Steam Production

39. Steam Production The fibre in the cane is generally sufficient to enable the quantity of bagasse produced by the mills to supply all the steam necessary for power production and for manufacture, when utilised as fuel in the boiler furnaces. With a normal fibre content (12-14%) and a well balanced and well designed factory, there remains in addition an excess of bagasse (or of steam) which may be used for other purposes. We may comment in passing, that the cane, which produces the sugar, supplies at the same time the fuel necessary for the factory which handles it; and also leaves, as by-products or waste-products: (a) Molasses, from which is made either rum, or motor fuel for tractors and lorries. (b) The muds, which form an excellent manure for the fields. (c) Ashes, produced by combustion of bagasse in the furnaces, and which supplement the muds in a most convenient fashion, to supply a complete fertiliser. Finally, it should not be forgotten that the cane itself provides the water necessary for steam production, which is very valuable for factories which do not have access to an unlimited supply of fresh water. We shall study successively: (a) Steam. (b) Bagasse. (c) The combustion of bagasse. (d) The boilers and their accessories. (e) The general conditions for steam production in the sugar factory. STEAM Tables

The principal properties of steam are given (in both British and metric units on account of their importance) in the following tables: (a) Table 149. This gives the properties of dry saturated steam. There exists a definite relationship between the pressure and the temperature of dry saturated water vapour: if one of these properties is known the other is immediately fixed. However, the liquid phase ceases at the critical point, which corresponds to a temperature of 374°C (705°F) and a pressure of 225 kg/cm2 (3200 p.s.i.); above this temperature, water can exist only in the gaseous form. It will be noted that the total heat necessary to form unit weight of steam by no means increases proportionately to the pressure. It increases at first, though slowly, passes through a maximum in the neighbourhood of 450 p.s.i., and then decreases:

39

659

TABLES

TABLE 149A PROPERTIES OF DRY SATURATED STEAM (METRIC UNITS)

I

=

temperature in °C

A = iii =

q = sensible heat to raise 1 kg of water from O°C to 1°C in kcal/kg p

d

=

r

p = absolute pressure of the vapour, in kg/cm 2 p' = gauge pressure of the vapour, in kg/cm 2 d = density of water at tOC, in kg/dm 3

v

=

latent heat of vaporisation of water at tOC in kca1jkg q + r = total heat of 1 kg of water at to density of vapour at to, in :kg/m 3 specific volume of vapour at to, in rn 3 /kg

q

r

A

w

l'

0 5 10 15 16 17 18 19 20

0.00623 0.00889 0.01251 0.01738 0.01853 0.01975 0.02103 0.02239 0.02383

0.99987 0.99999 0.99973 0.99913 0.99897 0.99880 0.99862 0.99843 0.99823

0 5 10 15 16 17 18 19 20

597.2 594.4 591.6 588.8 588.3 587.7 587.1 586.6 586

597.2 599.4 601.6 603.8 604.3 604.7 605.1 605.6 606

0.00485 0.00680 0.00940 0.01282 0.01363 0.01447 0.01536 0.01630 0.01729

206.3 147.2 106.4 77.99 73.39 69.10 65.10 61.35 57.84

21 22 23 24 25 26 27 28 29 30

0.02534 0.02694 0.02863 0.03041 0.03229 0.03426 0.03634 0.03853 0.04083 0.04325

0.99802 0.99780 0.99756 0.99732 0.99707 0.99681 0.99654 0.99626 0.99597 0.99567

21 22 23 24 25 26 27 28 29 30

585.5 584.9 584.3 583,8 583.2 582.6 582.1 581.5 581 580.4

606.5 606.9 607.3 607.8 608.2 608.6 609.1 609.5 610 610.4

0.01833 0.01942 0.02056 0.02177 0.02304 0.02437 0.02576 0.02723 0.02876 0.03036

54.56 51.49 48.63 45.94 43.41 41.04 38.82 36.73 34.77 32.93

31 32 33 34 35 36 37 38 39 40

0.04580 0.04847 0.05128 0.05423 0.05733 0.06057 0.06398 0.06755 0.07129 0.07520

0.99537 0.99505 0.99463 0.99440 0.99406 0.99371 0.99336 0.99299 0.99262 0.99224

31 32 33 34 35 36 37 38 39 40

579.8 579.3 578.7 578.1 577.5 577 576.4 575.9 575.3 574.7

610.8 611.3 611.7 612.1 612.5 613 613.4 613.9 614.3 614.7

0.03204 0.03380 0.03565 0.03758 0.03960 0.04171 0.04392 0.04622 0.04863 0.05114

31.20 29.58 28.05 26.61 25.25 23.97 22.77 21.63 20.56 19.55

41 42 43 44 45 46 47 48 49 50

0.07930 0.08360 0.08809 0.09279 0.09771 0.10284 0.10821 0.11382 0.11967 0.12578

0.99186 0.99147 0.99107 0.99066 0.99024 0.98982 0.98940 0.98896 0.98852 0.98807

41 42 43 44 45 46 47 48 49 50

574.2 573.6 573 572.4 571.8 571.2 570.7 570.1 569.5 569

615.2 615.6 616 616.4 616.8 617.2 617.7 618.1 618.5 619

0.05377 0.05650 0.05935 0.06233 0.06544 0.06867 0.07203 0.07553 0.07918 0.08298

18.60 17.70 16.85 16.04 15.28 14.56 13.88 13.24 12.63 12.05

51 52 53 54 55 56 57 58 59

0.1322 0.1388 0.1457 0.1530 0.1605 0.1684 0.1765 0.1850 0.1939 0.2031

0.98762 0.98715 0.98669 0.98621 0.98573 0.98524 0.98478 0.98425 0.98375 0.98324

50.9 51.9 52.9 53.9 54.9 55.9 56.9 57.9 58.9 59.9

568.4 567.8 567.3 566.7 566.1 565.6 565 564.4 563.8 563.3

619.3 619.7 620.2 620.6 621 621.5 621.9 622.3 622.7 623.2

0.0869 0.0910 0.0953 0.0997 0.1043 0.1091 0.1141 0.1193 0.1247 0.1302

11.50 10.98 10.49 10.02 9.584 9.164 8.764 8.385 8.025 7.682

60

660

39

STEAM PRODUCTION

TABLE 149A (continued) p

d

q

-

r

A

562.7 562.1 561.5 560.9 560.3 559.7 559.1 558.5 558 557.4

623.6 624 624.4 624.8 625.2 625.6 626 626.4 626.9 627.3

0.1359 0.1419 0.1481 0.1545 0.1611 0.1680 0.1752 0.1826 0.1902 0.1981

7.356 7.046 6.752 6.473 6.206 5.951 5.709 5.478 5.258 5.049

v

w

61 62 63 64 65 66 67 68 69 70

0.2127 0.2227 0.2330 0.2438 0.2550 0.2666 0.2787 0.2912 0.3042 0.3177

0.98272 0.98220 0.98167 0.98113 0.98059 0.98005 0.97950 0.97894 0.97838 0.977~H

60.9 61.9 62.9 63.9 64.9 65.9 66.9 67.9 68.9 69.9

71 72 73 74 75 76 77 78 79 80

0.3317 0.3463 0.3613 0.3769 0.3931 0.4098 0.4272 0.4451 0.4637 0.4829

0.97723 0.97666 0.97607 0.97548 0.97489 0.97429 0.97368 0.97307 0.97245 0.97183

70.9 71.9 72.9 73.9 74.9 75.9 76.9 77.9 78.9 80

556.8 556.2 555.6 555 554.4 553.8 553.2 552.6 552 551.3

627.7 628.1 628.5 628.9 629.3 629.7 630.1 630.5 630.9 631.3

0.2062 0.2146 0.2234 0.2324 0.2418 0.2514 0.2614 0.2717 0.2823 0.2933

4.849 4.658 4.476 4.302 4.136 3.977 3.826 3.681 3.543 3.410

81 82 83 84 85 86 87 88 89 90

0.5028 0.5234 '0.5447 0.5667 0.5894 0.6129 0.6372 0.6623 0.6882 0.7149

0.97121 0.97057 0.96994 0.96930 0.96865 0.96800 0.96734 0.96668 0.96601 0.96534

81 82 83 84 85 86 87 88 89 90

550.7 550.1 549.5 548.8 548.2 547.6 547 546.4 545.7 545.1

631.7 632.1 632.5 632.8 633.2 633.6 634 634.4 634.7 635.1

0.3046 0.3162 0.3282 0.3406 0.3534 0.3666 0.3802 0.3942 0.4086 0.4235

3.283 3.162 3.047 2.936 2.830 2.728 2.630 2.537 2.447 2.361

91 , 92 "93 94 95 96 97 98 99 100

0.7425 0.7710 0.8004 0.8307 0.8619 0.8942 0.9274 0.9616 0.9969 1.0332

0.96467 0.96399 0.96330 0.96261 0.96192 0.96122 0.96051 0.95981 0.95909 0.95838

91 92 93 94 95 96 97 98 99 100

544.5 543.9 543.3 542.7 542 541.4 540.8 540.2 539.5 538.9

635.5 635.9 636.3 636.7 637 637.4 637.8 638.2 638.5 638.9

0.4388 0.4545 0.4707 0.4873 0.5045 0.5221 0.5402 0.5588 0.5780 0.5977

2.279 2.200 2.124 2.051 1.981 1.914 1~851 1.789 1.730 1.673

q

r

A

w

p

p'

--

v

--_._-

101 102 103 104 105 106 107 108 109 110

1.0707 1.1092 1.1489 1.1898 1.2318 1.2751 1.3196 1.3654 1.4125 1.4609

0.0375 0.0760 0.1157 0.1566 0.1986 0.2419 0.2864 0.3322 0.3793 0.4277

101 102 103.1 104.1 105.1 106.1 107.1 108.1 109.1 110.1

538.3 537.6 536.9 536.2 535.6 535 534.3 533.6 533 532.4

639.3 639.6 640 640.3 640.7 641.1 641.4 641.7 642.1 642.5

0.6179 0.6387 0.6601 0.6820 0.7045 0.7276 0.7514 0.7758 0.8008 0.8265

1.618 1.565 1.515 1.466 1.419 1.374 1.331 1.289 1.249 1.210

111 112 113 114 115

1.5106 1.5618 1.6144 1.6684 1.7239

0.4774 0.5286 0.5812 0.6352 0.6907

111.1 112.1 113.2 114.2 115.2

531.8 531.1 530.4 529.7 529.1

642.9 643.2 643.6 643.9 644.3

0.8528 0.8798 0.9075 0.9359 0.9650

1.173 1.137 1.102 1.068 1.036

39

661

TABLES TABLE 149A (continued)

p

p'

q

r

A

--(J)

v

116 117 118 119 120

1.7809 1.8394 1.8995 1.9612 2.0245

0.7477 0.8062 0.8663 0.9280 0.9913

116.2 117.2 118.2 119.2 120.3

528.4 527.8 527.1 526.4 525.7

644.6 645 645.3 645.6 646

0.9947 1.026 1.057 1.089 1.122

1.005 0.9752 0.9462 0.9183 0.8914

121 122 123 124 125 126 127 128 129 130

2.0895 2.1561 2.2245 2.2947 2.3666 2.4404 2.5160 2.5935 2.6730 2.7544

1.0563 1.1229 1.1913 1.2615 1.3334 1.4072 1.4828 1.5603 1.6398 1.7212

121.3 122.3 123.3 124.3 125.3 126.4 127.4 128.4 129.4 130.4

525.1 524.4 523.7 523.1 522.4 521.6 520.9 520.3 519.6 518.9

646.4 646.7 647 647.4 647.7 648 648.3 648.7 649 649.3

1.156 1.190 1.225 1.262 1.299 1.337 1.376 1.415 1.455 1.496

0.8655 0.8404 0.8161 0.7927 0.7701 0.7482 0.7271 0.7068 0.6871 0.6680

131 132 133 134 135 136 137 138 139 140

2.8378 2.9233 3.011 3.101 3.192 3.286 3.382 3.481 3.582 3.685

1.8046 1.8901 1.978 2.068 2.159 2.253 2.349 2.448 2.549 2.652

131.4 132.5 133.5 134.5 135.5 136.6 137.6 138.6 139.6 140.6

518.2 517.4 516.7 516 515.3 514.6 513.9 513.3 512.6 511.9

649.6 649.9 650.2 650.5 650.8 651.2 651.5 651.9 652.2 652.5

1.539 1.583 1.628 1.673 1.719 1.767 1.815 1.864 1.915 1.967

0.6496 0.6318 0.6146 0.5979 0.5817 0.5661 0.5510 0.5363 0.5221 0.5084

141 142 143 144 145 146 147 148 149 150

3.790 3.898 4.009 4.122 4.237 4.355 4.476 4.599 4.725 4.854

2.757 2.865 2.976 3.089 3.204 3.322 3.443 3.566 3.692 3.821

141.7 142.7 143.7 144.8 145.8 146.8 147.8 148.9 149.9 150.9

511.1 510.4 509.7 508.9 508.2 507.5 506.8 506 505.3 504.6

652.8 653.1 653.4 653.7 654 654.3 564.6 654.9 655.2 655.5

2.020 2.074 2.129 2.185 2.243 2.302 2.362 2.423 2.485 2.548

0.4951 0.4823 0.4698 0.4577 0.4459 0.4345 0.4235 0.4128 0.4024 0.3924

p

p'

q

r

A

(J)

-

v

5 5.5 6 6.5 7 7.5 8 8.5 9 9.5

3.967 4.467 4.967 5.467 5.967 6.467 6.967 7.467 7.967 8.467

151.11 154.71 158.08 161.21 164.17 166.96 169.61 172.12 174.53 176.83

152.1 155.8 159.3 162.5 165.6 168.5 171.3 173.9 176.4 178.9

503.7 501.1 498.5 496.1 493.8 491.6 489.5 487.5 485.6 483.6

655.8 656.9 657.8 658.6 659.4 660.1 660.8 661.4 662 662.5

2.621 2.867 3.112 3.356 3.600 3.842 4.085 4.327 4.568 4.809

0.3816 0.3488 0.3213 0.2980 0.2778 0.2603 0.2448 0.2311 0.2189 0.2080

10 10.5 11 11.5 12 12.5 13 13.5 14 14.5

8.967 9.467 9.967 10.467 10.967 11.467 11.967 12.467 12.967 13.467

179.04 181.16 183.20 185.17 187.08 188.92 190.71 192.45 194.13 195.77

181.2 183.4 185.6 187.7 189.7 191.6 193.5 195.3 197.1 198.9

481.8 480.1 478.3 476.6 475 473.5 471.9 470.4 468.9 467.4

663 663.5 663.9 664.3 664.7 665.1 665.4 665.7 666 666.3

5.049 5.290 5.530 5.770 6.010 6.249 6.488 6.728 6.967 7.207

0.1981 0.1891 0.1808 0.1733 0.1664 0.1600 0.1541 0.1486 0.1435 0.1388

662

39

STEAM PRODUCTION

TABLE 149A (continued) q

r

A

w

197.36 200.43 203.35 206.14 208.81 211.38 213.85 216.23 218.53 220.75

200.6 203.9 207.1 210.1 213 215.8 218.5 221.2 223.6 226.1

466 463.2 460.4 457.8 455.2 452.7 450.2 447.7 445.5 443.2

666.6 667.1 667.5 667.9 668.2 668.5 668.7 668.9 669.1 669.3

7.446 7.925 8.405 8.886 9.366 9.846 10.33 10.81 11.29 11.78

0.1343 0.1262 0.1190 0.1126 0.1068 0.1016 0.09682 0.09251 0.08856 0.08492

23.967 24.967 25.967 26.96727.967 28.967 30.967 32.967 34.967 38.967

222.90 224.99 227.01 228.98 230.89 232.76 236.35 239.77 243.04 249.18

228.5 230.8 233 235.2 237.4 239.5 243.6 247.5 251.2 258.2

440.9 438.7 436.6 434.4 432.3 430.2 426.1 422.1 418.3 410.8

669.4 669.5 669.6 669.6 669.7 669.7 669.7 669.6 669.5 669

12.26 12.75 13.23 13.72 14.21 14.70 15.69 16.68 17.68 19.69

0.08157 0.07846 0.07557 0.07288 0.07037 0.06802 0.06375 0.05995 0.05658 0.05078

43.967 48.967 58.967 78.967 98.967 148.967 198.967 224.567

256.23 262.70 274.29 293.62 309.53 340.56 364.08 374.15

266.5 274.2 288.4 312.6 334 381.7 431.4 501.5

401.7 393.1 376.6 346.3 317.1 243.2 150.7 0

668.2 667.3 665 658.9 651.1 624.9 582.1 501.5

22.25 24.85 30.21 41.60 54.21 93.90 161.2 315

0.04495 0.04024 0.03310 0.02404 0.01845 0.01065 0.00620 0.00318

p

p'

15 16 17 18 19 20 21 22 23 24

13.967 14.967 15.967 16.967 17.967 18.967 19.967 20.967 21.967 22.967

25 26 27 28 29 30 32 34 36 40 45 50 60 80 100 150 200 225.6

v

TABLE 149B PROPERTIES OF DRY SATURATED STEAM (BRITISH UNITS)

t = temperature in OF p = absolute pressure of the steam, p.s.i.a. p' = gauge pressure of the steam, p.s.i.g., d = density of water at rOF, in lb./cu.ft. q

r

= sensible heat to raise 1 lb. of water from 32° to rOF in B.Th.U./lb. p

d

q

latent heat of vaporisation of water at tOp, in B.Th.V./lb. l = q + r = total heat of 1 lb. of" water at rOF iii = density of vapour at tOp, in lb./cu.ft. v = specific volume of vapour at rOF, in cu.ft./lb. =

r

l

-

w

v

32 40 50 60

0.0886 0.1217 0.1780 0.2563

62.420 62.428 62.411 62.368

0 8 18.1 28.1

1075 1071.3 1064.9 1059.3

1075 1079.3 1083 1087.4

0.000303 0.000409 0.000587 0.000828

3305 2445 1705 1208

62 64 66 68 70

0.2751 0.2951 0.3163 0.3389 0.3630

62.357 62.344 62.331 62.318 62.303

30.1 32.1 34.1 36.1 38.1

1058.2 1057.1 1056 1054.8 1053.7

1088.3 1089.2 1090.1 1090.9 1091.8

0.000886 0.000947 0.001011 0.001079 0.001152

1130 1057 989.2 926.5 868.4

72 74 76 78 80

0.3886 0.4156 0.4444 0.4749 0.5070

62.287 62.271 62.254 62.235 62.218

40.1 42 44 46 48

1052.6 1051.5 1050.4 1049.2 1048.1

1092.7 1093.5 1094.4 1095.2 1096.1

0.001228 0.001308 0.001393 0.001483 0.001579

814.5 764.6 717.9 674.2 633.5

39

663

TABLES

TABLE 149B (continued) -

v

d

q

r

A

0.5411 0.5771 0.6152 0.6556 0.6983

62.198 62.178 62.158 62.137 62.113

50 52 54 56 58

1046.9 1045.8 1044.7 1043.6 1042.5

1096.9 1097.8 1098.7 1099.6 1100.5

0.001679 0.001785 0.001896 0.002012 0.002135

595.7 560.4 527.5 496.9 468.3

92 94 96 98 100

0.7434 0.7910 0.8410 0.8939 0.9495

62.092 62.069 62.045 62.021 61.995

60 62 64 66 68

1041.3 1040.1 1039 1037.9 1036.8

1101.3 1102.1 1103 1103.9 1104.8

0.002265 0.002402 0.002545 0.002696 0.002853

441.5 416.4 393 371 350.5

102 104 106 108 110

1.0080 1.0696 1.1347 1.2033 1.2752

61.970 61.944 61.917 61.890 61.862

70 72 74 75.9 77.9

1035.7 1034.5 1033.4 1032.2 1031

1105.7 1106.5 1107.4 1108.2 1109

0.003019 0.003193 0.003376 0.003567 0.003767

331.2 313.2 296.3 280.4 265.5

112 114 116 118 120

1.3509 1.4304 1.5137 1.6012 1.6929

61.833 61.804 61.775 61.745 61.714

79.9 81.9 83.9 85.9 87.9

1029.8 1028.7 1027.6 1026.4 1025.2

1109.8 1110.6 1111.5 1112.3 1113.1

0.003977 0.004197 0.004427 0.004666 0.004918

251.4 238.3 225.9 214.3 203.4

122 124 126 128 130

1.789 1.890 1.996 2.107 2.223

61.683 61.652 61.619 61.587 61.554

89.9 91.9 93.9 95.9 97.9

1024.2 1023.1 1021.9 1020.8 1019.6

1114.1 1115 1115.8 1116.7 1117.5

0.005180 0.005455 0.005742 0.006042 0.006353

193 183.3 174.2 165.5 157.4

132 134 136 138 140

2.345 2.472 2.605 2.744 2.889

61.520 61.486 61.452 61.417 61.382

99.9 101.9 103.9 105.9 107.9

1018.5 1017.4 1016.2 1015.1 1013.9

1118.4 1119.3 1120.1 1121 1121.8

0.006676 0.007016 0.007371 0.007741 0.008127

149.8 142.5 135.7 129.2 123.1

142 144 146 148 150

3.041 3.200 3.365 3.538 3.719

61.346 61.309 61.272 61.235 61.198

109.9 111.9 113.9 115.9 117.9

1012.8 1011.6 1010.4 1009.2 1008

1122.7 1123.5 1124.3 1125.1 1125.9

0.008526 0.008943 0.009375 0.009824 0.01029

117.3 111.8 106.7 101.8 97.14

152 154 156 158 160

3.907 4.102 4.306 4.519 4.741

61.160 61.121 61.082 61.043 61.003

119.9 121.9 123.9 125.9 127.9

1006.9 1005.7 1004.5 1003.3 1002.1

1126.7 1127.6 1128.4 1129.2 1130

0.01078 0.01129 0.01182 0.01236 0.01293

92.74 88.57 84.62 80.88 77.32

162 164 166 168 170

4.970 5.213 5.463 5.723 5.994

60.963 60.922 60.881 60.840 60.798

129.9 131.9 133.9 135.9 137.9

1000.9 999.7 998.4 997.2 996

1130.8 1131.5 1132.3 1133.1 1133.9

0.01352 0.01413 0.01477 0.01542 0.01611

73.97 70.77 67.73 64.83 62.08

172 174 176 178 180

6.274 6.566 6.868 7.184 7.512

60.755 60.712 60.669 60.626 60.582

139.9 141.9 143.9 145.9 147.9

994.8 993.6 992.3 991.1 989.9

1134.7 1135.5 1136.2 1137 1137.8

0.01681 0.01755 0.01831 0.01909 0.01990

59.47 56.99 54.62 52.37 50.24

182 184 186 188 190

7.852 8.204 8.569 8.948 9.341

60.538 60.493 60.448 60.403 60.357

149.9 151.9 153.9 155.9 158

988.6 987.4 986.1 984.9 983.7

1138.5 1139.3 1140.1 1140.8 1141.6

0.02074 0.02161 0.02252 0.02345 0.02441

48.21 46.27 44.41 42.65 40.96

p

82" 84 86 88 90

(J)

