0.01040 13.700 12.100 10.700 9.530 8.490 7.580 7.81 8.81 9.82 10.80 11.90 12.90 90.4 91.2 92.0 92.8 93.7 94.6 0.0178 0.0200 0.0222 0.0243 0.0264 0.0286 . 0.1975 0.1974 0.1973 0.1971 0.1970 0.1969 93.9 94.7 95.5 96.3 97.2 98.0 0.2049 0.2047 0.2045 0.2043 0.2041 0.2039 50 F Superheat Enthalpy Entropy 97.4 97.2 99.0 99.8 100.7 101.5 0.2120 0.2117 0.2114 0.2111 0.2109 0.2107 30 35 40 45 < SO 5.57 6.27 7.03 7.88 8.79 0.01045 0.01049 0.01053 0.01057 0.01062 6.770 6.080 5.460 4.920 4.440 13.90 14.90 16.00 17.00 18.10 95.3 96.1 96.8 97.6 98.4 0.0307 0.0328 0.0349 0.0370 0.0391 0.1969 0.1968 0.1968 0.1967 0.1967 98.8 99.6 100.3 101.1 101.9 0.2038 0.2037 0.2036 0.2035 0.2034 102.3 103.1 103.8 104.6 105.4 0.2105 0.2103 0.2101 0.2099 0.2098 65 9.80 0.01066 4.020 19.10 99.2 . 0.0412 0.1967 102.7 0.2033 106.2 0.2097 >00 10.90 0.01071 3.640 20.20 100.0 0.0432 0.1967 103.5 0.2033 107.0 0.2096 65 12.10 0.01076 3.300 21.30 100.8 0.0453 0.1967 104.3 0.2032 107.8 0.2094 70 13.40 0.01061 3.000 22.40 101.5 0.0473 0.1967 105.0 0:2032 108.5 0.2093 <75 14.80 0.01086 2.740 23.50 102.2 0.0493 0.1967 105.7 0.2031 . 109.2 0.2092 80 85. 90 95 100 105 , 16.30 17.90 19.70 21.60 23.60 25.90. 0.01091 0.01096 0.01101 0.01106 0.01111 0.01116 2.500 2.280 2.090 1.918 1.761 1.620 24.50 25.60 26.70 27.80 28.90 30.10 102.9 103.6 104.4 105.1 105.7 106.4 0.0513 0.0533 0.0553 0.0573 0.0593 0.0613 0.1966 0.1966 0.1966 0.1966 0.1965 0.1965 106.4 107.1 107.9 108.6 109.2 109.9 0.2030 0.2029 0.2028 0.2028 0.2027 0.2026 109.9 110.6 111.4 112.1 112.7 113.4 0.2090 0.2089 0.2088 0.2087 0.2085 0.2084 . of liquid leaving the condenser from the enthalpy of super heated vapor going into it, this, <3/ -- 26,3 (89.34 - 29.68) - 1S69 Btu per minute.. .. (e)'. The cooling water rate (based on a gallon as 8.34 lb) is 1569 +' (8 X 854) - 235 gpm. t . (J) The compressor aixe is fixed by the volume of gas which must be drawn into the machine per unit time. Saturated vapor at 62.7 psia has a specific volume, from Table 1, of 0.779 cu ft per pound, hence 265 X 0.779 * 20.49 cfmof gas must be han dled. Assuming a volumetric efficiency of 90 percent, the com pressor must then displace 20.49 + 0.9 -- 225 cfm. The speed is given as 500 rpm and, as the unit is known to be double-act ing, the displacement is therefore (225 X 1728) + (2 X 500) ~ 39.4'cu in. lf-the unit were designed so that bore d and stroke were the same, . (W) V,4 - 39.4 d - 3.69 in. , . () . (CP) - (h,, - h/t) + (h, - A,.) - (8252 - 29.68) + (8954 - 8252) - 8.17 where &/c is the specifio enthalpy of liquid at discharge from the condenser. The coefficient of performance of Example 1 may be com pared with that of an ideal system operating on the Carnot cycle between the same temperature limits. Then T, = 501 F (which is 41 F + 460) and Tc = 554 F (which is 94 F + 460) and; The actual cycle is therefore 8.17 '4- 9.6 or 85 percent as effective as a Carnot cycle between the same temperature limits. Influence of Suction Pressure Brief consideration of the analytical procedure used in dis cussion of riie ample saturation cycle will bring out the need for maintaining the suction pressure on any refrigeration system as high as the load will permit. As the suction pressure increases, for fixed discharge pressure, the enthalpy of re frigerant entering the evaporator remains unchanged, but the leaving enthalpy increases and, hence, the refrigerating effect increases. Further, compressor energy input is reduced not merely because of the greater enthalpy of the gas at suction, but also because of a reduction in the enthalpy of the super heated gas at discharge. Since the refrigerating effect is greater and the work-less, it is obvious that there will be a substantial gain in the coefficient of the performance. See Chapter.54,'Fig. 1. The actual value of suction pressure on any system is ob viously determined by the required temperature which must be' maintained in the conditioned space. For a direct-expan sion system the evaporator can be held at a temperature not much less than that of the conditioned enclosure, except in cases where lower temperatures may be needed in order to establish a desired ratio of dehumidifying to cooling load. When dehumidification requirements dictate the use of un usually low evaporator temperatures, the increased