Document 2Gwk2x2d8wv6bqwjraRpzYQp

784 CHAPTER 36 1951 Guide 0.17330, BO the Buperheat, lmi, possessed by the actual gas discharged from this com pressor can be obtained by interpolation aa, Ui _ 0.16828 - 0.16603 25 " 0.17330 - 0.16608 from which U& = 7.6 deg. As the saturation temperature at 121 psia is 94 F the actual temperature, id, of the vapor leaving the compressor is, id -- 94 4* tad -- 94 + 7.6 = 101.6 F. By the same kind of interpolation the enthalpy of the discharged vapor can be determined from the enthalpies given for vapor superheated 25 F and for saturated vapor, (Ad - 88.10) (0.16828 - 0 16608) (92.16 - 88.10) " (0.17330 - 0.16608) from which, Ad = 89.34 Btu per pound. Then substituting in Equation 7, (hp) = 26.3 (89.34 - 82.82) 42.42 = 4.03. (d) The rate of heat loss from the condenser, Qc, must be equal to the sum of the energies picked up by the refrigerant in the evaporator and the compressor, Q,, = 53.14 + (89.34 - 82.82) '= 53.14 + 6.52 = 59.66 Btu per pound or 26.3 X 59.66 = 1569 Btu per minute. This same figure can, of course, be determined more Fig. 3. Pressure-Enthalpy Diagram for Simple Saturation Cycle directly by subtraction of the enthalpy of liquid leaving the condenser from the en thalpy of superheated vapor going into it, thus, Q. = 26.3 (89.34 -- 29.68) = 1569 Btu per minute. (e) The cooling water rate (based on a gallon as 8.34 lb) is 1569 -P (8 X 8.34) -- 23.5 gpm. N (/) The compressor size is fixed by the volume of gas which must be drawn into the machine per unit time. Saturated vapor at 52.7 psia has a specific volume, from Table 1, of 0.779 cu ft per pound, hence 26.3 X 0.779 = 20.49 cfm of gas must be handled. Assuming a volumetric efficiency of 90 percent, the compressor must then- displace 20.49 -5- 0.9 = 22.8 cfm. The speed is given as 500 rpm and, as the unit is known to be double-acting, the displacement is therefore (22.8 X 1728) -5- (2 X 500) = 39.4 cu in. If the unit were designed so that bore d and stroke were the same, (ird1) + 4 = 39.4 d = 3.69 in. (ff) (CP) = (Ar. - A,,,) + (Ad -- A,.) = (82.82 - 29.68) -f- (89.34 -- 82.82) = 8.17 where ht, is the specific enthalpy of liquid at discharge from the condenser. The coefficient of performance of Example 1 may be compared with that of an ideal system operating on the Carnot cycle between the same tempera- Refrigeratlon 785 ture limits. Then T. = 501 F (which is 41 F -f- 460) and T,, = 554 F (which is 94 F + 460) and, 501 (CP) = = 9.6 554 - 501 The actual cycle is therefore 8.17 -5- 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 discussion of the simple 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 refrigerant 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 en thalpy of the superheated 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. The actual value of suction pressure on any system is obviously de termined by the required temperature which must be maintained in the conditioned space. For a direct expansion 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 es tablish a desired ratio of dehumidifying to cooling load. When dehumidi fication requirements dictate the use of unusually low evaporator temperatures,' the increased operating cost should properly be charged against the dehumidification rather than the sensible cooling. Influence of Discharge Pressure In contrast to the 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 which 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 pressure) raises the enthalpy of the gas leaving the compressor; hence, increases the work of compression. Further, as the enthalpy of saturated liquid leaving the con denser 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. Influence of Water Jacket The preceding discussion has, in every case, assumed isentropic com pression. 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 occur