664

39

STEAM PRODUCTION

TABLE 149B (continued) p

d

q

r

--

l

m

v

192 194 196 198 200

9.747 10.168 10.606 11.059 11.528

60.311 60.264 60.217 60.170 60.123

160 162 164 166 168

982.4 981.2 979.9 978.7 977.5

1142.4 1143.1 1143.9 1144.7 1145.4

0.02541 0.02644 0.02750 0.02860 0.02973

39.36 37.82 36.36 34.97 33.64

202 204 206 208 210

12.013 12.514 13.034 13.569 14.124

60.075 60.026 59.977 59.928 59.879

170 172 174 176 178

976.2 975 973.7 972.5 971.3

1146.2 1147 1147.8 1148.5 1149.3

0.03090 0.03211 0.03335 0.03463 0.03595

32.36 31.14 29.99 28.88 27.82

212

14.696

59.830

180.1

970

1150.1

0.03731

26.80

p'

q

r

l

w

--

v

p

214 216 218 220

15.290 15.902 16.535 17.189

0.594 1.206 1.839 2.493

182.1 184.1 186.1 188.1

968.8 967.4 966 964.7

1150.9 1151.5 1152.1 1152.8

0.03872 0.04017 0.04166 0.04320

25.83 24.89 24 23.15

222 224 226 228 230

17.863 18.559 19.275 20.016 20.779

3.167 3.863 4.579 5.320 6.083

190.2 192.2 194.2 196.2 198.2

963.5 962.2 960.8 959.5 958.3

1153.7 1154.4 1155 1155.7 1156.5

0.04478 0.04641 0.04809 0.04982 0.05160

22.33 21.55 20.79 20.07 19.38

232 234 236 238 240

21.567 22.380 23.218 24.081 24.970

6.871 7.684 8.522 9.385 10.274

200.2 202.3 204.3 206.3 208.3

957.1 955.7 954.3 953 951.7

1157.3 1158 1158.6 1159.3 1160

0.05343 0.05531 0.05724 0.05923 0.06127

18.72 18.08 17.47 16.88 16.32

242 244 246 248 250

25.885 26.828 27.797 28.795 29.825

11.189 12.132 13.101 14.099 15.129

210.4 212.4 214.4 216.5 218.5

950.4 949.1 947.7 946.3 945

1160.8 1161.5 1162.1 1162.8 1163.5

0.06337 0.06553 0.06775 0.07003 0.07236

15.78 15.26 14.76 14.28 13.82

252 254 256 258 260

30.883 31.973 33.093 34.244 35.427

16.187 17.277 18.397 19.548 20.731

220.5 222.5 224.6 226.6 228.6

943.6 942.3 940.9 939.5 938.1

1164.1 1164.8 1165.5 1166.1 1166.7

0.07476 0.07723 0.07976 0.08237 0.08504

13.38 12.95 12.54 12.14 11.76

262 264 266 268 270

36.643 37.893 39.177 40.496 41.851

21.947 23.197 24.481 25.800 27.155

230.7 232.7 234.7 236.8 238.8

936.8 935.4 934 932.6 931.1

1167.5 1168.1 1168.7 1169.4 1169.9

0.08776 0.09057 0.09345 0.09640 0.09941

i 1.39 11.04 10.70 10.37 10.06

272 274 276 278 280

43.250 44.680 46.143 47.650 49.200

28.554 29.984 31.447 32.954 34.504

240.9 242.9 245 247.1 249.1

929.6 928.2 926.8 925.4 924.1

1170.5 1171.1 1171.8 1172.5 1173.2

0.1025 0.1057 0.1089 0.1123 0.1157

9.756 9.462 9.179 8.907 8.643

282 284 286 288 290

50.790 52.420 54.080 55.800 57.555

36.094 37.724 39.384 41.104 42.859

251.1 253.2 255.2 257.3 259.3

922.7 921.3 919.8 918.4 917

1173.8 1174.5 1175 1175.7 1176.3

0.1192 0.1228 0.1264 0.1302 0.1340

8.389 8.144 7.908 7.681 7.461

39

665

TABLES

TABLE 149B (continued)

p 292 294 296 298 300 p

59.355 61.200 63.090 65.025 67.006

p' 44.659 46.504 48.394 50.329 52.310

p'

q 261.4 263.4 265.5 267.6 269.6 q

A

r

915.5 914.1 912.7 911.2 909.7

r

-

(J)

1176.9 1177.5 1178.2 1178.8 1179.3

0.1380 0.1420 0.1462 0.1504 0.1547

A

(J)

-

v 7.247 7.040 6.841 6.650 6.464

v

74.7 84.7 94.7 104.7 114.7

60 70 80 90 100

307.33 316.05 323.92 331.17 337.89

277.2 286.2 294.4 301.8 308.9

904.3 897.7 891.6 886.1 880.6

1181.5 1183.9 1186 1187.9 1189.5

0.1714 0.1929 0.2143 0.2357 0.2570

5.836 5.184 4.666 4.244 3.892

124.7 134.7 144.7 154.7 164.7

110 120 130 140 150

344.16 350.05 355.59 360.85 365.85

315.4 321.5 327.5 333.1 338.4

'875.6 870.8 866.2 861.7 857.5

1191 1192.3 1193.7 1194.8 1195.9

0.2781 0.2993 0.3204 0.3415 0.3626

3.596 3.341 3.121 2.928 2.758

174.7 184.7 194.7 204.7 214.7

160 170 180 190 200

370.61 375.18 379.55 383.75 387.78

343.4 348.2 352.8 357.3 361.7

853.5 849.5 845.7 841.9 838.3

1196.9 1197.7 1198.5 1199.2 1200

0.3836 0.4046 0.4256 0.4467 0.4677

2.607 2.472 2.349 2.239 2.138

224.7 234.7 244.7 254.7 264.7

210 220 230 240 250

391.67 395.44 399.06 402.59 406

365.9 369.9 373.9 377.7 381.3

834.7 831.2 827.8 824.5 821.2

1200.6 1201.1 1201.7 1202.2 1202.5

0.4887 0.5098 0.5308 0.5519 0.5730

2.046 1.962 1.884 1.812 1.745

274.7 284.7 294.7 304.7 314.7

260 270 280 290 300

409.32 412.56 415.70 418.75 421.74

385 388.5 391.9 395.3 398.7

817.9 814.8 811.6 808.4 805.4

1202.9 1203.3 1203.5 1203.7 1204.1

0.5941 0.6151 0.6362 0.6575 0.6785

1.683 1.626 1.572 1.521 1.474

334.7 354.7 374.7 394.7 414.7

320 340 360 380 400

427.49 432.98 438.24 443.28 448.13

405 411 416.8 422.4 427.9

799.6 793.9 788.4 782.9 777.6

1204.6 1204.9 1205.2 1205.3 1205.5

0.7209 0.7634 0.8060 0.8489 0.8919

1.387 1.310 1.241 1.178 1.121

514.7 614.7 714.7 814.7 1,014.7

500 600 700 800 1,000

470 505.41 520.33 546.37

452.8 474.6 494.2 512.3 544.7

752.3 728.8 706.8 685.9 646.3

1205.1 1203.4 1201 1198.2 1191

1.109 1.332 1.559 1.793 2.281

0.9016 0.7509 0.6412 0.5576 0.4383

1,214.7 1,514.7 2,014.7 3,014.7 3,207.4

1,200 1,500 2,000 3,000 3,193

568.75 597.50 636.82 696.08 705.47

573.6 613 673 805.7 902.7

608.7 554 461 209.3 0

1182.3 1167 1134 1015 902.7

2.801 3.660 5.377 11.930 19.660

0.3570 0.2732 0.1860 0.0838 0.0509

For a pressure of

488.~

p'= o p.s.i.g. p'= 85 p' =:= 427 p' = 853

A= A= 1= 1=

1150 B.Th.V./lb. 1187 1205 1196

666

39

STEAM PRODUCTION

Now the possibilities of transformation and utilisation of the energy contained in this unit weight of steam increase very rapidly with the pressure. Hence the interest of high pressures: it costs hardly any more in the way of heat content to produce steam at 427 than at 85 p.s.i. (b) Table 150. This table is applicable to superheated steam, and gives the mean specific heat of the steam, at different pressures, between the temperature corresponding to saturation and various temperatures of superheat. This table is useful for calculations relating to superheaters, and permits of calculation of the following table. (c) Table 151. This table gives the total heat contained in unit weight of superheated steam, at various pressures and temperatures. (d) Table 152. This gives the specific volume of superheated steam, at various pressures and temperatures. This table is useful for calculations relating to steam piping. TABLE 150 MEAN SPECIFIC HEAT OF SUPERHEATED STEAM

(Knoblauch and Jakob) This table gives the mean specific heat c of steam between its saturation temperature t and the final temperature of superheat T (metric units). This mean specific heat permits of calculation of the total heat in the superheated steam, A: A =ë +

c(T—t)

c = mean specific heat of the steam, at constant pressure, between t° and T°C t = saturation temperature of the steam at the pressure p T = final temperature of the superheated steam ë = total heat of the saturated steam at pressure p, in kcal/kg A = total heat of the superheated steam at T° and pressure p p = absolute pressure of the steam, in kg/cm2 co = specific heat of saturated steam at pressure p p p' t co

1 0 99 0.487

2 1 120 0.501

4 3 143 0.528

6 5 158 0.555

8 7 170 0.584

10 9 179 0.613

12 11 187 0.642

14 13 194 0.670

16 15 200 0.699

18 17 206 0.729

20 19 211 0.760

25 24 223 0.848

30 29 233 0.940

T

c

c

c

c

c

c

c

c

c

c

c

c

c

120 140 160 180 200

0.483 0.480 0.478 0.476 0.475

0.496 0.491 0.488 0.486

0.521 0.515 0.509

0.544 0.534

0.576 0.561

0.590

0.623

0.660

220 240 260 280 300

0.475 0.474 0.474 0.474 0.474

0.485 0.484 0.483 0.482 0.482

0.505 0.501 0.499 0.497 0.496

0.526 0.519 0.514 0.510 0.508

0.548 0.538 0.530 0.525 0.521

0.572 0.558 0.548 0.540 0.534

0.599 0.580 0.567 0.556 0.548

0.629 0.605 0.588 0.575 0.565

0.661 0.631 0.610 0.594 0.582

0.697 0.660 0.634 0.615 0.600

0.738 0.694 0.660 0.637 0.619

0.783 0.729 0.692 0.665

0.898 0.808 0.752 0.714

320 340 360 380 400

0.475 0.476 0.477 0.478

0.482 0.482 0.483 0.483 0.484

0.495 0.494 0.494 0.494 0.494

0.505 0.504 0.504 0.503 0.503

0.517 0.515 0.514 0.512 0.511

0.530 0.527 0.524 0.522

0.543 0.538 0.535 0.533

0.558 0.552 0.548 0.545

0.572 0.565 0.560 0.556

0.589 0.580 0.574 0.568

0.606 0.596 0.587 0.580

0.645 0.630 0.617 0.607

0.685 0.565 0.647

_

39

TABLES TABLE 151 TOTAL HEAT OF SUPERHEATED STEAM

This table gives the total heat A of eqn. (548) using the same nomenclature. p' — corresponding approximate gauge pressure. A. METRIC UNITS P P T°C

200 210 220 230 240 250 260 270 280 290 300 310 320 330 340 350 360 370 380 390 400 410 420 430 440 450 460 470 480 490 500 510 520 530 540 550

9

8

677.5

683

688.4 693.5 698.7 703.8

11 10

13 12

674.9 680.8 686.3 691.6

671.8 678.2

697

702.2

684

698.6 695.2 700.6

709 714 719

724.1 729.2

707.5 712.6 717.7 722.8 727.9

711.2 716.3 721.5 726.7

734.3 739.3 744.3 749.4 754.4

733.1 738.2 743.2 748.4 753.5

731.9 737.1 742.2 747.4 752.5

759.4 764.5 769.6 774.6 779.6

758.5 763.7 768.7 773.8 778.9

757.6 762.8 7679

784.8 789.9

795

800.1 805.2 810.2 815.3 820.5 825.7 838.8

836

841.2 846.4 851.6 856.8

706

17 16

; 672.5 679.2 685.3 691.3

697

21 20

26 25

31 30

673.7 680.6 687.2 693.2

673.8 681.3 688.1

674.6 682.4

702.7 708.1 713.5 718.9 724.3

699.2 710.6 716.2 721.8

694.6 700.8 706.8 712.7 718.4

729.6 734.8

727.2 732.6 737.9 743.4 748.7

724.1 729.7 735.2 740.8 746.3

754

740

745.4 750.6

705

689.6 696.3 702.8

709 715 721

41 40

51 50

61 60

677.8 686.1 693.6 707.7 707.5

673.6 682.8 691.3 699.1

670.1 680.3 689.5

726.8 732.5 738.2 743.7

714.1 720.4 726.6 732.7 738.5

706.5 713.5 720.2 726.7

698

749.3 754.9 760.3 765.7 771.2

744.3 750.1 755.9 761.6 767.3

739.2 745.4 751.5 757.4 763.2

733.8 740.4 746.8

733

705.8 713.3 720.5 727.2

778.1

755.8 761.1 766.2 771.4 776.6

759.3 764.6 769.8 775.1

751.7 757.1 762.5 767.8 773.1

784.1 789.2 794.3 799.4 804.5

783.3 788.5 793.6 798.8 803.9

781.9 787.1 792.3 797.5 802.7

780.4 785.7 790.9 796.2 801.4

778.6 783.9 789.2 794.6 799.9

776.7 782.1 787.5 792.9 798.3

772.9 778.5 784.1 789.7 795.2

769.1 774.8 780.5 786.2

765.2 771.1 111

792

788.8

809.6 814.8

809

807.8

806.6 811.9 817.2 822.5 827.7

805.1 810.4 815.8 821.2 826.4

803.6

800.6 806.1 811.6 817.1 822.6

797.5 803.2 808.8 814.5

794,5 800.2

820

811.7 817.4

833

831.8 837.1 842.6 847.9 853.2

830.6 835.9 841.4 846.8 852.1

828.1 833.5

825.5

823

836.6 842.2 847.8

840

820

825.2 830.3 835.5 840.7

846

851.2 856.4

773

814.2 819.4 824.6 829.7

835

840.2 845.5 850.8

856

813

818.3 823.6 828.8 834.1 839.3 844.6 849.9 855.1

838.3 843.7

849

854.3

809

814.4 819.8 825.2

839

844.5 849.9

831

753

759.1

783

806

828.6 834.4 845.6

TABLE 151 (continued) B. BRITISH UNITS (Ë in B.Th.U./lb.) P (p.s.La.) 114.7 134.7 154.7 174.7 194.7 214.7

T°F (p.s.i.g.)

100 120 140 160 180 200

400 1226.0 1223.0 1219.8 1216.2 1212.6 1208.7

500 1277.0 1275.7 1273.6 1271.7 1269.6 1267.6

600

700

800

900

WOO

1327.6 1326.3 1324.8 1323.3 1321.9 1320.4

1378.0 1376.8 1375.8 1374.6 1373.6 1372.5

1428.0 1427.6 1426.8 1425.8 1425.1 1424.3

1479.6 1478.8 1478.1 1477.4 1476.7 1475.9

1531.0 1530.6 1530.1 1529.6 1528.9 1528.4

668

39

STEAM PRODUCTION TABLE 151 (continued) B. BRITISH UNITS (Ë in B.Th.U./lb.)

p' (p.s.i.g.)

P (p.sJ.a.)

T°F 400

250 300 350 400 450 500 600 800

264.7 314.7 364.7 414.7 464.7 514.7 614.7 814.7

500

600

700

800

900

1000

1262.3 1256.7 1250.9 1244.8 1238.0 1230.8 1214.5

1316.9 1313.1 1309.3 1305.5 1301.4 1297.4 1288.8 1269.8

1369.7 1367.0 1364.1 1361.4 1358.4 1355.5 1349.6 1337.3

1422.0 1419.8 1417.6 1415.4 1413.2 1411.0 1406.6 1397.5

1474.3 1472.5 1470.8 1469.1 1467.2 1465.5 1462.0 1454.9

1526.9 1525.6 1524.1 1522.7 1521.3 1520.1 1517.0 1511.3

TABLE 152A SPECIFIC VOLUME OF SUPERHEATED STEAM (METRIC UNITS)

p p' / T v

= at Dsolute pressure of the 1 steam, in kg/ci11* = approximate gauge pressure of steam, kg/cm 2 = corresponding saturation temperature at /?, °C = temperature of the superheated steam, °C = specific volume of steam at T°C and pressure /?, in m 3 /kg Sp. vol. v in m 3 /kg at a temperature 7X°C) of: p

p'

t°C

1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25 26 27 28 29 30 31 36 41 51 61

0 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25 26 27 28 29 30 35 40 50 60

99.1 119.6 132.9 142.9 151.1 158.1 164.2 169.6 174.5 179 183.2 187.1 190.7 194.1 197.4 200.4 203.3 206.1 208.8 211.4 213.8 216.2 218.5 220.7 222.9 225 227 229 230.9 232.8 234.6 243 250.6 263.9 275.4

t

150

1.725 0.902 0.617 0.471 0.382 0.321 0.278 0.245 0.219 0.198 0.181 0.166 0.154 0.144 0.134 0.126 0.119 0.113 0.107 0.102 0.0968 0.0925 0.0886 0.0849 0.0816 0.0785 0.0756 0.0729 0.0704 0.0680 0.0658 0.0566 0.0495 0.0394 0.0325

1.976 0.980 0.647 0.481

200 2.215 1.102 0.731 0.545 0.434 0.359 0.306 0.267 0.235 0.210 0.190 0.173 0.159 0.146 0.136

250 2.453 1.222 0.812 0.607 0.484 0.402 0.343 0.299 0.265 0.237 0.215 0.196 0.181 0.167 0.155 0.145 0.136 0.128 0.120 0.114 0.108 0.103 0.0978 0.0932 0.0890 0.0852 0.0817 0.0783 0.0752 0.0723 0.0696 0.0582

300 2.691 1.342 0.892 0.668 0.533 0.443 0.379 0.330 0.293 0.263 0.239 0.218 0.201 0.186 0.173 0.162 0.152 0.143 0.135 0.128 0.122 0.116 0.110 0.105 0.101 0.0968 0.0930 0.0894 0.0860 0.0829 0.0800 0.0678 0.0585 0.0454 0.0364

350

400

450

2.927 1.461 0.972 0.728 0.581 0.484 0.414 0.361 0.321 0.288 0.261 0.239 0.220 0.204 0.190 0.178 0.167 0.158 0.149 0.141 0.134 0.128 0.122 0.117 0.112 0.107 0.103 0.0994 0.0958 0.0924 0.0893 0.0761 0.0661 0.0520 0.0425

3.164 1.580 1.052 0.788 0.629 0.524 0.448 0.392 0.348 0.313 0.284 0.260 0.239 0.222 0.207 0.194 0.182 0.172 0.162 0.154 0.147 0.140 0.133 0.128 0.123 0.118 0.113 0.109 0.105 0.101 0.0979 0.0837 0.0730 0.0578 0.0476

3.400 1.698 1.131 0.847 0.677 0.564 0.483 0.422 0.375 0.337 0.306 0.280 0.258 0.240 0.223 0.209 0.197 0.186 0.176 0.167 0.159 0.151 0.144 0.138 0.133 0.127 0.123 0.118 0.114 0.110 0.106 0.0910 0.0795 0.0632 (.0523

500 3.636 1.816 1.210 0.907 0.725 0.604 0.517 0.452 0.401 0.361 0.328 0.300 0.277 0.257 0.240 0.225 0.211 0.199 0.189 0.179 0.170 0.162 0.155 0.149 0.143 0.137 0.132 0.127 0.123 0.118 0.115 0.0982 0.0859 0.0685 0.0568

39

TABLES TABLE 152B SPECIFIC VOLUME OF SUPERHEATED STEAM (BRITISH UNITS)

p p' t T v

= absolute pressure of steam, in p.s.i.a. — gauge pressure, p.s.i.g. = temperature (°F) of saturated steam at pressure p = temperature of the superheated steam, °F = specific volume of the steam, at 7°F in cu.ft./lb.