operating cost should properly be charged against the dehumidification rather than the sensible cooling. Refrigeration 549 Temp Pressure pm Table 4 .... Properties of Tridilorotriftuoroethane (CjCl*F*)" Volume Density Enthalpy from --40*F liquid cu ft/lb Vapor at ft/lb Liquid (b/cu ft Vapor ib/cu ft Liquid Btu/lb latent Btu/lb Vapor Btu/lb Entropy from -- 40*F Liquid F deg Vapor F deg 0 0.8377 28.21* 0.00966 31.31 4 .9503 27.99* .00968 27.84 8 1.075 27.73* .00971 24.81 12 1.213 27.45* .00974 22.17 16 1.366 27.14* .00977 19.84 20 1.634 26.80* 0.00979 17.81 24 1.719 26.42* .00982 16.02 28 1.922 26.01* .00985 14.43 32 2.145 25.55* .00988 13.03 36 2.388 25.06* .00991 11.79 40 2.655 24.52* 0.00994 10.68 44 2.944 23.93* .00997 9.703 48 3.258 23.29* .01000 8.830 52 3.602 22.59* .01003 8.044 56 3.973 21.83* .01006 7.342, 60 4.374 21.02* 0.01010 6.713 64 4.807 20.14* .01013 6.149 68 5.275 19.18* .01016 5.640 72 5.780 18.16* .01019 . 5.180 76 6.320 17.06* .01023 4.769 80 6.902 15.87* 0.01026 4.392 84 7.527 14.60* .01030 4.05V 88 8.194 13.24* .01033 3.742 92 8.908 11.79* .01037 3.463 96 9.668, 10.24* .01040 3.208 100 10.48 104 11.35 108 12.28 110 12.70 8.59* 6.82* 4.93* 3.95* 0.01044 .01048 .01051 0.01053 2.976 2.762 2.567 - 2.477 * Inches of mercury below one ebnoepbere. t StuuUrd cycle tempenturm. ASilE declination--Rdricorant 1U, 103.56 103.27 102.98 102.69 102.40 102.10 101.81 101.51 101.21 100.91 100.60 100.30 99.99 99.68 99.37 99.05 98.73 98.42 98.10 97.77 97.45 97.12 96.79 96.46 96.13 95.79 95.46 95.12 94.95 0.03194 .03592 .04031 .04511 .05040 7.98 8.78 9.59 10.41 11.22 0.05616 .06243 .06929 .07675 .08483 12.03 12.85 13.67 14.49 16.32 0.09361 .1031 .1133 .1243 .1362 16.16 16.99 17.82 18.66 - 19.50 0.1490 .1626 .1773 .1931 .2097 20.35 21.19 22.05 22.90 23.76 0.2277 .2468 .2672 .2888 .3117 24.63 25.49 26.36 27.24 28.11 0.3360 .3620 .3896 0.4038 28.99 29.89 30.78 31.22 70.92 70.68 70.44 70.20 69.96 69.72 69.48 69.24 69.00 68.75 68.50 68.25 68.00 67.74 -67.48 67.22 66.96 66.69 66.43 66.16 65.88 65.60 65.32 65.04 64.75 64.46 64.16 63.86 63.71 78.89 79.46 80.03 80.61 81.18 0.0182 .0199 .0216 .0234 .0251 0.1725 .1724 .1723 .1722 .1722 81.75 82.33 82.91 83.49 84.07 .0.0268 .0285 .0302 .0318 .0335 0.1722 .1721 .1722 .1722 .1722 84.65 85.24 85.82 86.40 86.98 0.0352 .0368 .0385 .0401 .0418 0.1723 .1723 .1724 .1728 .1727 87.57 88.15 88.74 89.33 89.92 0.0434 - .0450 .0467 .0483 .0499 0.1728 .1729 .1731 .1732 .1734 90.51 91.09 91.68 92.28 92.86 0.0515 .0531 .0547 .0563 .0578 0.1736 .1738 .1740 . .1742 .1744 93.45 94.05 94.64 94.93 0.0594 .0610 .0626 0.0634 0.1746 .1748 .1751 0.1752 Influence of Discharge Pressure In contrast to suction pressure, the compressor discharge pressure should be kept as low as operating conditions will allow. This pressure must be high enough to provide a saturation temperature of refrigerant within the condenser that is greater than the exit temperature of the cooling water. The discharge pressure therefore is a direct function of the temperature of the cooling fluid, and will automatically rise whenever the temperature of cooling water (or air) rises; it will also rise when the flow rate of the :cooling medium is decreased. Increase in discharge pressure (for fixed suction pres sure) raises the enthalpy of the gas leaving the compressor; hence, increases the work of compression. Further, as the enthalpy of saturated liquid leaving the condenser increases with pressure, the refrigerating effect must decrease. Thus the effect of such a pressure rise is to require more work per pound of refrigerant handled, and at the same time to necessitate an increase in the refrigerant flow rate. (See Chap ter 54, Fig. 1.) Influence of Water Jacket The preceding discussion has, in every case, assumed lsentropic compression. Where exact performance data are not available, this assumption is a desirable one since it leads to a conservatively large determination of the power required. In most actual systems, the compression process departs from isentropic due to irreversible heat transfers which oc cur between the vapor in the cylinder and the cylinder wall, and also because of intentional heat dissipation from the outside of the cylinder walls to the surroundings,- or to a cooling fluid passing through a water jacket around the