Document NeObgY1bzKjam4zgpJ3xxNdqb
566
CHAPTER 38
1960 Guide
Sat. Tamp. F
Abi. Pftu. Lb par Sq In.
Table 3 .... Properties of Trichloromonoftuoromethane (CC1*F)"
Vsbnt
Enthalpy
Enthalpy and Entropy Takan fro*i -40 F
Entropy
25 f Superheat
Liquid
Vapor
liquid
Vapor
liquid
Vapor
Enthalpy Entropy
50 F Superheat Enthalpy Entropy
0
2.59
0.01020 13.700
7.81
90.4
0.0178
0.1975
93.9
0.2049
97.4
0.2120
5
2.96
0.01024 12.100
8.81
91.2
0.0200
0.1974
94.7
0.2047
97.2
0.2117
10
3.38
0.01028 10.700
9.82
92.0
0.0222
0.1973
95.5
0.2045
99.0
0.2114
15
3.85
0.01032
9.530 10.80
92.8
0.0243
0.1971
96.3
0.2043
99.8
0.2111
20
4.36
0.01036
8.490 11.90
93.7
0.0264
0.1970
97.2
0.2041
100.7
0.2109
25
4.94
0.01040
7.580 12.90
94.5
0.0286
0.1969
98.0
0.2039
101.5
0.2107
30
5.57
0.01045
6.770
13.90
95.3
0.0307
0.196ft
98.8
0.2038
102.3
0.2105
35
6.27
0.01049
6.080
14.90
96.1
0.0328
0.1968
99.6
0.2037
103.1
0.2103
40
7.03
0.01053
5.460
16.00
96.8
0.0349
0.1968
100.3
0.2036
103.8
0.2101
45
7.88
0.01057
4.920
17.00
97.6
0.0370
0.1967
101.1
0.2035
104.6
0.2099
50
8.79
0.01062
4.440 18.10
98.4
0.0391
0.1967
101.9
0.2034
105.4
0.2098
55
9.80
0.01066
4.020 19.10
99.2
0.0412
0.1967
102.7
0.2033
106.2
0.2097
60
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
0.01081
3.000 22.40
101.5
0.0473
O.1967
105.0
0.2032
10S.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
16.30
0.01091
85
17.90
0.01096
90
19.70
0.01101
95
21.60
0.01106
100
23.60
0.01111
105
25.90
0.01116
* ASHRAE deuigmUoa--Rcfcigtreat 11.
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 1099
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, thus,
Qt - 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 -i- (8 X 8.34) - 23.5 gpm.
(f) 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 2o-3 X 0.779 = 20.49 cfm of gas must be han dled- Assuming a volumetric efficiency of 90 percent, the com pressor 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) + (2 X 500) 39.4 cu in. If the unit were designed so that bore d and stroke were the same,
(wd*) + .4 - 39.4 d - 3.69 in.
(?) (CP) - (K< - A,,) + (A, - A..) = (82.82 - 29.68) + (89-34 - 82.82) - 8.17
where Arc is the specific enthalpy of liquid at discharge from the condenser.
Tbe 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 Tt = 501 F (which is 41F + 460) aod Tc = 554 F (which is 94 F + 460) and,
The actual cycle is therefore 8.17 ~ 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 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 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
Temp
Praam*
' Table A .... Properties of Trichlorotrifluoroethane (CjCljFj)*
Volume
Density
Enthalpy from -- 40*F
F
psia
psifl
Uquid cu ft/tb
Vapor cu ft/lb
liquid Ib/cu ft
Vapor tb/cu ft
liquid Btu/lb
Lotent Btu/lb
Vapor Btu/lb
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
1.534
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* 001010
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
4.051
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 12.76
8.59* 6.82* 4.93* 3.95*
0.01044 .01048 .01051
0.01053
* Incite* of mercury below oae eUzusphere. t Btendeni cyde tempenum*. * ASHRAE desixaetioo--Relricenst.lU.
2.976 2.762 2.567 2.477
103.50 103.27 102.98 102.69 102.40
0.03194 .03592 .04031 .04511 .05040
7.98 8.78 9.59 10.41 11.22
102.10 202.81 101.51 101.21 100.91
0.05616 .06243 .06929 .07675 .08483
12.03 12.85 13.67 14.49 15.32
100.60 100.30 99.99
99.68 99.37
0.09361 .1031 .1133 .1243 .1362
16.16 16.99 17.82 18.66 19.50
99.05 98.73 98.42 98.10 97.77
0.1490 .1626 .1773 .1931 .2097
20.35 21.19 22.05 22.90 23.76
97.45 97.12 96.79 96.4696.13
0.2277 .2468 .2872 .2888 .3117 *
24.63 25.49 26.36 27.24 28.11
95.79 95.46 95.12 94.95
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
81.75 82.33 82.91 83.49 84.07
84.65 85.24 85.82 86.40 86.98
87.57 88.15 88.74 89.33 89.92
90.51 91.09 91.68 92.28 92.86
93.45 94.05 94.64 94.93
567
Entropy from --40*F
liquid Btu/lb/ F deg
Vapor Btu/lb/ F deg
0.0182 .0199 .0216 .0234 .0251
0.1725 .1724 .1723 .1722 .1722
0.0268 .0285 .0302 .0318 .0335
0.1722 .1721 .1722 .1722 .1722
0.0352 .0368 .0385 .0401 .0418
0.1723 .1723 .1724 .1726 .1727
0.0434 .0450 .0467 .0483 .0499
0.1728 .1729 .1731 .1732 .1734
0.0515 .0531 .0547 .0563 .0578
0.1736 .1738 .1740 .1742 .1744
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 thaD 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 isentropic 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 beat transfers which oc cur between the vapor in tbe cylinder and the cylinder walk 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