P

p'

t°F

Sp. vol. v in cu.ft./lb. at a temperature T(°F) of:

t

300

400

500

600

700

800

900

212

26.8 16.5 12.0 7.83 5.84 4.67 3.89

30.5 18.1 12.8 8.00

34.7 20.6 14.6 9.18 6.67 5.22 4.28

38.8 23.0 16.3 10.3 7.51 5.90 4.85

42.9 25.5 18.1 11.4 8.35 6.56 5.40

46.9 27.9 19.8 12.6 9.17 7.22 5.94

51.0 30.3 21.6 13.7 9.98 7.86 6.48

55.1 32.7 23.3 14.8 10.8 8.50 7.01

59.1 35.2 25.0 15.9 11.6 9.14 7.54

1000

14.7 24.7 34.7 54.7 74.7 94.7 114.7

0 10 20 40 60 80 100

239r.4 258.8 286.7 307.3 323.9 337.9

134.7 154.7 174.7 194.7 214.7

120 140 160 180 200

360.9 370.6 379.5 387.8

350

3.34 2.93 2.61 2.35 2.14

3.61 3.12 2.74 2.42 2.19

4.11 3.56 3.13 2.80 2.53

4.58 3.98 3.51 3.14 2.84

5.05 4.39 3.88 3.47 3.14

5.51 4.79 4.23 3.79 3.43

5.96 5.19 4.59 4.11 3.72

6.42 5.58 4.94 4.43 4.01

234.7 254.7 274.7 294.7 314.7

220 240 260 280 300

395.4 402.6 409.3 415.7 421.7

1.96 1.81 1.68 1.57 1.47

1.98

2.30 2.11 1.94 1.80 1.68

2.59 2.38 2.20 2.04 1.90

2.87 2.63 2.44 2.27 2.12

3.14 2.89 2.67 2.49 2.32

3.40 3.13 2.90 2.70 2.52

3.66 3.37 3.13 2.91 2.73

334.7 354.7 374.7 394.7 414.7

320 340 360 380 400

427.5

1.39 1.31 1.24 1.18 1.12

1.57 1.47 1.38 1.30 1.23

1.78 1.68 1.58 1.50 1.42

1.98 1.87 1.77 1.67 1.59

2.18 2.06 1.94 1.84 1.75

2.37 2.23 2.11 2.00 1.90

2.56 2.41 2.28 2.16 2.06

464.7 514.7 614.7 714.7 814.7

450 500 600 700 800

1.000 0.902 0.751 0.641 0.558

1.08 0.956 0.772

1.25 1.12 0.921 0.775 0.664

1.41 1.26 1.05 0.887 0.768

1.55 1.40 1.16 0.988 0.859

1.69 1.49 1.27 1.08 0.945

1.83 1.65 1.37 1.18 1.03

433

438.2 443.3 448.1 459.5

470

488.8 505.4 520.3

Formulae

The formulae (273), (274) and (275), while valuable for low pressures, are only approximate above about 300 p.s.i., and should not be used beyond 425 p.s.i. For high pressures, they may be replaced by the following formulae: Ë = 1205.5 ■ r=

(447.6 — 0 2 1980

138.8(705 — 0 0 · 3 1 5

(546) (547)

X = total heat content of saturated steam at f expressed in B.Th.U./lb. for absolute pressures above 425 p.s.i. r = latent heat of vaporisation of the steam at pressures above 425 p.s.i. t = saturation temperature of the steam, in °F.

670

STEAM PRODUCTION

39

Dryness fraction

Steam produced in boilers not equipped with superheaters is generally not dry, since it carries with it droplets of water. The dryness fraction x of a wet steam is the quantity of dry steam contained in unit weight of that steam. Steam of a dryness fraction x = 0.95 contains 95% of vapour and 5% of water entrained as droplets. Tromp (p. 268) gives as mean values in the sugar factory, when there are no superheaters: High-pressure steam Exhaust steam

x = 0.95 to 0.97 x = 0.80 to 0.90

(SAR\ K '

However, many designers assume 0.95 for the exhaust steam from engines. A calorimeter for determining dryness fraction is fully specified by Tromp (p. 268) and moreover.is readily installed; in the absence of such an apparatus, we may adopt as a first approximation: Live steam x = 0.95 \ ($AQ\ K Exhaust steam x = 0.90 j ' Entropy diagrams

The engineer who is studying plans for installations of boilers, turbines or steam engines cannot do without a steam diagram. This is a graph on which are set out the most important properties of steam. The most valuable diagrams are: (a) The Moilier diagram, which plots as abscissa the entropy of the steam and as ordinate its total heat ë. (b) The^entropy diagram, which plots as abscissa the entropy and as ordinate the temperature t. The entropy of steam is the quantity: ö

=

dß \ -ø~

(550)

J 32 1

t = steam temperature T = absolute temperature of the steam = (459.4 + f)°F Q = quantity of heat involved in the transformation under consideration. Jhis somewhat unfamiliar concept of entropy presents great advantages: (1) The diagram may be drawn at a suitable scale for giving the properties of steam under all conditions encountered in practice. (2) Isothermal transformations are represented on the entropy diagram by horizontal lines. (3) Adiabatic changes are represented by vertical lines. (4) Changes which involve no external work are represented by lines of equal total heat. It is therefore easy to trace the properties of steam before and after a given change of con­ ditions. We may consider as: (a) Changes at constant total heat: Throttling in steam engines. Expansion without external work (in a pressure reducer, for example). The variation in internal energy of the steam, corresponding to the drop in pressure, is utilised to evaporate the water contained in the steam (wet steam) or to superheat it (dry steam).

39

ENTROPY DIAGRAMS

671

On the Mollier diagram, these changes are represented by a horizontal line. (b) Isothermal changes: Evaporation at constant pressure. (c) Adiabatic changes: Compression. Expansion in a turbine or steam engine, with production of external work. This is the most important and the most interesting case in the present discussion.

Entropy

d

Fig. 363. Entropy diagram. Significance of graphs.

The external work is obtained from the heat given up by the steam. The total heat of one pound of steam the state of which is represented by the point D (Fig. 363) by reason of the choice of co-ordinates, is equal to the area OABCDd. In the same way, the total heat relative to the point D' will be equal to the area OAB'D'd. But it is not necessary to measure these areas, since ë is given by curves placed on the diagram and may be read off by interpolation. Example. Assume a steam turbine receiving steam at 327 p.s.i.a. and 662°F. This condition is represented on the diagram by the point D, intersection of the isobar 327 p.s.i.a. with the horizontal 662°. The corresponding total heat is read off from the diagram: = 1,347 B.Th.U./lb. The expansion in the turbine will be adiabatic, i.e. it will take place along a vertical line on the diagram. If the back pressure is 21.8 p.s.i.a. the final state of the steam will be represented by the point D\ at the intersection of the vertical through D and the horizontal of 21.8 p.s.i.a. or 232.5°F. It is readily seen on the diagram that this point D' corresponds to wet steam of dryness fraction x == 0.935, and of total heat V == 1,090 B.Th.U./lb. The external work accomplished per pound of steam is then: T = (A — ë') 778 ft.lb./lb. or here: 7 = (1,347 — 1,090)778 = 200,000 ft.lb./lb.

(551)

672

39

STEAM PRODUCTION

778 = mechanical equivalent of the B.Th.U. = number of foot pounds corresponding to 1 B.Th.U. In practice, it is obviously necessary to take into account the efficiency of the operation. Moreover, the change is not exactly adiabatic; instead of following a vertical line DD'

Superheated steam zone

Wet steam zone Entropy Fig. 364. Mollier diagram. Significance of graphs.

(Fig. 365), it follows a curve DD": there is a slight degradation of energy leading to a slight superheat (or to an increase in dryness fraction). We shall have several occasions to return to the entropy diagram. Fig. 366 gives a large scale diagram which will permit of preliminary solutions to problems concerned with the use of steam in the factory.

Fig. 365. Adiabatic expansion. Equivalents

Table 153 will serve to recall the equivalent figures for heat and mechanical work. TABLE 153 THERMO-MECHANICAL EQUIVALENTS

£, the work corresponding to 1 B.Th.U. = 778.3 ft.lb. ( =

427kgm)

u 1 t_ i. · 550x3,600 ~*ËËð™¾ô, ^ „ „ , ,x hence: 1 h.p.h. requires ==~ = 2,544 B.Th.U. ( = 632 kcal) 778.3 2,544 1 kWh requires = 3,413 B.Th.U. ( = 860 kcal) 0.745

0.25

1.00

1.25

Entropy B.T.U. per lb. Deg. Fa hr.

1.75

2.00

Fig. 366. Temperature-entropy chart. Reprinted with permission from JOSEPH H. KEENAN AND FREDERICK G. KEYES, Thermodynamic Properties of Steam, 1936, John Wiley & Sons, Inc.

A MOLLIER CHART FOR STEAM Modified and greatly reduced from Keenan and Keye's Thermodynamic Properties of Steam, published (1936) by John Wiley and Sons, Inc. Reproduced by permission of the publishers. Abs. press.• in.Hg.

Sat. te~.:.

I

p~:'~:.

of

101.74 0.20 34.56 I 1.0 40.23 2 0.25 126.08 44.96 I 3 141.48 0.30 0.35 152.97 49.06 I 4 52.64 5 162.24 0.40 0.45 170.06 55.87 [ 6 58.80 7 0.50 176.85 61.48 8 182.86 0.55 0.60 63.95 9 188.28 0.65 66.26 10 193.21 0.70 0.75 0.80 72.32114.696 I 212.00 0.85 74.13 16 216.32 75.84 18 222.41 0.90 0.95 77.47 ' 20 227.96 1.00 240.07 79.03 25 1.10 250.13 259.28 1.20 267.25 1.30 274.44 1.40 I 87.17140 89.51 45 1.50 281.01 1.60 1292.71 1.70 1 95.77 70 302.92 97.65 I 80 1.80 312.03 99.43 90 320.27 1.90 327.81 2.00 101.14

I

I ~g:~

i

~~

ig~:~~

i

~t~~

i ~~

m~ ~

Abs. press., lb. per SQ. in.

Sat. temp.•

lb. per sq. in.

i

120 140 160 180 200 220

Sat. temp.• of

341::I5

353.02 363.53 373.06 381.79 389.86 397.37 404.42 411.05 417.33 444.59 467.01

240

260 280 300 400 500 I 600 700 800 1000 1200 i

I

486 21

1 503.10 .

518.23 544.61 567.22 587.10 604.90 621.03 635.82

1400

1600 I 1800 I 2000

12200 2400

649.46 666.12

2600 673.94 12800 684.99 3000 609.36 , 3-206.21,05.40

lux)

:9 fo--+---I'o-+-+"-~'--~

m

ii

Q.

:::i 1250'Jo-l-lHT1~fol"-If1-£++-ftJt-....f,ol---+'-lR-+++~~

". Cli ~

~

...c:-

~'200

!.JJ

1.2

'8-

:::> t-:

:[l

1.1

1300

1.3

1.4 Entropy

1.5

Fig. 367

39

QUANTITY OF BAGASSE

673

BAGASSE Final bagasse, or simply bagasse, is the solid fibrous material which leaves the delivery opening of the last mill of the tandem, after extraction of the juice. It is the residue from the milling of cane. Physical composition

In spite of the diversity of milling plants and machines employed, the physical composition of bagasse varies between rather narrow limits. Its most important property, from the point of view of steam production, is its moisture content. Now very poor work at the mills will give a bagasse of 50% moisture while very good work will give a bagasse of 40% moisture. Certain factories in Hawaii and Formosa report moistures of the order of 38%, but these are exceptional figures. It is still very difficult, even in a modern mill, to obtain figures below 42%. The most frequent values are: w

= 42^8%

(552)

w = moisture in bagasse and generally we shall not involve any great error in adopting for practically all cases the standard value: w = 45%

(553)

In addition to water, the bagasse contains: (a) insoluble material, consisting mainly of cellulose, and comprising the fibre content of the bagasse, (b) substances in solution in the water (this water obviously originating in the juice), consist­ ing of sugar and of impurities. These dissolved substances are present in small quantities, ranging from 2 to 5%. If we designate their proportion by weight as D%, there remains of fibre: F = 100 — w — D = 47-56%

(554)

F= 50%

(555)

a fairly frequent value being: F = fibre% bagasse. Quantity of bagasse The extreme values of the mean fibre content of cane are close to: / = 10 a n d / = 16%; but it generally lies in the region of 12-14%. We obtain the quantity B of bagasse obtained from 100 parts of cane, by equating the weight of fibre entering the mills to that leaving: 100/= B · F

(556)

B= 100-^

(557)

hence:

F = fibre in bagasse.

674

STEAM PRODUCTION

39

It will be seen that the quantity of the bagasse varies between 24 and 28% by weight of cane, or approximately one quarter. We shall have available then about 560 pounds of bagasse per ton cane. Bulk density

Tromp (p. 270) gives as bulk density of bagasse: 10-15 lb./cu.ft. when stacked 5-7.5 lb./cu.ft. in a loose condition. This density depends mainly on its moisture content, thus bagasse is lighter when it contains less water. For a mean value w = 45%, we may take: d = 12.5 lb./cu.ft. for bagasse in a stack d = 7.5 lb./cu.ft. for loose bagasse. We may take 10-11 lb./cu.ft. for loose bagasse loaded on a lorry to a height of 5-7 feet. Storage of bagasse

The bulk density of bagasse makes it a very bulky material. Storage of excess bagasse from the factory presents a difficult problem for this reason. Except in very dry localities, bagasse cannot be left out in the open, since it ferments, decays and loses a large proportion of its value even for use as fuel. However, it may if necessary be stored in the open, on condition that it is placed in the form of a conical or pyramidal stack, with the top at a slope of at least 30° (Fig. 368), and that a roof is formed by means of cane leaves overlapping like roofing slates, running lengthwise in the direction of the slope, as in the thatched roofs of native cottages.

Fig. 368.

Generally it will be of advantage to store the bagasse under a roof. The angle of repose of bagasse is rather variable, but generally in the range of 45-50°. To avoid the necessity for large and expensive buildings, stored bagasse should be compress­ ed, in order to reduce its volume. Bagasse presses

There are two principal types of bagasse press: (1) Baling press. Bales of 12 x 12 X 24 in., or 18 X 24 X 24 in. are formed, and are tied with two or three wires, to prevent them from re-expanding. For this purpose a hydraulic press is employed, similar to a press for hay. The bales of fresh bagasse have a density of 25-45 lb./cu.ft., according to the pressure employed. They are then stacked in "draughtboard" fashion (Fig. 369) so as to allow air to circulate between the bales.

39

BAGASSE PRESSES

675

After two or three months, the bagasse has lost part of its moisture content, which drops from 45% to about 25%, and the density of the bales is no more than 20-30 lb./cu.ft. According to Tromp (p. 320), bagasse stacked at its normal moisture content leaving the mills, is liable to give rise to spontaneous combustion. We have never experienced or heard of accidents of this type. The baling press is the only solution possible for factories having a normal excess of bagasse. It is the only system which permits fresh bagasse, reaching the end of the bagasse conveyor at the boilers, to be stored progressively, a few tons at a time, as its production continues. Bagasse, baled and stacked in "draughtboard" fashion and sheltered from the weather, may be stored for a long time. When a long period of storage is foreseen, it is of advantage to sprinkle powdered boric acid on the stack in proportion as the stack is built up.

Fig. 369. "Draught board" stacking of baled bagasse.

(2) Briquetting press. Bagasse in its original state cannot be made to remain compressed: whatever pressure be employed, it does not retain any cohesion. If on the contrary it is dried, it is possible to obtain briquettes: (a) On condition the moisture content has been reduced below 15%. Best results are obtained with bagasse of 7-8% water. If the drying is taken further the briquettes would re-absorb moisture from the air and would expand. (b) Provided pressures are used of the order of 5,000 p.s.i., or preferably 14,000 p.s.i. The briquettes are made in the form of small plates of 8 x 4 x 1 in., which can be stacked at a bulk density of 45-55 lb./cu.ft. Tromp (p. 321) specifies a press of 70 h.p. for 5-10 tons of bagasse per hour. Briquetting is facilitated and cohesive properties of the briquettes are improved by adding a binder, molasses for example, in the ratio of one part for four of bagasse. In Cuba a pressure of 120 p.s.i. is considered sufficient, when a binder is incorporated (I.S.J., (1944) p. 237). The disadvantage of briquetting is the necessity for drying the bagasse, which requires an area of 2,200 sq.ft./ton/day, and expensive handling, when air-drying is employed. It is therefore applicable only to small quantities. The best method available at present seems to be that supplied by the firm of Pawert of Basle. The bagasse is first dried to about 10% moisture, by hot gases taken from the boiler flues, in vertical drying tubes, being conveyed by fans in contact with the hot gases. It passes then into a very powerful press where it is compressed in a cylinder: the simple friction of the walls produces a pressure in the neighbourhood of 20,000 p.s.i. The compressed bagasse leaves this cylinder in the form of a "sausage" of 2\ or 3J inches in diameter which can then be broken into any length desired, for example from 1 inch to 4 inches. It has then a specific gravity of

676

39

STEAM PRODUCTION

1.1 and its bulk density, even when these round briquettes are stored in a pile, is of the order of 30 lb./cu.ft. It is thus a fuel which is readily handled, similar to compressed pellets of fine charcoal, and very clean. Breaking up bales. When bagasse stored in bales is to be re-used, the bales are broken up by passing them through a light crusher. It has been reported (I.SJ., (1944) p. 301) that this operation may cause fibrosis of the lungs in the workers handling this operation. It is necessary therefore to provide adequate ventilation. Uses of bagasse

In addition to its use as fuel, excess bagasse may find very profitable use as; (a) Raw material for manufacture of fireproofed insulating boards, used for building purposes. (b) Raw material for fabrication of paper pulp. (c) Raw material for manufacture of various solvents utilised in industry. Chemical composition of bagasse

The chemical composition of dry bagasse varies slightly according to different authors (see Table 154). TABLE 154 CHEMICAL COMPOSITION OF BAGASSE

N. Deerr Tromp (p. 455) (I.SJ., (1938) p. 175) Carbon C Hydrogen H Oxygen O Ash å

46.5 6.5 46 1 100

44 6 48 2 100

Kelly (F.A.S., (1938) p. 61)

M.R* (I.S.J., (1939) p. 69)

48.5 6 43.3 2.2

47.5 6.1 44.4 2

100

100

Danes (I.S.J., (1947) p. 103) 47.9 6.7 45.4 100

Gregory (F.A.S., (Dec. 1944) p. 26) 49 7.4 41.8 1.8 1ÖÖ

48.1 6.1 43.3 2.5 ÜÖÖ

* M . R . = m e a n o f results o f Kersten, Prinsen Geerligs, Bolk, v. d. K r e k e and Spencer.

(The results given in the fourth column originally reported as percent on fibre, have been converted to a basis of 2% ash). Since these variations are not large, we shall use in our calculations the following mean standard composition: C= H= 0=

47 % 6.5% 44 %

a =

2.5%

(558)

100 Calorific value of bagasse

The calorific value (or C.V.), is the quantity of heat which will be released by combustion of unit weight of the fuel under consideration. We distinguish two different calorific values:

39

CALORIFIC VALUE OF BAGASSE

677

(a) The gross calorific value, or higher calorific value (G.C.V.): this is the heat liberated by the combustion of one pound of the fuel, taken at 32°F and under 29.92 in. of mercury, all the products of combustion being reduced to the same conditions. The water present in the fuel, as well as the water formed by combustion of the hydrogen entering into its composition is consequently condensed. The gross calorific value is readily measured in the laboratory with the aid of the Mahler bomb calorimeter. (b) The nett calorific value, or lower calorific value (N.C.V.) which assumes on the contrary that the water formed by combustion, as also the water of constitution of the fuel, remains in the vapour state. The gross calorific value gives a good measure of the heat theoretically available from the fuel, but since, in industrial practice, it has not yet been found practicable to reduce the temperature of the combustion products below the dew point, the nett C.V. gives a more accurate indication of the heat practically obtainable. It is the nett C.V. which therefore should be used in practice; but there is no means of determining it directly. It is necessary therefore to calculate this value. Now, in this calculation, there is a certain contradiction between the condition that the combustion gases should be reduced to 32°F and 29.92 in., and the condition that the water should not be condensed. The convention is adopted of subtracting from the gross C.V. the weight of hydrogen in one pound of fuel, multiplied by 1,080 B.Th.U. This amounts to assuming that for the gross C.V., condensation takes place in the neighbour­ hood of 50°F, in the course of cooling to 32°F, since (eqn. 274): r = 1,093 — (0.7 X 18°) = 1,080 B.Th.U. It is moreover easy to verify, from the latent heat and the specific heats of water and water vapour, that the temperature at which the condensation (or in the inverse sense, evaporation) is considered to take place, has no great influence on the total heat liberated. A temperature in the neighbourhood of 32°F is therefore adopted while taking into account, in the calculation, only vapour and not liquid water; this simplifies matters. The nett C.V. of a fuel is therefore given by the formula N.C.V. = G.C.V. — 1,080£

(559)

E — weight of water vapour present in the gases produced by combustion of one pound of the fuel, expressed in pounds. Furthermore, combustion of hydrogen takes place according to the reaction: 4

32

36

The weight of water formed is thus equal to nine times the weight of hydrogen. We have then for a dry fuel: E=9H (561) ¹ = weight of hydrogen contained in one pound of the fuel. Hence: N.C.V. = G.C.V. — 9,720//

(562)

678

39

STEAM PRODUCTION

this equation applies only to a dry fuel. For a wet fuel, it is necessary to take into account also the original water present, which is then added to the water formed by combustion. Gross calorific value of dry bagasse. In spite of considerable differences in appearance be­ tween different varieties of cane, the G.C.V. of dry bagasse is remarkably constant in all countries and for all varieties of cane. Table 155 for example gives several values selected at random. TABLE 155 G.C.V. OF DRY BAGASSE

Author Behne Hedley ? Gregory Gregory

I.S.J., (1935) I.S.J., (1936) I.S.J., (1934) F.A.S., (Dec. F.A.S., (Dec.

Queensland S. Africa Hawaii Cuba Porto Rico

G.C.V. of dry bagasse

Reference

Country

kcal/kg

p. 160 p. 349 p. 126 1944) p. 26 1944) p. 26

Mean

B.Th.U./lb.

4,542 4,585 4,622 4,691 4,594

8,177 8,253 8,320 8,444 8,270

4,607

8,293

It will scarcely involve an error of more than 2%, therefore, if we adopt as a universal value for the G.C.V. of dry bagasse: G.C.V. = 8,280 B.Th.U./lb. = 4,600 kcal/kg

(563)

Nett calorific value of dry bagasse. We have seen (eqn. 558) that dry bagasse contains 6-7% of hydrogen, and have taken the mean figures as 6.5%. Eqn. (562) then gives: N.C.V. = G.C.V. — (0.065 x 9,720) = 8,280 — 630 = 7,650 B.Th.U./lb. N.C.V. = 7,650 B.Th.U./lb. = 4,250 kcal/kg

(564)

Calorific value of wet bagasse. We now know the calorific value of dry bagasse. How are we to deduce the value for the wet bagasse with which we have to deal in practice? Theoretical calculation. We may first base our figures on the percentage composition of wet bagasse, which is given in Table 156, TABLE 156 CALORIFIC VALUE OF CONSTITUENTS OF BAGASSE

Constituent Fibre Sugar Water

% F s w

C.V. kcal/kg

B.Th.U./lb.

4,600 3,955 (~ 4,000) 0

8,280 7,120 0

Not only has water no calorific value, but on the other hand it absorbs heat in being vaporiz­ ed during the combustion (N.C.V.).

39

679

CALORIFIC VALUE OF BAGASSE

G.C.V. = 8,280-^- + 7,120- S 100 ' 100 f 100 — w s N.C.V. - 8,280 —— + 7,120 — — 630 — 100 100 100

w 1,080 ♦ ' 100

or: G.C.V. = 8 2 . 8 F H - 7 1 . 2 J

}I (B.Th.U./lb.)

N.C.V. = 82.8F H- 71.2s—4.5w—630 j

(565)

In addition to the fibre F, the water w and the sucrose s, the bagasse contains only reducing sugars and other impurities. Although these represent only small quantities, it is necessary to allow for them. Since the calorific value of glucose (6,737 B.Th.U./lb.) is slightly lower than that of sucrose, and the value for dry molasses (7,380 B.Th.U./lb.), which is considered as representing the total impurities, is slightly higher, we shall include all these soluble materials with the sucrose, and write: F= 100 — s' — w s' = sucrose + combustible non-sugars, of assumed C.V. = 7,200 B.Th.U./lb. With a residual juice purity of 80, we have: 0.80

■=

1.25s

Hence: F= 100— 1.255 — w Replacing s by s' in (565) and F by its value, we have: British units G.C.V. = 8,280 — 13.5s — 82.8H> N.C.V. = 7,650 — 13.5s — 87.3w

Metric units G.C.V. = 4,600 — 7.5* — 46 w N.C.V. = 4,250 — 7.5* — 48.5w

(566)

G.C.V. = gross calorific value of the bagasse, in B.Th.U./lb. (or kcal/kg) N.C.V. = nett calorific value of the bagasse in B.Th.U./lb. (or kcal/kg) s = sucrose % bagasse w = moisture % bagasse. These formulae are theoretical, and their application gives results differing somewhat from the true values. Another formula is sometimes employed (I.S.J., (1938) p. 78): G.C.V. = 8,280 — 14.45 —

95.4H>

(567)

G.C.V. = 8,345 — 22s — 83.45H> N.C.V. = 7,783 — 22s — 88.27H>

(568)

Hessey (6th Congr. I.S.S.C.T., p. 1054) suggests:

which have been determined experimentally, and verified fairly well in practice. Example. The N.C.V. of a bagasse of 2% sucrose and 45% moisture would be: N.C.V. = 7,783 — (22 x 2) — (88.27 x 45) = 3,767 B.Th.U./lb.

680

39

STEAM PRODUCTION

Simple calculation. The preceding formulae are interesting, but in practice a simpler method may be used; (1) G.C.V. We may use: G.C.V. = 8,280 (1 — w) (Br. units)

G.C.V. = 4,600 (1 — w) (m. units)

(569)

w = moisture content of bagasse relative to unity (w = 0.45 for example, and not w = 45% as was the case in the preceding formulae). (2) N.C.V. Like the G.C.V., the N.C.V. is applicable to the crude and wet fuel. For the wet bagasse it is necessary to take into account the heat removed, not only by the water vapour produced by combustion of the hydrogen constituent of the bagasse, but also by the water vapour originating in the moisture in bagasse. We have then (eqn. 559): N.C.V. = G.C.V. — 1,080£ (Br. units)

N.C.V. = G.C.V. — 600E (m. units)

with: E = water from hydrogen of constitution plus moisture content E = 9H{\ — w) + w Hence: N.C.V. = 8,280(1 — w) — 1,080(9 x 0,065) (1 — w) — l,080w or: m. units N . C . V . = 4,250 — 4,850H>

Br. units N . C . V . = 7,650 — 8,730w>

(570)

We may comment that, in this formula, we have already taken into account the following heat losses, from the point of view of steam production in the factory: (a) Latent heat of vaporization of the water formed by combustion of the hydrogen con­ tained in the bagasse, which is lost in the flue gases with this water vapour if the latter is not condensed. (b) Latent heat of vaporization of the water content of bagasse, which in the same way is lost with the flue gases. We have then to take into account only the following losses: (a) sensible heat lost in the flue gases (b) losses by radiation (c) losses in unburnt solids (d) loss by incomplete combustion of carbon giving CO instead of CO2. COMBUSTION OF BAGASSE Composition of air

The composition of dry air is given in Table 157. TABLE 157 COMPOSITION OF DRY AIR

Oxygen Nitrogen and inerts

% By weight

% By volume

23.15 76.85

20.84 79.16

Reactions of t h e c o m b u s t i o n

The combustible elements in bagasse are carbon and hydrogen. In burning they give:

39

681

PROPERTIES OF GASEOUS PRODUCTS OF COMBUSTION C + O2 -+ CO2

\ H 2 + O -> H2O

(571)

By weight 12 g + 32 g = 44 g

2 g + 16 g = 18 g

Or: 1 + 2.67 = 3.67

1+8=9

Properties of gaseous products of combustion

Table 158 gives the principal properties which we shall use in the study of the combustion, for the principal components of the flue gases. TABLE 158 PROPERTIES OF GASEOUS PRODUCTS OF COMBUSTION

C02 H20 N2 (atmospheric) O2 CO Air

Mol.wt.

constant constant Constant êêR p.s.u and cu.ft.llb.

44 18 28 32 28 29

0.2434 0.5944 0.3822 0.3347 0.3826 0.3697

Density 3rp

md

2gg2

.

(Ib./cu.ft.)

3rp

Specific vol. . and 2Qg2 (eu.ft.flb.)

0.1234 0.0502 0.0784 0.0892 0.0780 0.0807

8.10 19.92 12.75 11.21 12.81 12.39

(A) Combustion of dry bagasse without excess air

The composition (558) of bagasse and the proportions given by the reactions (571) allow us to calculate the quantity of oxygen necessary for combustion. (a) Oxygen. In order to burn 1 lb. of dry bagasse, we require: C 0.470 lb. x 2.67 = H 2 0.065 lb. x 8 =

1.250 lb. O2 = 14.01 cu.ft. of oxygen 0.520 lb. O2 = 5.83 cu.ft. of oxygen

or a total of

1.770 lb.

O2 = 19.84 cu.ft. of oxygen

but the bagasse already contains 0.440 lb.

O2 =

Hence the air must supply

O2 = 14.91 cu.ft. of oxygen

1.330 lb.

4.93 cu.ft. of oxygen

(b) Nitrogen. This oxygen brings with it (cf. Table 157): 1.330 x — - — = 4.420 lb. nitrogen = 56.36 cu.ft. nitrogen Hence the total weight of air required is: 5.750 lb. air = 71.27 cu.ft., all these volumes of oxygen, nitrogen or air being reckoned at 32°F and 29.92 inches. Since the composition of bagasse is not strictly constant, it is futile to retain decimals which have no real significance. We shall say then: Total weight of air necessary = 5.75 lb. Total volume of air necessary = 71.3 cu.ft. (at 32°F and 29.92 in.)

Csir\ p/z;

682

STEAM PRODUCTION

39

And the quantity of water formed is: 0.065 + 0.520 = 0.585 lb. water = 11.65 cu.ft. water vapour

(573)

(B) Combustion of w e t bagasse with excess air

We can now pass on to the general case. It is not possible in practice to burn a fuel in industrial conditions by supplying only the quantity of air theoretically necessary: combustion would be poor and incomplete. In order to obtain complete combustion, without unburnt, and so that all the carbon burns to the form of CO2, it is necessary to supply a certain excess of air. A large proportion of the available heat would be lost if portion of the carbon is allowed to burn only to the form of CO. The reaction:

c + o ->- co liberates only 4,450 B.Th.U./lb. carbon, whereas the normal combustion: C + 02->C02 releases 14,650 B.Th.U./lb. We shall write: w Moisture of bagasse relative to unity Ratio of weight of air used to weight theoretically necessary m and we shall designate by: Pa = weight and Va = volume of air used per pound of bagasse Pg = weight and Vg = volume of the gaseous products of combustion Pgs = weight and Vgs = volume of the gases (assumed dry) all these volumes being reduced to 32°F and 29.92 inches. (a) Pa. We have: Pa = 5.75(1 — w)m (b) P

g.

(574)

And in consequence:

Pg = 5.75(1 — w)m + 1 (575) since the products of combustion consist of: (1) the unit weight of fuel the combustion of which we are studying; (2) the air for combustion which serves to burn this fuel. (c) Pgs. We shall obtain this quantity by deducting from Pg the water formed by combustion of the hydrogen and the water of composition of the bagasse: Pgs = Pg — 0.585(1 — w) — w

(576)

Pgs = (1 — w) (5.75m + 0.415)

(577)

or (d) Va. We have similarly: Va= 71.3(1 — w)m

(578)

(e) Vg. To calculate Vg, we must note: (1) that we have more oxygen than that strictly necessary to burn all the carbon and hydrogen in the bagasse;

39

683

COMPOSITION OF FLUE GASES

(2) that: volume of CO2 = volume of oxygen from which it is formed, and volume of H2O = volume of oxygen from which it is formed, multiplied by 2. The combustion thus gives, per pound of dry combustible (see above), the quantities given in Table 159, for a quantity (1 — w) of dry substance. TABLE 159 VOLUME OF GASEOUS PRODUCTS OF COMBUSTION

Vg = volumes of air introduced -b volume of oxygen originating in the bagasse — volume of O2 used to form water — volume of O2 used to form CO2 + volume of water formed + volume of CO2 formed

Air + 4.93 — 5.83 — 14.01 + 11.65 + 14.01

giving as total: volume of air introduced

+ 10.75

It is also necessary to add the volume of water vapour originating in the moisture contained in the bagasse. We have finally: Vg = 71.3(1 — w)m + 10.75(1 — w) +

(579)

19.92M>

or: Vg = 71.3(1 — w)m + 9.17w + 10.75

(580)

(f) Vgs. To obtain this quantity, it is sufficient to subtract from expression (579) above the water of composition, 19.92w and the water formed by combustion, 11.65(1 — w). There remains: Vgs = 71.3(1 — w)m + 10.75(1 — w) — 11.65(1 — w) or: Vgs = 71.3(1 — w)m — 0.9(1 — w)

(581)

Comment. It must not be forgotten that all the volumes given above have been calculated at 32°F and 29.92 in. of mercury. To obtain the volumes at any temperature t, it will be necessary to apply Mariotte's law: pv = RT (where T = 459.4 + t). Since R is a coefficient, and p is constant (atmospheric pressure), we have: 459.4 + / Vt = V (582) ° 459.4 + 32 vt = volume at temperature t vo = volume at 32°F. Composition of flue gases

We know the total weight of gases: Pg = 5.75(1 — w)m + 1

(575)

The weight of the individual gases is given by: (a) Nitrogen: N 8 = 1.330 x

'

(1 — w)m

or

N2=4A2(\

— w)m

684

39

STEAM PRODUCTION

(b) Oxygen:

O2 derived from the air 1.330(1 — w)m +O2 derived from the bagasse 0.440(1 -w). —O2 used to form water —0.520(1 — w) —O2 used to form CO2 —1.250(1 — w) 0 2 = 1.33(1 —w) (/n—1)

or: (c) Water:

Water formed 0.585(1 — w) + water of constitution w

H2O = 0.585(1 — w) + w

(d) Carbon dioxide: CO2 = 0.47 x 3.67(1 — w)

COa = 1.72(1 — w)

Substituting the values of m and w, and dividing by Pg, we may readily calculate the propor­ tion by weights of each of these constituents in the flue gases. Example. If w = 0.45 and m = 1.5: Pg = (5.75 x 0.55 x 1.5) + 1 = 5.74 lb. and:

N2 4.42 x O2 1.33 x H2O (0.585 x CO2 1.72 x

0.55 x 1.5 = 3.647 1b. or 0.55 x 0.5 = 0.3661b. or 0.55)-f 0.45= 0.7721b. or 0.55 = 0.945 lb. or 5.730 lb.

63.6% 6.4% 13.5% 16.5% 100 %

The slight difference between the total and Pg is due to the ash content, as the sum of the components C + H + O of the bagasse which we have assumed (558) does not amount to one pound, while we have taken one pound in eqn. (575). To be very accurate, it would have been necessary to take 1 — å. CO2 content of flue gas. We have just seen the composition of the combustion gases by weight. Their composition by volume is equally interesting, since their CO2 content gives information on the excess air used. The quantity of CO2 formed by combustion of one pound of dry bagasse is constant, since the carbon content of the bagasse is assumed constant (C = 47% approximately). If only the theoretical quantity of air were used, the CO2 content of the flue gas would be at a maximum; if there is excess air, the quantity of CO2 will remain constant in a volume of air which becomes greater as the excess air becomes higher. The proportion of CO2 will therefore decrease as the excess air increases. Now, very simple types of apparatus are used which give a determination of the percentage of CO2 in the flue gases. We seek the relation between this percentage and the quantity of excess air. We shall assume that we are dealing with an apparatus giving the CO2 as percentage of the dry gases, i.e. that it condenses or fixes the water vapour contained in the gases before analysing them (if we use an apparatus giving the percentage of CO2 in the wet gas, this percentage will be appreciably lower). The total volume of the dry gases Vgs is given by eqn. (581). The volume of CO2 which these gases contain is obtained immediately by means of the weight of CO2 found above: Volume of C0 2 in flue gas = 1.72(1 — w) x 8.10 = 14(1 — w)

39

685

COMPOSITION OF FLUE GASES

Since the coefficient of expansion is the same for all gases, the proportion calculated from volumes at 32°F will remain the same at any temperature of the measurement. The CO2 content of the flue gas volume is therefore: 14(1 — w) V9s

y

_

14(1 -w) 71.3(1 — w)m — 0.9(1 — w)

(

'

Hence we may derive m: m=

0.196 y

+ 0.0126

(584)

Since the second term is neglible relative to the first, we may write: 0.196 m=

(585)

7

weight of air used

m =

weight of air theoretically necessary ã = CO2 content per unit volume of dry gases. Table 160 gives several values of ã for the corresponding values of m. TABLE 160 RELATION BETWEEN C O 2 CONTENT OF FLUE GASES (y) AND EXCESS AIR (m)

ã m

0.06 3.27

0.07 2.80

0.08 2.45

0.09 2.18

0.10 1.96

0.11 1.78

0.12 1.63

0.13 1.51

0.14 1.40

0.15 1.30

0.16 1.22

0.17 1.15

Optimum proportion of CO2. The theoretical maximum CO2 content of the flue gases is 19.6%, according to eqn. (585). If it is desired to obtain complete combustion, without ap­ preciable formation of C O , it is necessary to work with a minimum of excess air. Eigenhuis considers (I.SJ., (1937) p. 477), following his experience in Java and in Queens­ land, that it is possible to maintain an average of 15% CO2 without solid or gaseous unburnt. Later tests with modern boiler furnaces in Queensland have reported up to 16% CO2 without unburnt. Shillington (I.SJ., (1939) p. 259) considers that good combustion can be obtained only with a CO2 content maintained between 10 and 14%. Above 14%, the proportion of C O would become much too high. In South Africa (I.SJ., (1940) p. 349), Hayes considers that heat losses due to excess air are not serious so long as the CO2 content is maintained above 12%; but below that figure the loss in efficiency becomes considerable. Conversely, above 14%, he confirms having found the presence of C O in the gas. He considers then that good combustion corresponds to a CO2 content lying between 12 and 14%. In Java (I.SJ., (1942) p. 267), the mean excess air is 63%. In Jamaica (I.SJ., (1947) p. 102), Davies considers that the optimum excess air in practice corresponds to 50%. If in addition to these authoritative opinions we consider results of tests carried out in various sugar countries we are led to conclude that the best boiler efficiencies are obtained for:

686

39

STEAM PRODUCTION

a CO2 content of an excess air of

12-14% 40-60%

R ,. pöo;

Loss in efficiency due to CO. It is considered in South Africa (I.S.J., (1940) p. 349) that each per cent of CO in the gases of combustion corresponds to a loss of heat of 4.5%. In Cuba (I.S.J., (1944) p. 235), it is indicated with more precision: 4.36% of the calorific value of the bagasse. Calculation of combustion t e m p e r a t u r e

The combustion temperature T prevailing in the bagasse furnace is readily calculated from the fact that the heat developed in the combustion is recovered in the gases passing from the furnace to the boiler. Since the calorific value and all the heat quantities are expressed with reference to a basic temperature of 32°F, it is necessary to take into account the heat already stored in the fuel and in the combustion air, at an ambient temperature of t degrees. f 1 · cc · at + f Pa · ca · at + Nt = Ó Ã P · c · at

J 32

J 32

(587)

J 32

We have then, for 1 lb. of fuel: f = ambient temperature at which the air and the fuel arrive in the furnace T° = combustion temperature sought Pa = weight of air used/lb. fuel P = weight of each of the gaseous products of combustion/lb. fuel cc = specific heat of fuel ca = specific heat of the air c = specific heat of each of the gaseous products Ni — lower calorific value of the fuel. The nett calorific value is employed, since in practice the water vapour contained in the gases is not condensed. The expression Ó/3Ã2° Pc . at represents the sum of the heat capacities of the different gases constituting the flue gas. It is necessary to use the integral since the specific heat is a function of temperature. We may avoid integrals by using Table 161, which gives the mean specific heat of the various substances with which we have to deal, between 32°F and any temperature f or T°. With the aid of this table we can readily read off the mean specific heat between any two temperatures t° and T°. Equation (587) then becomes: But:

(1 · cc + Pa · ca)t + Nt = Ô[Ó Ñû]æ0

(588)

1 + Pa = Ó P = Pg and, since these specific heats of air and the gases do not differ greatly, we may take: 1 'Cc +Pa'Ca=

Ó PC

(589)

Finally, taking losses into account: aßoNi

ô=,+

íúú&

<59ï>

39

CALCULATION OF COMBUSTION TEMPERATURE

a = coefficient taking into account unburnt solids ßo = coefficient taking into account losses by radiation in the furnace. We may take for bagasse furnaces: a = 0.98-0.99, according to draught ß0 = 0.98-0.99. TABLE 161 COMBUSTION GASES

(Habif, pp. 23-26) Temp. °C

Total heat to heat from 0 to T° at const, press, (kcal/kg)*

Mean sp. heat between 0 and T°

Specific iheat at T°

co2

N2, CO

02

0.468 0.476 0.484 0.492 0.499 0.507

0.246 0.247 0.248 0.249 0.250 0.251

0.214 0.215 0.216 0.217 0.218 0.219

0 10 20 31 43 55

0 23 48 73 100 126

0 12 24 37 50 62

0 10 21 32 43 54

0.224 0.228 0.232 0.236 0.240

0.515 0.523 0.530 0.538 0.546

0.252 0.253 0.254 0.255 0.256

0.220 0.221 0.222 0.223 0.223

67 79 92 106 120

154 183 212 242 273

75 88 101 114 128

66 77 88 100 111

0.234 0.236 0.238 0.240 0.241

0.244 0.248 0.252 0.256 0.260

0.554 0.562 0.569 0.577 0.585

0.257 0.258 0.259 0.260 0.261

0.224 0.225 0.226 0.227 0.228

134 149 164 179 195

304 337 370 404 439

141 154 168 182 195

123 135 147 159 171

0.278 0.280 0.282 0.284 0.286

0.243 0.245 0.247 0.248 0.250

0.264 0.268 0.272 0.277 0.281

0.593 0.600 0.608 0.616 0.624

0.262 0.263 0.264 0.265 0.266

0.229 0.230 0.231 0.232 0.232

211 228 245 263 281

475 510 547 585 624

209 223 237 251 266

183 195 208 220 232

0.795 0.810 0.826 0.841 0.857

0.288 0.290 0.292 0.294 0.296

0.252 0.254 0.256 0.258 0.259

0.285 0.289 0.293 0.297 0.301

0.631 0.639 0.647 0.655 0.663

0.267 0.268 0.269 0.270 0.271

0.233 0.234 0.235 0.236 0.237

299 318 337 356 376

663 703 744 786 828

280 294 309 324 338

245 258 270 283 296

0.411 0.419 0.427 0.435 0.444

0.872 0.888 0.903 0.919 0.935

0.298 0.300 0.302 0.304 0.306

0.261 0.263 0.265 0.267 0.268

0.305 0.309 0.313 0.317 0.321

0.670 0.678 0.686 0.694 0.701

0.272 0.273 0.274 0.275 0.276

0.238 0.239 0.240 0.241 0.241

397 417 439 460 482

872 916 960 1,006 1,052

353 368 383 398 414

309 322 336 349 362

0.452 0.460 0.468 0.476 0.484

0.950 0.966 0.981 0.997 1.012

0.308 0.310 0.312 0.314 0.316

0.270 0.272 0.274 0.276 0.277

0.325 0.330 0.334 0.338 0.342

0.709 0.717 0.725 0.732 0.740

0.277 0.278 0.279 0.280 0.281

0.242 0.243 0.244 0.245 0.246

505 528 551 574 598

1,099 1,147 1,196 1,246 1,296

429 444 460 476 491

376 389 403 417 431

C02

H20

N2, CO

02

CÖ2

H20

0 50 100 150 200 250

0 199 0.207 0.215 0.224 0.232 0.240

0.468 0.483 0.499 0.515 0.530 0.546

0.246 0.248 0.250 0.252 0.254 0.256

0.214 0.216 0.218 0.220 0.222 0.223

0.199 0.203 0.207 0.211 0.215 0.219

300 350 400 450 500

0.248 0.256 0.264 0.272 0.281

0.562 0.577 0.593 0.608 0.624

0.258 0.260 0.262 0.264 0.266

0.225 0.227 0.229 0.231 0.232

550 600 650 700 750

0.289 0.297 0.305 0.313 0.321

0.639 0.655 0.670 0.686 0.701

0.268 0.270 0.272 0.274 0.276

800 850 900 950 1,000

0.330 0.338 0.346 0.354 0.362

0.717 0.732 0.748 0.764 0.779

1,050 1,100 1,150 1,200 1,250

0.370 0.378 0.387 0.395 0.403

1,300 1,350 1,400 1,450 1,500 1,550 1,600 1,650 1,700 1,750

* To convert to B.Th.U./lb., multiply by 1.8.

We may comment that: (1) the combustion temperature increases as t increases,

H20

N2,CO

02

688

39

STEAM PRODUCTION

(2) it decreases as UPc increases. In other words excess air has a marked influence on the temperature of combustion. In the same way, the moisture of the bagasse will similarly lower this temperature greatly, on account of the additional water vapour present, and all the more so since the specific heat of water vapour is nearly double that of the other gases. Further, it must still not be forgotten (cf. eqn. 570) that the moisture also reduces Nu and consequently has a double effect on the temperature of combustion. Example. We shall take again the example on p. 684 and calculate TiPc for the composition of flue gases found. We shall assume: / = 30°C \ a = 0 99 ß = 0*99 | aß°Ni Ni = 2fil0 kcalj

== 2

>°30 kcal

We obtain TtPc by determining, for each of the products for combustion, the corresponding term (T—t)Pc. The simplest method is to operate by interpolation, which is very quickly done when we already have some idea of the temperature sought. If for example it is considered that the temperature lies between 1100° and 1200°C, we calculate ÓÑ
1,100° N2 02 H20 CO2

3.647 0.366 0.772 0.945

kg kg kg kg

x x x x

294 258 703 318

T

(LPc]o

1,200°

= 1,070 kcal = 94 kcal = 540 kcal = 300 kcal

3.647 0.366 0.772 0.945

kg kg kg kg

x x x x

324 283 786 356

= 1,180 kcal = 103 kcal = 606 kcal = 336 kcal

2,004 kcal

2,225 kcal

Interpolating for 2,030 kcal, we find: T— t = 1,112°C (2,034°F), T= 1,112 + 30 = 1,142°C (2,088°F) Mean specific heat. We may dispense with Table 161 by utilizing the formulae given in Table 163. TABLE 163 MEAN SPECIFIC HEAT OF COMBUSTION GASES BETWEEN 32° (OR 100°) F AND T°F

(deduced from Table 161 from Izart and Habif) Nitrogen Oxygen Water vapour C0 2

0.246 0.214 0.468 0.199

+ + + +

0.000011T 0.000010Ã 0.000087Ã 0.000046Ã

It will then be necessary to multiply by T the mean specific heat found in order to obtain the heat necessary to raise the temperature of the gas under consideration from 32° (or 100)°F to T°.

39

689

TEMPERATURES OBTAINED IN PRACTICE

No appreciable error is involved if the mean specific heat between 0°C (32°F) and T° is used in place ofthat between 30°C and T° (86°F"and T°). Since the composition of the flue gases varies only between relatively narrow limits, we may speak of the mean specific heat of the mixed gases. We may thus take approximately: For mixed combustion gases from bagasse: Mean sp. heat = 0.27 + 0.00003Ã(7ºç °F) (591) or, somewhat less accurately: M.S.H. = 0.3 (592) Estimation of furnace temperature

The furnace temperature is measured with the aid of pyrometers or Seger cones. It may also be estimated by eye with a certain degree of accuracy, particularly if the observer has trained himself to compare his estimates with pyrometer readings. The colour of the furnace is more brilliant as the temperature increases (see Table 164). TABLE 164 COLOUR SCALE OF TEMPERATURE

Dazzling white Welding heat White Straw yellow Pale yellow Golden yellow Orange Light cherry red Cherry red Incipient cherry red Dark red

1,500 1,400 1,300 1,200 1,150 1,100 1,050 1,000 900 800 650

2,700 2,500 2,400 2,200 2,100 2,000 1,900 1,800 1,650 1,500 1,200

Temperatures obtained in practice

It is considered in Cuba (F.A.S., (December 1944) p. 27) that the average temperature of a bagasse furnace is approximately 2,000°F. Inefficient furnaces give 1,500-1,900°F. The highest temperatures reported are 2,350°F for a period of some minutes, and 2,275°F in continuous operation. The author has recorded 2,375°F in a modern Cail-Steinmuller furnace with rotary bagasse feeder, with ambient air supplied under forced draught, and with 42% moisture in bagasse. However, the average temperature varied between 2,000 and 2,200°F. With bagasse of 48% moisture, an identical furnace, without forced draught, maintains an average temperature in the vicinity of 1,800°F. TABLE 165 COMBUSTION TEMPERATURE IN BAGASSE FURNACES ( ° F )

Excess air Moisture in bagasse

50% (m = 1.5)

75% (m = 1.75)

100% (m = 2)

40% (w = 0.40) 45% (w = 0.45) 50% (w = 0.50)

2,200 2,100 2,000

2,010 1,920 1,830

1,850 1,780 1,700

690

39

STEAM PRODUCTION

Table 165 summarises approximately the influence of excess air and moisture in bagasse on furnace temperature. This table assumes an ambient temperature of 86°F and a product aßoof0.98. Heat losses in flue gases Our formula (570), giving the nett calorific value of the bagasse, already takes into account the loss of heat of the water vapour passing with the gases to the chimney. The most important of the losses which remain to be accounted for is that corresponding to the sensible heat lost in these flue gases, and we shall seek to evaluate this loss. We know the composition of the flue gases, and the specific heat of its component gases. We thus have immediately the corresponding heat loss. The mean specific heat of the flue gases between 32°F and the flue gas temperature varies only slightly because this temperature itself is limited. In a modern installation, with eco­ nomizer or air heater, it is easy to obtain flue gas temperatures below 400°F, for example 350°F: but, although it is possible to reach 270°F, there is little interest in going below 300°F, which may be considered as the lower economic limit. Conversely, it would be only a very old or inadequate installation which would allow the gases to leave at more than 570°F. We shall therefore not involve any appreciable error if we take the mean specific heat be­ tween 32°F and the flue gas temperature, as equal to the specific heat at 212°F. (We may comment that our reference temperature should be 32°F, and not the ambient air temperature, since the calorific value is calculated as from 32°F, and hence it is necessary to refer the whole heat balance to this same basic temperature.) Calculation. Starting from the weights found for the components of the flue gases (ef. p. 683), the sensible heat q carried by each of these gases will be, per pound of bagasse burnt: N2 O2 H2O CO2

q\ = 4.42(1 — w)m x 0.250/ q* = 1.33(1 — w) (m — 1) X 0.218/ qs = [0.585(1 — w) + w] 0.499/ qA = 1.72(1 — w) x 0.215/

(593)

Adding, simplifying, and approximating so as not'to conserve decimals without practical significance, we obtain the total sensible heat lost (reckoned as from 32° F): Br. units 0.50 32) (1 - w) ( 1.4m + q=(tq / w m

-0.12

0.50 m. units / ^ q= t(\ — w) \\Am + — w

= sensible heat lost in flue gases, in B.Th.U./lb. (kcal/kg) of bagasse == temperature of the flue gases in °F (°C) — moisture of bagasse relative to unity = ratio of weight of air used for combustion to weight theoretically necessary. Example. For: m = 1.5

w = 0.45

/ = 356°F we shall have:

q= (356 — 32) x 0.55il.4 X 1.5 + ^ 1 ^ — 0.12)

515 B.Th.U./lb.

or approximately 14% of N.C.V. of the bagasse (3,725 B.Th.U./lb.)

0.12)

J

(594)

39

QUANTITY OF STEAM OBTAINABLE

691

Quantity of steam obtainable We may now calculate the quantity of steam which we can obtain from unit weight of bagasse. The losses of heat in the furnace and at the boiler consist of the following: (a) Latent heat of the water formed by combustion of hydrogen in the bagasse. (b) Latent heat of the water contained in the bagasse. (c) Sensible heat of the flue gas leaving the boiler. (d) Losses in unburnt solids. (e) Losses by radiation from the furnace and especially from the boiler. (f) Losses due to bad combustion of carbon giving CO instead of CO2. Now the use of the nett calorific value (formula 570) has already taken into account losses (a) and (b). The loss (c) is given by eqn. (594). The three other losses are taken into account by means of coefficients applied to the total quantity of heat which is still available after the first three losses: a = coefficient taking into account losses by unburnt solids ß = coefficient taking into account losses due to radiation ç = coefficient taking into account losses due to incomplete combustion. The quantity of heat remaining to be transferred to the steam is therefore given by the expression: Br. units Mv = (7,650

8,730H> — q)aßy

M. units Mv = (4,250 — 4,850w — q)aßv

(595)

Mv = heat transferred to the steam per lb. (kg) of bagasse burned, in B.Th.U. (kcal) w = moisture per unit of bagasse q = sensible heat of flue gases given by eqn. (594). a (solid unburnt) is of the order 0.99. It rarely descends lower, unless a very high draught is used, carrying away to the chimney relatively large pieces of bagasse. ß (radiation) varies from 0.90 to 0.95, according to the more or less efficient lagging of the boiler. For a well lagged boiler a figure of 0.95 may be taken. ç (imperfect combustion) may vary from 0.99 down to 0.80. This coefficient will be better according as: (1) the bagasse has a lower moisture, v (2) the excess air is lower, (3) the furnace temperature is higher (this condition moreover is a consequence of the two preceding ones). This coefficient falls rapidly when the moisture in bagasse exceeds 50%, or similarly when the furnace temperature falls below 1,650°F. In a good modern furnace, it easily exceeds 0.90. For a well conducted combustion, we may use a figure of 0.95-0.97. Overall efficiency. The ratio: Mv Heat units transferred to the steam ñ* = N = (596) G.C.V. of the bagasse s is called the efficiency of the boiler. This efficiency generally varies between 50 and 65%. Tromp (I.S.J., (1940) p. 90) gives the figures shown in Table 166.

692

STEAM PRODUCTION

39

TABLE 166 BOILER EFFICIENCIES

Type of boiler Elephant boiler Water-tube Water-tube

Type of grate

Overall efficiency

Step grate Horseshoe Ward furnace

50% 60% 66%

Tromp reports that this value of 66% is the highest which he had encountered. It corresponds to a value of the coefficient ç of about 0.99. In Queensland (F.A.S., (1931) p. 540), tests reported by Behne gave the figures given in Table 167. TABLE 167 BOILER EFFICIENCIES

Type of boiler

lb. steam/lb. bagasse

Overall efficiency

2.3 2.4 1.8

53.1 56.2 42.2

Thompson (1931) B. &W. (1931) Semi-tubular (1931)

/o

More recent tests in Queensland, reported by Jenkins (Tech. Commun. Sugar Expt. Stations, (1938) No. 1) give figures of 2.9 lb. steam/lb. bagasse and overall efficiency of 68% on G.C.V. for a boiler with large combustion chamber and large air heater. It is considered in Cuba (I.S.J., (1944) p. 235) that the overall efficiency never exceeds 61.3% and attains this value only in the best installations. Weight of steam per unit weight of bagasse. Table 168 gives figures for the total heat required to produce unit weight of steam, for different steam conditions obtaining in the factory. TABLE 168 HEAT REQUIRED TO PRODUCE UNIT WEIGHT OF STEAM

(Temperature of feed water 194°F) Steam pressure 2

kg/cm 6 8 10 10 15 20 25

p.s.i. 85 114 142 142 213 284 355

Steam temperature °C

225 300 325 350

Saturated Saturated Saturated

Heat required

°F

kcal/kg

B.Th.U./lb.

437 572 617 662

569 572 574 599 635 645 656

1,025 1,030 1,033 1,078 1,143 1,163 1,181

We have assumed in all cases a feed water temperature of 194°F, which is a common tempera­ ture, within a few degrees, at the feed water tank, before the water is pumped to the boiler or to the economizer. The total quantity of heat from the bagasse which is effectively utilized and is recovered in the steam depends on four principal factors: w m ç t (temperature of flue gases)

39

693

QUANTITY OF STEAM OBTAINABLE

Adopting the following figures as average or readily obtainable values: w= ma = ß= ç =

0.45 1.5 0.99 0.95 0.95

we shall have for Mv and for the weight of vapour obtainable per unit weight of bagasse the corresponding values given in Table 169. TABLE 169 HEAT TRANSMITTED TO STEAM PER POUND OF BAGASSE, AND WEIGHT OF STEAM PRODUCED PER UNIT WEIGHT OF BAGASSE, FOR DIFFERENT TEMPERATURES t OF FLUE GASES

(w = 0.45; m = 1.5; a = 0.99; â = 0.95; ç = 0.95) t, °F Mv, B.Th.U./lb. Steam produced 85 p.s.i. satd. 142 p.s.i. satd. 142 p.s.i. 437°F 213 p.s.i. 572°F 284 p.s.i. 617°F 355 p.s.i. 662°F

302 2,941

347 2,876

392 2,813

482 2,686

572 2,558

2.87 2.85 2.73 2.57 2.53 2.49

2.81 2.78 2.67 2.52 2.48 2.44

2.75 2.72 2.61 2.46 2.42 2.38

2.62 2.60 2.49 2.35 2.31 2.28

2.50 2.48 2.37 2.24 2.20 2.17

To supplement this table, which corresponds to good normal conditions, we give in Table 170 the quantities of steam obtained per unit weight of bagasse in much less favourable conditions: w= m= a = â= ç =

0.50 2 0.99 0.95 0.90

TABLE 170 HEAT TRANSMITTED TO STEAM PER POUND OF BAGASSE, A N D WEIGHT OF STEAM PRODUCED PER UNIT WEIGHT OF BAGASSE, FOR DIFFERENT TEMPERATURES / OF FLUE GASES

(w = 0.50; m = 2; a = 0.99; â = 0.95; ç = 0.90) t, °F Mv, B.Th.U./lb. Steam produced 85 p.s.i. satd. 142 p.s.i. satd. 142 p.s.i. 437°F 213 p.s.i. 572°F 284 p.s.i. 617°F 355 p.s.i. 662°F

302 2,360 2.30 2.28 -2.19 2.06 2.03 2

347 2,290

392 2,219

482 2,079

572 1,940

2.24 2.22 2.12 2 1.97 1.94

2.17 2.15 2.06 1.94 1.91 1.88

2.03 2.01 1.93 1.82 1.79 1.76

1.89 1.88 1.80 1.70 1.67 1.64

In general, it is seen that the quantity of steam produced per unit weight of bagasse varies from: 1.80 to 2.75 lb./lb. bagasse

(597)

Figures ranging from 2.25 to 2.50 lb./lb. are often obtained. According to the fibre in cane and the conditions of operation of the boiler station, we thus obtain 1,000-1,700 lb. steam/t.c.

694

STEAM PRODUCTION

39

FUELS OTHER THAN BAGASSE

Owing to insufficient fibre content in cane, excess moisture in the bagasse, a low efficiency at the boiler station, or wastage of steam, certain factories have not sufficient bagasse to produce all the steam which they require. They have therefore to use some supplementary fuel: fire­ wood, coal, fuel oil, etc. Wood

The N.C.V. of cellulose is fairly constant at about 7,650 B.Th.U./lb. However, the water con­ tent of different woods varies greatly, according to the species (and even the variety) and the time which has elapsed since it was cut. According to Izart (p. 824), a very dry wood contains 20% of water; a dry wood 30% and a green wood 40-50% of water. Firewoods generally used have between 30 and 40% water. Eqn. (570) may be applied to firewood without great error. Certain woods are very superior to others for fuel purposes. The tree "Casuarina equisetifolia" would have, at a given age, 5-10% less water than most other woods. We may take approximately: Ordinary wood moderately dry Casuarina moderately dry

N.C.V. = 4,500 B.Th.U./lb. N.C.V. = 5,850 B.Th.U./lb.

Coal and fuel oil

The calorific value of these two fuels is: Coal N.C.V. = 12,600-14,400 B.Th.U./lb. Fuel oil N.C.V. = 19,250 B.Th.U./lb. the difference between N.C.V. and G.C.V. being moreover rather small, except for coals having a high hydrogen content. Fuel oil presents the advantage that it can be burnt in bagasse furnaces, by means of special burners. Molasses

The calorific value of molasses per pound of dry substance, is: (N.C.V.) = 6,000-7,000 B.Th.U./lb., according to Noel Deerr (p. 471) (G.C.V.) - 6,840 B.Th.U./lb., (Prinsen Geerligs, p. 328) (G.C.V.) = 7,650 B.Th.U./lb., (Naus Bey, I.S.J., (1938) p. 326). The latter value would correspond to a N.C.V. of 7,020 B.Th.U./lb. We may reckon ap­ proximately, for an average molasses as fired: N.C.V. = 67 B — 900 B.Th.U./lb.

(598)

where B = dry substance in molasses. Provided it is heated and diluted (the latter practice, however, being most unfortunate from the point of view of fuel value), molasses may be burnt in burners similar to those used for fuel oil. It has the disadvantage of giving such quantities of ashes that their removal sometimes poses a problem difficult to solve.

39

695

TYPES OF FURNACE FURNACES

Bagasse feed

Older bagasse furnaces were provided with a rectangular hopper, a simple opening through the arch of the furnace. Through this wide open passage, air was drawn in with the bagasse. It encountered practically no resistance, unlike the combustion air, which was forced to pass through the grate and traverse the bed of bagasse. This air, passing through the top opening, entered the furnace directly, contributed practically nothing to the combustion, and served only to increase greatly the excess air. To avoid this drawback, bagasse furnaces are fitted with a hopper and a rotary bagasse feeder (Fig. 370). By preventing the unrestricted access of free air, this simple accessory is valuable and indispensable.

Fig. 370. Rotary feeder for bagasse furnace (Fives- Lille).

Fig. 371. Step grate.

Types of furnace

There are four principal types of bagasse furnaces: (a) the step-grate furnace, (b) the Cook or horseshoe furnace, (c) the Ward furnace, (d) the spreader-stoker furnace. (a) Step-grate furnace. (Fig. 371). This is the classic type of furnace and the most widely used.

696

STEAM PRODUCTION

Fig. 372. Horseshoe furnace (Cail).

39

39

TYPES OF FURNACE

697

The grate consists of small plates of cast iron, arranged in steps. Its inclination to the horizontal should be 52°, the value recommended by the Java experi­ ment station, and by Tromp (p. 275), and adopted by most manufacturing firms. The grate consists of three parts: (1) The upper part, or dead plate, without steps or openings for passage of air, and on which the bagasse is dried before passing on to the grate proper. (2) The grate proper, corresponding to the steps or grate bars. Certain manufacturers progressively increase the space between bars, so as to proportion the quantity of air passing through the bagasse to the degree of combustion required. Generally the bagasse ignites at the first bars of the grate, and burns from the first bars along the whole length of the grille proper. (3) The portion of slight slope, or ash grate, at the lower end of the grate, on which combus­ tion of the bagasse is completed, leaving ashes, which fall between the bars of the grate, into the ash-pit. Certain manufacturers make this ash grate horizontal, but it is of advantage to give it a slight slope. Tromp (p. 275) recommends 20°. French firms generally adopt 15°. We give (Fig. 371) an example of dimensions (in metric units) for a step grate furnace.

Fig. 373. Conical pile of bagasse.

(b) Cook furnace. (Fig. 372). This type of furnace is also described as the horseshoe furnace. It is more recent, but tends to be more widely adopted at the expense of the step grate. The bagasse falls directly from the feed chute into the furnace, which takes the form of a tank in the shape of a horseshoe. The bagasse forms a conical heap, in the form of a sugar loaf (Fig. 373). By means of tuyeres placed all around the horseshoe, and more numerous near the bottom, air, which may be cold but is preferably preheated, is blown into the pile of bagasse, which burns away and collapses on itself. The ashes accumulate at the bottom of the furnace, the height of which should be designed accordingly. The horseshoe furnace requires slightly more height than does the step-grate furnace, but has the advantage of dispensing with the grate bars and bearer bars. It permits of very high combustion rates and gives excellent results from the point of view of efficiency.

698

STEAM PRODUCTION

39

(c) Ward furnace. (Fig. 374). This is a recent design, which first appeared in the American zone and is becoming widely adopted. The hearth and the furnace are very similar to those of the Cook furnace. The hearth consists of a cast iron plate with provision for circulation of a small proportion of the air, surmounted by an oblong furnace carrying three rows of tuyeres. The bagasse burns in the middle of this in the form of a conical heap.

l·////////□////,

Fig. 374. Ward furnace (straight-tube boiler) (Babcock & Wilcox).

The original feature of this furnace lies in its upper portion, and its advantage is its simplicity. The disadvantage of bagasse furnaces in general lies in the necessity to install a separate furnace in such a way that the heat of the furnace is concentrated on the burning bagasse, while the hot gases are made to follow a circuitous path before reaching the relatively cold heating surface of the boiler, which is placed out of "sight" of the furnace. With the Ward design, the furnace is placed squarely below the heating surface; and so that combustion will not suffer in consequence, a small inclined arch is interposed over the greater part of the vertical space immediately above the furnace, so as to reflect heat on the burning bagasse while screen­ ing the furnace from the cold tubes. In boilers with straight tubes of Babcock or Steinmuller types, two such arches sloping in opposite directions are provided (Fig. 374) with a space between them about 20 inches wide across the width of the furnace. A small part of the heating surface is thus in sight of the fuel bed. With bent-tube boilers, of the Stirling type, the first pass for the gases is placed ahead of the boiler itself (Fig. 375); since the heating surface is not in sight of the furnace, a single

39

TYPES OF FURNACE

699

small arch is sufficient. The space left between it and the opposite wall is then of the order of 40 inches. This restriction in the gas path, especially in the former case, obviously causes a high gas velocity at this point, and an appreciable draught loss. In spite of the simplification of the gas circuit, the Ward furnace requires a higher draught than other types of furnace. Like the Cook furnace, it is particularly well suited for use with an air-heater. A height of about 13 ft. is allowed between the hearth and the arch. The Ward furnace slightly increases the height required, but affords a great saving in the total floor space necessary for furnace and boiler, and a substantial saving in refractories. It has given excellent results and high efficiencies.

Fig. 375. Ward furnace (bent-tube boiler) (Stirling).

(d) Spreader-stoker furnace. This is the most recent type of furnace. It has no enclosing wall as has the Cook or Ward furnace, and consists simply of the space situated between the boiler tubes and a special flat grate. The latter may be: (a) fixed; (b) of rocking type with removal of ashes by hand; (c) mechani­ cal, with continuous discharge of ashes. This third type is to be recommended for capacities of 35 tons of steam per hour or higher (I.S.J., (August 1954) p. 222). It is the method of feeding the bagasse which constitutes the most original feature of the spreader-stoker furnace. It is effected as in a Ward furnace, by means of a steeply sloping chute; but at the moment this chute discharges the bagasse into the furnace, air under pressure,

700

STEAM PRODUCTION

39

supplied by a pipe located just behind the chute, is blown in through a longitudinal slot running the whole length of the bottom end of the chute, and throws the bagasse violently into the furnace. The bagasse is thus dried and burnt as it falls into the furnace, and the largest pieces complete their combustion on the grate. Whatever type of grate is used, the air spaces through the grate are generally made 3-5% of the grate area, instead of 25-40% for step grates. The greater part of the combustion air is supplied by the air under pressure which serves to throw the bagasse into the furnace. It is considered that the spreader-stoker furnace permits of reducing the normal excess air to 30% (instead of 40-50%) and consequently of improving the efficiency substantially. Moreover, this type of furnace permits a very superior combustion rate to that of other types. For example, an evaporation rate of 7-8 lb./h/sq.ft. of heating surface may be obtained compared with 5-6 lb./h/sq.ft. for furnaces of Cook or Ward type. Of course, the dimensions of the combustion chamber must be planned to suit the type of boiler chosen; it will be advisa­ ble to keep to figures of about 25,000 B.Th.U./cu.ft./h, and not to go above 40,000 B.Th.U./ cu.ft./h. Finally, the spreader-stoker makes ash removal easy, is easy to clean, and, having no arches nor separate furnace, is economical in brickwork. Combustion chamber

The furnace is the space included between the grate (or the hearth, for a hearth type of furnace), the side walls of the furnace, the arch, the front wall and the bridge wall. The combustion chamber includes the furnace plus the free space traversed by the gases between leaving the furnace and reaching the boiler tubes. The ashpits are not included, if such are provided. Combustion chamber volume. The volume of the combustion chamber should be proportioned to the volume of gases necessary for combustion. This volume is therefore generally fixed in relation to the quantity of heat liberated per hour by the fuel used (Nf). Since there exists a certain ratio between the heating surface of the boiler and the quantity of steam which it can produce (cf. p. 705), the combustion chamber volume may also be related to the heating surface of the boiler. Noel Deerr (p. 469) recommends 10-30 cu.ft. of combustion chamber volume per 100 sq.ft. of heating surface, and this, according to his figures, would correspond to 10,000-30,000 B.Th.U./h/cu.ft. Tromp (p. 295) suggests 7-33 cu.ft./100 sq.ft. of heating surface, the highest figures corre­ sponding to the highest rates of combustion. However, on account of the evaporation rates which he indicates, this would correspond to about 30,000-70,000 B.Th.U./h/cu.ft. In Cuba, figures of 13-16 cu.ft./100 sq.ft. of heating surface are used; in Hawaii, about 25 cu.ft. These values are not applicable to furnaces of the Ward type. For spreader-stoker furnaces of Riley type, Miller (F.A.S., (April 1954) p. 46) gives a maxi­ mum of 40,000 B.Th.U./cu.ft./h, but recommends keeping in the region of 25,000 B.Th.U./ cu.ft./h as an optimum value, in order to improve the efficiency and decrease formation of fly ash. French manufacturing firms design for 23,000-28,000 B.Th.U./cu.ft. of combustion chamber volume/hour. Relating this to the heating surface of the boiler, and expressing combustion volumes in cu.ft./100 sq.ft. of heating surface, this corresponds to approximately:

39

701

DIMENSIONS OF THE FURNACE

water-tube boilers v = 26-40 cu.ft./lOO sq.ft.h.s. semi-tubular boilers v = 20-32 cu.ft./100 sq.ft.h.s.

(599)

To our knowledge the best results are obtained with 28,000 B.Th.U./cu.ft./h and 26 cu.ft./ 100 sq.ft. h.s., with water-tube boilers, but the combustion chamber volume seems to us to be a very secondary factor. The volume V of the combustion chamber is then given by:

„

B-Nt

(600)

28,000 V = combustion chamber volume, in cu.ft. B = weight of bagasse burned, in lb./h Nt = N.C.V. of the bagasse, in B.Th.U./lb. Dimensions of the furnace

All the furnace dimensions are fixed by the necessity to observe the following conditions: (a) Length of flame. The length of passage for the burning gases, between the grate and boiler tubes, should be at least 16 ft., and preferabley 23-26 ft. It should not exceed 30 ft. Below 23 ft. and particularly below 16 ft., the gases would not be completely burnt on reaching the cold water tubes, and the sudden cooling caused by their passage between the tubes would to a great extent arrest the combustion, thus increasing the proportion of CO, and decreasing the efficiency. Furthermore, below 16 ft., the ash entrained with the gases will not be completely burned, and will tend to adhere to the tubes, thus becoming harmful and dangerous (see Fig. 376).

Fig. 376. Length of flames. On the other hand, if the length of path for the gases is unnecessarily increased, there will be increased losses by radiation and by air leakage, as well as an increase in the space required and in the cost of the installation. (b) Widtfrof the boiler. The various types of water-tube boilers generally have a given heating surface per unit width of furnace. For example: Cail-Steinmuller boilers with headers Fives-Stirling boilers with 3 drums

360 sq.ft./ft. width 443 sq.ft./ft. width

To avoid the drawbacks of a complicated shape, the combustion chamber should have the

702

STEAM PRODUCTION

39

same interior width as the boiler, and the total width of the furnace or furnaces should be at most equal to the latter. (c) Volume of combustion chamber. Finally, the total volume of the combustion chamber should observe the conditions given by eqn. (599). It will be seen that the length, width and volume of the combustion chamber must conform to certain conditions, leaving little liberty to the designer of the furnace. The least imperative condition and the most elastic of the three is however that of the volume, which can, without great inconvenience, differ appreciably from the values given by eqn. (599). Secondary air. In all modern furnaces, there is provided, in addition to the normal air or "primary air", entering directly through the grate or by the tuyeres of the hearth furnace, a complementary air supply for "secondary air", behind the bridge wall, and consequently after the furnace proper. This air is introduced by a small duct built into the bridge wall (Figs. 378 and 385). The introduction of this supplementary air for combustion has the object of ensuring com­ plete combustion by changing to CO2 the CO which may remain after combustion in the furnace. Secondary air is generally made 5-15% of the total air supplied, averaging 10%. There is no advantage in exceeding this amount; if combustion is good in the furnace, i.e. if the com­ bustion temperature is high, it forms very little CO, and there would be risk of causing a useless increase in excess air. THE GRATE Grate area

The grate area is an important figure. At first sight, its value would seem to be immaterial, since large variations in grate area are encountered. Actually, a good furnace efficiency is obtainable only if the grate area conforms to certain conditions. Unfortunately, the method of reckoning grate area is not clearly defined. Many authors quote figures without specifying whether they refer to the grate proper, i.e. to the actual step grate, or whether they include the dead plate and the ash grate. The most important part and by far the most active portion of the grate is obviously that formed by the grate bars. However, the dead plate plays a part also: if it is too short, the bagasse will arrive at the point where it commences to burn with too high a moisture content, and its combustion will be less rapid and less efficient. However, it could not be argued that one square foot of dead plate was of equal importance to a square foot of the grate proper. It is the same with the ash grate: if it is insufficient, unburnt fuel will fall into the ashpit, but the importance of this disadvantage is much less than the importance of an adequate surface of the active part of the grate. In short, an additional length of grate bars could, if it were placed towards the top, replace the dead plate and, if it were obtained by an extension towards the lower end, could replace the ash grate; whereas neither an extension of the dead plate nor an extension of the ash grate could compensate for inadequate area of the grate proper. The role of the dead plate is in fact somewhat relative: the true drying effect is obtained almost as soon by the passage of air between the first grate bars as by radiation from the

39

WEIGHT OF BAGASSE BURNED

703

furnace, which acts only on the upper surface of the bagasse layer, and which, moreover, would act equally well if the dead plate were replaced by grate bars. The true reason for the presence of the dead plate is mainly to avoid entry of unwanted air between grate bars on which no combustion takes place; but combustion cannot commence except on the grate proper, and after one or two grate bars. With regard to the ash grate, if it is below a certain minimum area, its area would be im­ portant. However, with a normal ash grate, it is of little importance whether it be 8 in. shorter or 18 in. longer: the combustion would be very little influenced in consequence. If it is desired to define grate area precisely, the true solution would consist in reckoning: (a) the surface of the grate proper with a coefficient 1 (b) the surface of the dead plate and that of the ash grate multiplied by a coefficient to be determined in each case, which could be of the order of 0.3-0.5 for example. However, such a system would be very complicated. On the understanding that a certain minimum is required for the dead plate and for the ash grate, we are of the opinion that it is simpler and more logical to express the grate area by the area of the grate proper only, rather than by the total area: dead plate + grate proper + ashgrate. Proportions between different parts of the grate

The relative proportions of the dead plate, the grate proper and the ash grate are very variable. A good average would correspond to: dead plate = 0.4 -0.5 of the length of the grate proper ash grate = 0.25-0.35 of the length of the grate proper

(6o\\ ^ '

Weight of bagasse burned per unit grate area and per hour

The rate of working of a furnace grate is expressed by the quantity of bagasse burnt on this grate per square foot of grate area. We shall designate this rate by: ί=

weight of bagasse burnt per hour on the grate area of the grate proper (bars)

(602)

This "combustion rate" is interesting, since, to obtain sufficient combustion, it is necessary to proportion the grate area to the quantity of bagasse to be burnt per hour. If the draught were constant, the optimum value of the weight of bagasse burnt per square foot of grate area per hour would be easily fixed. However, this value increases rapidly with the rate of operation adopted, whereas the efficiency of the combustion varies little with the rate. In other words, the graph of combustion efficiency as a function of rate of working is very flat, and we may obtain a very good combustion: (a) at low draught and low combustion rate (b) at high draught and high combustion rate. According to Habif (p. 77), the optimum combustion rate would be obtained when burning 51 pounds of green bagasse per square foot of grate and per hour, including in grate area the dead plate and the ash grate. This would correspond to approximately: 87 lb./sq.ft./h for grate area proper Noel Deerr (p. 469) suggests approximately 100 lb./sq.ft./h Tromp (p. 278) 250-300 lb./sq.ft./h Shillington (I.S.J., (1939) p. 260): 300 lb./sq.ft./h

704

39

STEAM PRODUCTION

The furnace combustion rates which in our experience have given the best results, range from 125-185 lb./sq.ft./h, according to the more or less forced operation; but these furnaces could have carried an appreciably higher rate of feeding. We would suggest as optimum rate for step grates, the figures given in Table 171, on the understanding that the grate area given is only that of the grate proper. TABLE 171 OPTIMAL COMBUSTION RATES PER UNIT GRATE AREA FOR BAGASSE

Low Moderate High

kg/m2/h

Ib./sq.ft.lh

500- 700 700- 800 800-1,000

100-140 140-160 160-200

Evaporation rate

The "evaporation rate" or "rating" of a boiler is the weight of dry steam which it produces per unit heating surface per hour. We shall designate it by the Greek letter ô. Standard evaporation rate. We have already seen (Table 168) that the quantity of heat units required for production of unit weight of steam varies according to the temperature of the feed water for the boilers, and according to the pressure and also the superheat of the steam. To render evaporation rates comparable with one another the British practice is always to relate the quantity of steam produced to that which would be produced if the feed water were at 212°F and if evaporation were carried out at atmospheric pressure. In other words, the quantity of heat transferred to one pound of the steam produced is taken, divided by 970 (latent heat at 212°F) and the weight of steam generated per square foot of heating surface is then multiplied by the coefficient so obtained. The standard conditions in French practice on the other hand are: feed water at 0°C (32°F), steam at 100°C (212°F). In other words, French engineers replace the divisor 970 by the figure 1,150. This value has the advantage of approaching more closely to the order of magnitude of the quantity of heat supplied per pound of steam in modern industrial practice (cf. Table 168). Example. Consider a boiler receiving feed water at 194°F and producing 5 lb. steam/sq.ft. heating surface/h at 356 p.s.i., superheated to 662°F. Each pound of steam has thus received 1,180 B.Th.U. (cf. Table 168). The standard evapo­ ration rate will then be: (a) English rating: T'Q = 5 x -^— = 6.08 lb./sq.ft./h (b) French rating: TQ = 5 x ñô^=

5.13 lb./sq.ft./h

We may comment that the English rating, to be rendered comparable with the French rating, should be multiplied by: 970 (603) é 7 5 ä = 0.8435 and figures according to the French rating, to convert to English rating, should be multiplied by the reciprocal of this or 1.186.

39

705

RATIO OF HEATING SURFACE TO GRATE AREA

If the steam produced is wet, it will obviously be necessary to take into account the lower quantity of heat which has been supplied to each pound of steam. In the following discussion, the evaporation rates indicated will be either standard English ratings (from and at 212°F, ôï') or actual practical ratings obtained (ô). Average evaporation rates. Natal reports, for 1933, the following averages: Semi-tubular boilers T'Q = 2.9 lb./sq.ft./h Water-tube boilers ô^ = 3-5 lb./sq.ft./h Tromp (p. 295) gives: 3-4 lb./sq.ft./h, in general, and 7.5 lb./sq.ft./h for the best values recorded in modern installations. We shall take the figures given in Table 172. TABLE 172 MEAN EVAPORATION RATES (lb./sq.ft./h)

Normal Semi-tubular boilers Water-tube boilers

3 4.4

Forced 3.6 6

Any water-tube boiler can evaporate 4.4 lb./sq.ft./h in normal operation, but boilers with two circulations, of the Naeyer type for example, cannot be forced to the same ratings as boilers with three circulations. They exceed 5-5.2 lb./sq.ft./h only with difficulty. With very clean heating surfaces, with air heaters and forced draught below the grate, modern boilers with several circulations can achieve or exceed 7.3 lb./sq.ft./h. With economizer and air-heater of generous dimensions, a rating of 10 lb./sq.ft./h may even be attained. When the draught is altered, it may be estimated approximately that the evaporation rate will vary proportional to the square root of the draught expressed in inches of water. Ratio of heating surface to grate area

We shall designate this ratio by ó

heating surface of the boiler

=

^ð:by grate r-c— area s ofr-^: grate occupied bars

(604)

Let: P = weight of bagasse burned per hour Q = weight of steam generated per hour We shall have: PMv = 970Q

(605)

Mv = fraction of the C.V. of the bagasse which is recovered in the steam, expressed in But: P= and: Whence:

ß=<

ß-s 'S

_ S _ Mvß ° ~ T ~ 970ô' o

706

39

STEAM PRODUCTION

(a) Semi-tubular boilers. It is not desirable to seek a high evaporation rate with semitubular boilers. Moreover, their thermal efficiency is mediocre only. We may therefore take the following figures: M* =£2,340 B.Th.U./lb. ) ß = 100-140 lb./sq.ft./h whence: ó = 80-100 ô'0 = 3-3.6 lb./sq.ft./h '

(609)

(b) Water-tube boilers. We may take here: Mv = 2,700 B.Th.U./lb. ) â = 140-160 lb./sq.ft./h whence: ó = 75-100 ' ô'ï = 4.4-6 lb./sq.ft./h

(610)

In both cases then, we see that the values of a, corresponding to the most economical rates of combustion, are in the range 75-100. However, there is no need to attribute to a an exaggerated importance. The important and interesting values are those of ô
Extreme values

Average

Mauritius South Africa Egypt (Habif, p. 79)

48-146 4 0 - 90 52-113

88 70 66

Noel Deerr (p. 469) Tromp (p. 294)

70-120 50- 90

100

The figures given by Tromp however are not consistent with the combustion rates which he indicates (p. 278: /S = 250-300 lb./sq.ft./h). HORSESHOE FURNACE

This type of furnace offers little essential difference from the step-grate furnace. The principal points to remember are the following:

39

WEIGHT OF BAGASSE BURNED PER UNIT AREA

707

Weight of bagasse burned per unit grate area

In taking as area of the furnace the plane surface at the bottom, comprising the interior of the horseshoe, this furnace permits a combustion rate appreciably higher than the step grate. Tromp (p. 278) reports 450-650 lb. bagasse/sq.ft./h. French manufacturing firms prefer to base their designs on 270-360 lb./sq.ft./h, or an average of 300 lb./sq.ft./h, and consider rates of 400-450 lb./sq.ft./h as maxima. As will be seen from the two groups of limits indicated, this rating may be modified, by altering the draught or the arrangement of tuyeres, to a greater extent than with an ordinary grate. It is thus a more flexible design of furnace. Efficiency. This is also one of the types of furnace in which the highest temperatures and efficiencies have been recorded. It offers moreover a further advantage, which is by no means negligible, in that it gives ashes which are very rich in potash, and nicely pulverised. Height. Since it has no ashpit, the horseshoe furnace requires a greater height. A total height of 15-17 feet, between hearth and arch, is not excessive. This height should not be less than 10 ft. The furnace proper, that is the portion enclosed in the horseshoe, should have a height of about 8 ft., or 6.5 ft. as a minimum. Draught. Ample draught should be provided so that, when high ratings are required, the air jets from the tuyeres may penetrate deeply into the burning pile of bagasse. Otherwise, some bagasse would remain unburned, or at least combustion would be incomplete, with formation of CO, since the air has no other means of reaching the fuel, in this type of furnace. Air velocity in the tuyeres. The total cross-section of the tuyeres will be calculated so that the mean velocity of the air will lie between 50 and 65 ft./sec. A value of less than 50 should not be used, otherwise the tuyeres will become fouled and blocked. WARD FURNACE Weight of bagasse burned per unit area

A value of 200-250 lb./sq.ft./h is taken as a normal figure. In exceptional cases 300 lb./sq.ft./h may be attained but, at this rating, a large amount of unburnt bagasse particles is carried through to the chimney. HEATING SURFACE

The heating surface required in a sugar factory per t.c.h., depends on three factors: (a) The steam consumption of the factory per t.c. This may vary from a maximum of 1,680 pounds in a poorly equipped factory which is manufacturing white sugar, to a minimum of 1,000 pounds in a modern factory making raw sugar, and using pressure evaporation or thermo-compressors. Generally, steam consumption varies between 1,000 and 1,250 lb./t.c. (b) The type of boiler used, the normal evaporation rate of which may be high or low. (c) The greater or lesser intensity of boiler operation. Since it is advisable to allow some reserve capacity, it will be preferable to provide a heating surface corresponding to a normal rating. Hence we give in Tables 174 and 175, figures corresponding to normal and high ratings.

708

39

STEAM PRODUCTION

Tromp (p. 294) distinguishes factories according to their methods of clarification, and suggests the values of Table 176 as normal heating surfaces. TABLE 174 BOILER HEATING SURFACE REQUIRED

(eating surface • required (sq.ft./t.c.,h.)

T.

Steam required (lb It.c.) 1,008 1,120 1,232 1,344 1,456

Semi-tubular

Water-tube

410 460 503 547 591

273 306 328 361 394

TABLE 175 MINIMUM BOILER HEATING SURFACE FOR CONTINUOUS OPERATION AT HIGH RATINGS

Minimum heating surface (sq.ft./t.ch.)

Steam requirements (lb It c ) 1,008 1,120 1,232 1,344 1,456

Semi-tubular

Water-tube

328 361 405 438 470

208 230 252 273 295

TABLE 176 NORMAL HEATING SURFACE OF BOILERS,, ACCORDING TO TROMP

356 sq .ft./t.c.h. 404 sq .ft./t.c.h. 451 sq .ft./t.c.h.

Defecation factories Sulphitation Carbonatation

There exist cases where he has found 260 sq.ft./t.c.h. as sufficient, but these are in factories with an excellent heat balance. Noel Deerr (p. 469) gives very completefigures,which however do not correspond to presentday values. Statistics collected from more than sixty factories in Cuba (I.S.J., (1941) p. 192) have reported installed heating surface varying from 302-431 sq.ft./t.c.h., averaging 371 sq.ft./t.c.h. A survey of 36 factories in Porto Rico (T.S.J., (October 1950) p. 53) gives 203-655 sq.ft./t.c.h., averaging 336 sq.ft./t.c.h. SUPERHEAT AND SUPERHEATERS Use of superheated steam

Use of superheated steam is indispensable where power supply to the factory is by a steam turbine. Table 177 gives steam conditions often used. TABLE 177 CHARACTERISTICS OF STEAM FOR USE IN TURBINES

18 kg/cm 2 gauge: 20 kg/cm 2 gauge: 25 kg/cm 2 gauge:

300-350°C 325-375°C 350-400°C

250 p.s.i.g.: 300 p.s.i.g.: 350 p.s.i.g.:

600-650°F 625-700°F 650-750°F

39

CALCULATIONS FOR SUPERHEATERS

709

However, superheated steam is equally interesting in factories which have only reciprocating engines. Superheat has the advantage of avoiding or decreasing condensation losses on the cylinder walls, and of eliminating risks of water hammer at the engines. However, excessive superheat would cause trouble with the steam distribution valves and with stuffing-boxes of the engines, and would render lubrication difficult. Thus, in this case, superheat should be limited to 45-90°F; Izart (p. 908) states that the temperatures given in Table 178 should not be exceeded. TABLE 178 MAXIMUM TEMPERATURES ADMISSIBLE FOR SUPERHEATED STEAM

Engines with ordinary slide valves Engines with piston or Corliss valves Engines with drop-valves or sleeve valves

250°C or 480°F 275°C or 525°F 300°C or 570°F

There will then be no disadvantage in using a temperature of 437°F, and it will be advisable to keep to that figure. If this condition is realised, there will be every advantage in adopting the 90°F of superheat, recommended above as a maximum. Limit of superheat

With carbon steels, temperatures above 800°F in the superheater tubes cannot be used. With special steels, temperatures of 930°F are reached and exceeded, but the latter limit is of little interest in the cane sugar factory. Superheaters

Superheaters are heat exchangers placed in the path of the hot gasses. They are generally located in any suitable free space in the neighbourhood of the boiler tubes (Figs. 378 and 385). They receive the saturated or slightly wet steam coming from the boiler drum and deliver it in a superheated state to the general steam main of the factory. They are generally formed of tubes of small diameter, all of the same shape with several bends, interposed between two headers. Calculations for superheaters

We have two principal equations: M = aPC(Ti — T2) = p(\ — x)r + pc(T— t) M = quantity of heat transmitted by the superheater, in B.Th.U./h a = coefficient < 1, generally 0.90 p = weight of gas passing over the superheater, in lb./h C = specific heat of these gases T\ = temperature of gases at entry to the superheater, in °F Ti = temperature of gases leaving the superheater, in °F p = weight of steam to be superheated, in lb./h x = dryness fraction of the saturated steam (0.80-0.98 in general) r = latent heat of vaporisation, at the boiler pressure, in B.Th.U./lb. (cf. Table 149) c = mean specific heat of the superheated steam, between / and T° (cf. Table 150) t = temperature of the saturated steam, at the boiler pressure T = temperature of superheat desired or obtained

(611)

710

STEAM PRODUCTION

39

k = coefficient of heat transfer, in B.Th.U./sq.ft./h/°F, which varies from 4 to 6 according to the temperature and velocity of the hot gases. With bagasse, a value of 4 may be used. S = heating surface of the superheater, in sq.ft.

Eliminating M between (611) and (612), we have: s

~

p(\-x)r+pc(T-t) ,{Ô1 + ô2—ôÔÔÃ

(613)

Dryness fraction. It may be remarked that the influence of droplets of water entrained in the steam, p{\ — x)r often constitutes the greater part of the heat to be supplied to the steam. The dryness fraction of the steam for superheating is generally 0.96-0.98. In the cane sugar factory, figures are sometimes quoted of 18% moisture in steam (x = 0.82). It will be advisable therefore to measure the dryness fraction or to make generous allowance for it. Design of a superheater

We shall apply the preceding formulae to a concrete example: Battery of boilers consisting of 5 identical units Crushing rate of the factory 59 t.c.h. Working pressure of boilers 356 p.s.i. (437°F) Steam consumption per ton cane 1,176 lb./t.c. Weight of bagasse per ton cane 560 lb./t.c. Moisture in bagasse 45% (w = 0.45) Excess air used 50% (m = 1.5) Temperature of superheated steam desired 662°F Dryness fraction of saturated steam 0.95 Temperature of gases at the point where the superheater will be located 1,112°F It is required to calculate the superheater to be placed in each of the 5 similar boilers of this factory, which all operate together. Calculation ofPg. We have (eqn. 575): Pg = 5.75(1 — w)m + 1 = 5.74 lb./lb. bagasse Or a total of: 5.74 x 560 x 59 = 189,650 lb./h or per boiler: P = J ^ - =

37,930 lb./h

Calculation of p. The total weight of steam is: pt = 1,176 x 59 = 69,384 lb./h or per boiler: 69,384 p= —9-— = 13,877 lb./h

39

SUPERHEAT TEMPERATURE FURNISHED BY A SUPERHEATER

711

Calculation ofTz. From the eqns. (611), we have: ÷ = 7º or:

- | ^ - [(1 - x)r + c(T-1)]

(614)

13 877 T2 = 1,112 — —— — 1 — -—[(0.05 X 790) + 0.557(662 — 437)]J 0.90 x 37,930 x 0.30

or: Ã2 = 889°F We have taken C from eqn. (591) and c from Table 161. We may also, for the latter, apply the formula: (H2O) c = 0.468 + 0.000173(> — 32) (615) c = specific heat of water vapour at temperature t. Calculation ofS. The heating surface of the superheater to be installed in each of the 5 boilers is then (eqn. 613). c S =

13,877 [(0.05 x 790) + 0.557(662 - 437)] / 1,112 +889 662 + 437 x 4 ( 2 2 — )

=

^ 2 6 8 ■**'

Area of steam passage. The superheater should offer to the steam sufficient cross-sectional area of passage to give a velocity of the steam, calculated at the mean temperature (T + t)/2, of about 65 ft./sec (30-130 ft./sec). In order to obtain this velocity, which it is desirable to realise, it is often necessary to sub­ divide the superheater into a certain number of sections traversed by the steam in series. Value of Ti. If the temperature of the gases at entry to the superheater is not known, it may be measured, or an approximate value may be obtained by a calculation analogous to that for juice heaters (where the gases play the role of the heating steam in the heater, and the water in the boiler that of the juice), based on the fraction of the heating surface of the boiler already swept by the gases before reaching the position where the superheater is to be installed. The value of T\ should lie between 950 and 1300°F, i.e. the position chosen for the super­ heater should be one where the temperature of the combustion gases lies between those limits. Empirical rule. If higher temperatures of superheat are adopted for higher steam pressures, as is suggested by the values indicated in Table 177 and which is a logical practice, it is often satisfactory to dispense with the calculations for the superheater and to apply the following empirical rule: the heating surface of the superheater should be one third of the boiler heating surface: J = y

(616)

s = heating surface of the superheater S = heating surface of the boiler. Superheat temperature furnished by a superheater

Problem. A superheater is installed and it is desired to know what temperature it is capable of giving to the steam in certain conditions of operation.

712

39

STEAM PRODUCTION

We shall retain the nomenclature used previously. Here S is given, and the unknowns are: T2 and T. We eliminate M between eqns. (611) and (612), replacing T% by its value, eqn. (614), and solve for T. We have:

4£- (27º - /) + p[ct - r(l - *)l(V^F + 0 2

X2a C

kS 2

( kS

\

^

(617)

which gives the temperature sought. Variation of superheat with pressure. This same formula (617) will serve for studying the variation of superheat with the pressure at the boiler stop valve, the variables being then /and c. This problem is of interest in the rather frequent case where a high-pressure boiler is installed, and where it is necessary to use it temporarily at a lower pressure than that for which it has been designed. To solve this problem, it is of advantage to express eqn. (617) in the form: laPCkSTi —pr(\ — x) (laPC + kS) + [ilaPC + kS)pc — aPCkS]t aPCkS + (2aPC + kS)pc

(618)

while remarking that: (1) c is a linear function of /. The specific heat of water vapour is in fact given by the formula: c = 0.468 + 0.000311/ (m. units)

(615)

/ being the temperature of the steam under consideration. In consequence, the mean specific heat between / and T° has the value c = 0.468 + 0.000311 ß^-ã^)

(619)

(2) r is similarly a linear function of /: r = 607 — 0.7/ (m. units)

(274)

Taking these considerations into account, we obtain the equation (which we give in metric units only): 0.0001555/7Ã2 + (è.468/> +

ks^2aPC

)

T +

^6070 ~

x)

~

0J(1

"~ x)t

kSaPC -0.468/-0.0001555/2] _ _ _ _ _ _ ( which gives only one positive solution for T as function of /. Example. Numerical application. Given: S = k = x = Ti = p =

220 m2 20 kcal/m2/°C/h 0.97 600°C 14,000 kg/h

) (620)

2

Ã÷-/)= 0 J

39

713

MAINTENANCE OF SUPERHEATERS

P = 45,000 kg/h a =0.9 C = 0.3. Substituting in eqn. (620) the symbols by their values: 2.18Ã2 + 8,410×2.18*2 — 4,400* — 2,000,000 = 0 The positive root has the value: T= j/^640^00 + 2,000* + t2 —1,930 Using progressively higher pressures in the boiler, we shall obtain the following correspond­ ing values: Pressure = 15 kg/cm2 20 kg/cm2 25 kg/cm2 30 kg/cm2 40 kg/cm2

t = 200°C T = 214° 225° 235° 251°

323°C 332° 339° 344° 353°

Maintenance of superheaters

On account of their complicated form, it is not possible to clean the tubes of the superheater. Now, the droplets of water entrained with the wet steam are completely evaporated in the superheater. It is therefore in the superheater that any dissolved materials will be found which may have been introduced into the boiler with the feed water. It is necessary therefore to take great care that no water is allowed to enter a boiler fitted with a superheater, which may contain sugar or other material in solution. If the superheater becomes fouled with a deposit of carbon or a sort of tarry deposit of burnt sugar due to this cause, the cross-section of the tubes will be reduced, with consequent restriction on steam flow, while at the same time its heat transfer coefficient will be reduced. There is then no alternative but to replace the tubes, and to take all precautions to avoid a recurrence of the incident. ECONOMISERS

In the sugar factory, the boiler feed water is generally at a temperature of about 190°F. Now, the saturation temperature at which the water should be fed into the boiler to be transformed into steam varies, according to the pressure used, between 327°F (85 p.s.i.) and 455°F (427 p.s.i.). There is thus a large margin of temperature to be made up in the boiler. This margin means that a substantial proportion of the total heat has to be supplied to the water before evaporation proper commences. Assuming the feed water is at 194°F, Table 179 gives for three typical cases, the fraction of the total heat represented by this sensible heat. TABLE 179 FRACTION OF THE TOTAL HEAT REPRESENTED BY THE SENSIBLE HEAT TO BE SUPPLIED TO THE WATER {to = 194°F)

Steam press. (p.s.i.g.)

, c zupemeai

Saturation temp.

85 142 356

Saturated Saturated 662°F

327 362 437

(0f)

Total heat (B.Th.U./lb.)

Sensible heat (B.Th.U./lb.)

1,026 1,033 1,180

136 172 254

Sensible heat % total heat 13 17 21

714

STEAM PRODUCTION

39

Now the combustion gases leave the boiler at a temperature which is still relatively high, and generally above the saturation temperature. This sensible heat content of the gases would be lost in the stack. Hence the idea of utilising their sensible heat content to raise the tempera­ ture of the feed water, and so decrease the quantity of heat to be supplied to the steam in the boiler. This is the principle of the economiser. It is a heat exchanger placed in the path of the flue gases leaving the boiler, and through which the feed water is circulated between the feed pump and the boiler. It generally takes the form of tubes, in most cases with fins, through

Fig. 377. Finned tube for economiser (Cail-Steinmuller).

rr^

Fig. 378. Stirling boiler with horseshoe furnace, superheater and economiser.

39

715

CALCULATIONS FOR ECONOMISER

which the water circulates in series (Fig. 377). They are arranged in groups, the water passing from one tube to the following by means of a 180° bend. Two principal arrangements are adopted for economisers: (a) General economiser. When a single economiser is installed for the whole boiler station, it is placed in the main flue (Fig. 379). A by-pass is sometimes provided for the gases so that they may pass through the economiser in normal operation or may go direct to the stack, to permit of cleaning the economiser.

r ~\ Water

Gas

^

v

\

Fig. 379. General economiser installed in main flue.

(b) Individual economisers. It is now generally preferred to provide each boiler with its separate economiser. It is then placed in the last pass of the combustion gases leaving the boiler (Fig. 378). Calculations for economiser

We have 2 main equations: M = aPC(To — T) = pc(t — to) -fc(l

T— to = (To — t)e

-

«PC

(621)

r)S

(622)

M = quantity of heat transmitted by the economiser in B.Th.U./h a = coefficient < 1, generally between 0.90 and 0.95; frequently 0.93 P = weight of gas passing through the economiser, in lb./h C = specific heat of the gases Jo = temperature of gases entering the economiser, in °F T = temperature of gases leaving the economiser, in °F p = weight of water to be heated, in lb./h c = specific heat of the water = 1 to = temperature of water entering the economiser, in °F t = temperature of water leaving the economiser, in °F k = heat transfer coefficient for the economiser in B.Th.U./sq.ft./h/°F, generally between 1.5 and 3, according to the degree of cleanness of the exterior of the tubes, the velocity of the water in the tubes and of the gases around the tubes, and the temperature of the gases. As a mean value, we may take 2. S = heating surface of the economiser, in sq.ft. r = ratio: aPC pc

t — to To-

(623)

716

39

STEAM PRODUCTION

Equation (622) may again be written: To — t

, „ „

k(\ — r) In T—to

(624)

aPC

t

We may recall that: In x = 2.3 log x Eliminating aPC between (621) and (624), we have also: M-kS.r-O-V-'*

(625)

lo — t

ln-^= T— to First problem. Checking an existing installation. Determination of the heat transfer coef­ ficient k. Data: PC; pc; T0; T; t0; t\ S. Unknown: k. From eqn. (624), we have: aPC t To — t k = -^ r · In — — S(\—r) T—to Second problem. Design of an economiser. Determination of heating surface. Data: PC; pc; To; to and the desired temperature t. Unknowns: T; S. From eqn. (621), we have: T-To-**'-'* - 7 , - i ^ aPC r

,^x (626)

(627)

Hence S by eqn. (624), where we may take: jfc = 2 B.Th.U./sq.ft./h/°F. Numerical example. We require to calculate an economiser for the following case: Battery of boilers consisting of 5 identical units Crushing rate 59 t.c.h. Wood burnt 45 lb./t.c. Working pressure at boilers 356 p.s.i. Saturation temperature corresponding 437°F Steam consumption per ton cane 1,176 lb./t.c. Weight of bagasse per ton cane 560 lb./t.c. Moisture in bagasse 45% (w = 0.45) Excess air used for combustion 50% (m = 1.5) Temperature of gases leaving the boiler 464°F Temperature of water at feed tank 194°F Feed temperature required 284°F We require to calculate the heating surface of an economiser to be installed for each boiler. Calculation of P. We shall not take into account the quantity of firewood burned, since the object of installing an economiser is to eliminate this, which will probably be achieved. We have then (eqn. 575): Pg = 5.75(1 — w)m + 1 = 5.74 lb./lb. bagasse or, as total: Pt = 5.74 x 560 x 59 = 189,650 lb./h

39

717

SAFETY MARGIN

or for each boiler: P=

189,650 p ^ - = 37,930 lb./h

Calculation of p. The total quantity of water necessary as boiler feed has the value: pt= 1,176 x 59= 69,384 lb./h If the 5 boilers are working at the same rating, they will each receive: , - « £ ! * - 13,877 lb./h Calculation ofT. We have: 0.93 x 37,930 x 0.28 = 0.71 13,877 and: T = 464

284—194 — = 337°F 0.71

Calculation of S. We have then: _

0.93 x 37.930 x 0.282 2(1—0.71)

X

464 — 284 * ° g 337 — 194

or: S = 39,447 log 1.25 = 3,800 sq.ft. Variation of (f — f0) as function of S

It is of interest to ascertain in what proportion t varies, that is to say (/ — /o), when S is varied. If we adopt the following conditions assumed as standard: Moisture in bagasse Excess air Steam consumption per ton cane Weight of bagasse per ton cane Temperature of gases Temperature of water in feed tank

w = 0.45 m = 1.5 1,232 lb./t.c. 560 lb./t.c. To = 482°F to = 194°F

and if we calculate the heating surface for the economiser corresponding to a crushing rate of 10 t.c.h. (say for example an individual economiser for a boiler evaporating one-fifth of the steam for a factory of 50 t.c.h.), we obtain the graph of Fig. 380. This is obviously an exponential curve, and demonstrates that there is no advantage in seeking too great a gain in temperature; this would require an enormous heating surface, which would harldy be payable. With 8,000 sq.ft., a temperature rise of 144°F is obtained; with half this heating surface (4,000 sq.ft.) we obtain three quarters of this temperature rise, or 108°F. It requires approximately 24 sq.ft. to gain each of the first 10°F, whereas it requires 132 sq.ft. for a temperature rise of 1°F between 130 and 140°F. Safety margin

On the other hand, economisers are generally made of cast iron, to decrease the corrosive

718

STEAM PRODUCTION

39

action of flue gases and ashes. It is necessary then to avoid evaporation of water in the eco­ nomiser, for which it is not designed. It is necessary therefore to keep a safety margin between the temperature t of the water leaving the economiser and the boiling point corresponding to the working pressure of the boiler. Von Pritzelwitz (I.S.J., (1941) p. 143) recommends a margin of 72°F, Eigenhuis 36°F. We consider that it is wise not to go below 54°F.

Fig. 380. Variation in heating surface of economiser as function of temperature rise of water.

Ratio r

This is the ratio (t — to)l(T0 — T) between the rise in temperature of the water and the drop in temperature of the gases. It varies between 0.6 and 0.8, generally between 0.65 and 0.75. That is, the water temperature increases by 0.7° when the gases drop by Ã. Maximum economy

The maximum economy which can be obtained with an economiser is represented by the fraction of the total heat of the steam corresponding to the sensible heat, such as is given for example in Table 179, reduced by the heat quantity corresponding to the safety margin which must be allowed. It therefore depends on the steam pressure. With saturated steam, and keeping the safety margin of temperature at its minimum value of 54°F, it would be approximately as given in Table 180.

"ROTECO" ECONOMISER TABLE 180 MAXIMUM ECONOMY REALISED BY ECONOMISER

(Safety margin 54°F, saturated steam, feed water at 194°F) Steam pressure (gauge) ———— kgI cm2 6 8 10 12 14 16

.

Maximum economy possible

p.s.i.g. 85 114 142 171 199 228

8 % 10 % 11.5% 12.8% 14 % 15 %

Velocities

Economises are generally designed for the following velocities: (a) Water: 1.5-3 ft./sec, preferably 1.5-2 ft./sec. (b) Gases: 13-23 ft./sec, preferably 16-20 ft./sec. These values are those which offer the best compromise between the optimum conditions for heat transfer and permissible pressure drops.

Fig. 381. "Roteco" economiser (Delattre & Frouard). "Roteco" economiser

This is a very interesting patent combining in the same equipment an economiser and a fan. It may be described briefly as a rotary economiser, with water entry by a hollow shaft and water outlet similarly arranged at the other end (Fig. 381). The tubes are of copper, parallel to the axis, and furnished with fins of sheet metal, which produce the frictional effect required for the draught. The power required is very low, the heat transfer coefficient very high since the tubes are kept clean by the draught, and also by reason of the gas velocities in the equip-

720

STEAM PRODUCTION

39

ment. The efficiency is high (98-99%) the space required is very small (scarcely larger than that required for an ordinary fan), and the cost is lower than for a stationary economiser. AIR-HEATERS

In addition to the economiser, there is another type of equipment permitting of the partial recovery of the sensible heat of the combustion gases, which are still hot as they pass to the chimney. Instead of absorbing this heat in boiler feed water, it is absorbed by the air which is to be used for combustion in the furnace. The equipment is then called an "air-heater" or a "pre-heater". Types of air-heaters

There are 3 principal types of air-heater: (a) Tubular air-heaters. (b) Plate type air-heaters (Fig. 382). (c) Regenerative heaters.

Fig. 382. Plate type air-heater (Cail).

The first two types are ordinary heat exchangers. The difference between them is simply that the heat exchange surface consists in one case of tubes, and in the other case of plates of sheet metal. The third type is based on a different principle: flue gases and air alternatively are passed through flues containing brick chequer-work. The brickwork is heated by the passage of the gases, and gives up the heat thus stored when the gases are replaced by air. Alternatively, a suitable mass for absorbing the heat may be made to pass from the gas stream to the air stream and conversely. This is done in the Ljungstrom rotary air-heater, which is commonly used with steam boilers. These recuperative heaters have a rather low efficiency, but present the advantage of not being subject to corrosion, and of not requiring any cleaning. They are not employed in sugar factories to our knowledge. The term "pre-heater" is generally reserved for air-heaters of the first two types, with heat transfer surfaces of metal.

39

CALCULATIONS FOR AIR-HEATERS

721

Such heaters cannot be used with flue gas temperatures above about 930°F. At higher temperatures, the tubes or the plates, even if of cast iron, would rapidly become unserviceable. Tubular air-heaters are made with tubes of ordinary or chrome-copper steel, of 2\ in. o.d. The gases pass through the tubes and the air around them. Calculations for air-heaters

We have here exactly the same equations as for economisers (621 and following) where: p, c, t and txs apply to the air instead of to the water. Ratio r. In an air-heater: (1) The weight of flue gases is greater than the weight of air (since Pg = Pa + 1). (2) The specific heat of the gases is higher than that of the air (C > c). It follows that, in spite of the coefficient a, the ratio: aPC

' - " -

pc

To — T

(623)

is generally greater than 1. It varies between 1.2 and 1.6; its most frequent value is 1.3. That is, when the flue gases drop by Ã, the air rises in temperature by 1.3°. We may comment that if: > 1

To- -T

it follows similarly that: T- -to 10

t

so that the expression for S (eqn. 624) always has a plus sign. We may in fact write: aPC T—to aPC , To — t aPC T—to Ë n S = k{\ — r) ·' In T—to = ~ — k(r—l) — ·' In To — t

/iMOX

(628)

since:

"T-

-in A a

It will be necessary to make the same changes of sign in eqns. (625) and (626). Value of a. For air-heaters of metal with effective circulation, a value of 0.90 may be adopted for the coefficient a. Value of k. The coefficient of heat transfer in the air-heater is of the same order of magnitude as in an economiser. Tromp (p. 290) gives: Jfc = 2 B.Th.U./sq.ft./h/°F. Clayton (I.S.J., (1939) p. 389) gives, as a function of gas velocity: velocity of gases 16 33 ft./sec k 1.84 3.28 B.Th.U./sq.ft./h/°F. French firms take the rather low figures of: k = 1 - 2.5 B. Th. U. / sq. ft. / h/°F

722

39

STEAM PRODUCTION

In general, we may use: k = 1.6-2 B.Th.U./sq.ft./h/°F Velocities. The velocities generally used are as follows: (a) Air: 6-13 ft./sec; 13 ft./sec should not be exceeded. These velocities are given for cold air. The same cross-section may be retained for hot air, the velocity of which will therefore be higher, but should not exceed 33 ft./sec. (b) Gases: 16-26 ft./sec. Design of air-heater

It is required to calculate the heating surface of a metal air-heater with systematic circulation. Data: PC; pc; TO; to and the temperature desired t. Unknowns: T; S. Numerical example. We shall calculate the heating surface for the following case: Battery of boilers consisting of 5 identical Crushing rate of factory Wood burned Weight of bagasse per t.c. Moisture in bagasse Excess air used for combustion Temperature of gases leaving the boiler Temperature of ambient air Temperature of hot air required

units 59 t.c.h. 45 lb./t.c. 560 lb./t.c. 45% (w = 0.45) 50% (m = 1.5) 464°F 86°F 275°F

We require the heating surface to be installed for each boiler. Solution. We may assume that the above data are those applicable before installation of the air-heater. It is logical to calculate the air-heater on the assumption that it will give a certain improvement in operating conditions: (a) Elimination of consumption of firewood. (b) Reduction of excess air from 50 to 40%: m = 1.4. (c) Increase of the temperature of gases leaving the boiler from 464° to 482°F. Calculation of P. We have (eqn. 575) taking m = 1.4: Pg = 5.75(1 — w)m + 1 = 5.43 lb./lb. bagasse or as total: Pt = 5.43 x 560 x 59 = 179,407 lb./h and per boiler: P=

179 407 y = 35,881 lb./h

Calculation of p. The weight of air necessary will be (eqn. 574): Pa = 5.75(1 — w)m = 4.43 lb./lb. bagasse or as total: pt = 4.43 x 560 x 59 = 146,367 lb./h

39

111

VARIATION OF (t — to) AS A FUNCTION OF S

and per boiler: p

146 7

=

f

- 29,273 lb./h

Calculation ofT. We have: 0.90 x 35,881 x 0.283 29,273 X 0.24

1.30

We have calculated C according to eqn. (591). For -c, the value varies only slightly with temperature, as is readily seen from the columns N2 and O2 of Table 161, and no appreciable error will be made, in calculations for the air-heater, if we take in all cases: c = 0.24. To­

to

r

275 — 86

= 482-

337°F

1.30

Calculation of S. Taking for heat transfer coefficient the value: k = 2 B.Th.U./sq.ft./h/°F: S=

0.90 x 35,881 x 0.283 ôôÃ-^ 2(1.30—1)

_ 337 — 86 x 2.3 log -482 — 275

or: S = 35,032 log 1.213 = 2,938 sq.ft. 36

72

108

144

180

216

252

288 °F

8000 S sq.ft 7000 6000

H5000 -4000 3000

H2000 H1000

Fig. 383. Variation of heating surface of air heater as function of temperature rise of air.

Variation of (t — to) as a function of S

We may now calculate, as for the economiser, the variation of the rise in temperature of the air (t — to) as function of the heating surface S employed.

724

39

STEAM PRODUCTION

Adopting the following conditions, which have been purposely chosen to correspond to those adopted for the economiser: Moisture in bagasse Excess air Weight of bagasse per ton cane Temperature of flue gases Temperature of ambient air

w = 0.45 m = 1.4 560 lb./t.c. To = 500°F to = 86°F

and plotting as ordinates the air-heater surface corresponding to a crushing rate of 10 t.c.h., we obtain the graph of Fig. 383. It exhibits exactly the same form as that found for the economiser, and prompts the same comments. Limit of temperature obtainable

Up to what point is it possible to pre-heat the combustion air? Tromp (p. 293) indicates 400°F as a temperature obtained in practice. Clayton (I.S.J., (1939) p. 387) states that preheating up to 482°F may easily be obtained, but considers it wise not to exceed 356°F, otherwise damage will be caused to the grate bars and to refractories in the furnace, on account of the increased temperature of combustion. However, special refractories are made for very high temperatures; but their cost would appreciably reduce the advantage of the additional gain in temperature sought. We consider that with good quality refractory bricks as generally used in French cane sugar factories, there would be no disadvantage in adopting 400°F as the limit of temperature to which the air may be preheated. Increase in combustion temperature

What is the effect of the increase in the temperature of the combustion air on the temperature of combustion? Referring back to eqn. (590), we might consider that the combustion temperature would increase by exactly the same amount as the increase in air temperature. However, this formula contains an approximation which was justified when the temperature of the fuel was the same as that of the air and varied with it. This is not the case here, so we shall repeat the calculation. We shall designate by: to = the temperature of the bagasse and the ambient air / = the temperature of the heated air To = the temperature of combustion which would be obtained with the cold air T = the temperature of combustion obtained with the heated air. Replacing t and Tby to and To, eqn. (587) gives the combustion temperature To in the case of cold air. If the air is heated, the supplementary terms to be added to the two sides of eqn. (587) should themselves be equal. Hence: f Jt

Pa · cadt - Ó

o

f

PC · at O

(629)

J T

or, considering the mean specific heat: Ñá[€á]10- (t — to) = Ó P[C]TQ-

( Ã — To)

(630)

39

725

CORROSION

Hence: T—To t — to

or:

Pa[Ca]t m _ô ncjr0

0.24P« 0.33Ñ,

0.24 x 5.75(1 — w)m [0.33 x 5.75(1 — w)m] + 0.33

T—To t — to

1

1.375 +

(631)

1

4.18(1 — w)m

Giving extreme values to w and to m, it will be seen that the ratio:

»lies between:

T T

~°

(632)

t — to

0.55 and 0.64

For mean values: w = 0.45; m = 1.5; we have: 0 = 0.6

(633)

In other words, in a bagasse furnace, the increase in the combustion temperature due to pre-heating of the air is 60% of the increase in air temperature. Corrosion

Dew-point. Air-heaters are generally fabricated in steel. The dew-point of the combustion gases lies in the region of 140-150°F, hence it may be supposed that the danger of condensation on the walls of the heater, and consequently of corrosion, would be very limited, since the gases leave the air-heater at a temperature which is always greater than 212°F. In practice, there always exist points where pockets of gas are formed, the temperature of which can then drop to the dew-point. Recirculation. To avoid such risk of corrosion, it is arranged that the temperature of the air itself does not fall below 150°F. This is obtained by using "recirculation" of portion of the hot air. This consists simply of making a certain proportion of the hot air, leaving the heater, describe a close circuit: it is taken and returned to the air inlet.

(t) , '''//''///ú'-(t)■'

(to) (t'o)

/

■'

Fig. 384. Proportion of air re-circulated.

726

39

STEAM PRODUCTION

Calculation of proportion of air to be recirculated. Let x be the fraction of the air used which is to be recirculated (Fig. 384). We have then: 1 . , 0 + x . / = (1 + x)t'o

(634)

whence:

'.-'„ ' - < ;

to' = dew-point (approximately 150°F) to = temperature of ambient air (generally 86°F) t = temperature of hot air leaving the heater. Or, in the general case: =

150 — 86

64

* -7-isr = T=i»

(

}

It is seen that, for: t = 212°F, the quantity of air to be passed through the heater will be double the quantity used for combustion. Consequences of recirculation. When the temperature of the hot air /, is sufficiently high, recirculation presents little difficulty. The efficiency of the heater is not appreciably altered, since the increase in velocity compensates for the decrease in temperature difference between gas and air. The pre-heater will therefore be designed as if there were no recirculation, and the tempera­ ture of hot air obtained will hardly be altered. Lagging

It is necessary to insulate carefully the hot air duct passing from the air-heater outlet to the furnace, otherwise a large proportion of the heat recovered will be lost. Maximum economy

The maximum economy obtainable from an air-heater does not depend, as is the case with an economiser, on the steam pressure of the boilers, but only on the maximum temperature permissible for the air. Adopting a value of 400°F as the maximum permissible temperature, the maximum economy may go as high as 10 or 12%. It is seen then that the maximum heat economy obtainable by means of a heat exchanger in the flue gases is greater: (a) with an air-heater, if the boilers operate at low pressure, (b) with an economiser, in the case of high-pressure boilers. Choice between economiser and air-heater

What determines the economy realised, is the temperature to which the flue gases may be reduced. Any comparison between economiser and air-heater should therefore be made on a basis of equal final temperatures of the flue gase* passing to the stack. We may comment first that, in the case of good modern boilers, each of these two heat exchangers permits of the maximum or optimum econo:ny being realised, which corresponds to gases discharged at about 300°F.

39

CHOICE BETWEEN ECONOMISER AND AIR-HEATER

727

Consequently the factors which guide the choice to be made are the following: (1) Heating surface, (a) The heat transfer coefficient should be much better in an economiser (gas to water) than in an air-heater (gas to gas). In practice, we have seen that the difference is not significant: approximately 10% in favour of the economiser. (b) The mean difference in temperature between heating fluid and heated fluid, which influ­ ences the logarithmic term in the formulae (624) and (628) is generally greater for the air-heater, so reducing the heating surface required. Finally, taking these two factors into account, the air-heater will have a smaller heating surface for an equal saving. (2) Cost, (a) The economiser operates under pressure, and is of cast iron. The air-heater works at atmospheric pressure, hence is not subject to pressure (and is of steel). It should therefore cost much less. (b) However, it requires an additional fan, in the case where forced draught is not already used. Finally, the purchase price is rather definitely in favour of the air-heater for equal heating surface, and still more definitely for equal savings, by reason of the conclusion of paragraph (1) above. (3) Upkeep and working life. Either device requires little maintenance, but the economiser, made of thick cast iron, has a much longer life than the air-heater, which is made of thin sheet steel. (4) Supplementary advantages. The air-heater improves combustion, and generally permits of a better control of excess air. This is its greatest trump card. Conclusion. In the final reckoning, the advantages and disadvantages balance one another very closely. The choice will most often be guided by questions of detail, of convenience of installation, or of ease of insertion in a new installation. A horseshoe furnace, for example, requires forced draught in any case and so favours the air-heater. A neat solution consists of installing both types of equipment in the path of the flue gases. We may for example place the economiser first, followed by the air-heater. This is the ar­ rangement which is theoretically most logical: it is of advantage to stop the heat exchange in any heat exchanger when the temperature of the hot fluid approaches too close to that of the cold fluid, and then to transfer to exchange of heat with a cooler fluid. However, by placing the air-heater first in the gas circuit (Fig. 385), it may be possible to dispense with recirculation; this is a great advantage and, for certain types of boiler, can be conducive to a simpler instal­ lation. Such use of the two heat exchangers presents several advantages: (a) A large safety margin for the economiser, since only part of the possible temperature rise is being sought. (b) No risk of deterioration of refractories due to the high air temperature, for the same reason. (c) In the case of a breakdown of the first heat exchanger, there is a possibility that the second exchanger will compensate for it to a certain extent, consequent on the increase of temperature in the gases entering it.

728

STEAM PRODUCTION

39

Fig. 385. Cail-Steinmuller boiler with horseshoe furnace, superheater, air-heater and economiser.

Efficiency

When the efficiency of an economiser or air-heater is calculated, it is sometimes found to reach or exceed 1, instead of lying in the neighbourhood of the value expected: a = 0.90. This phenomenon is explained by the fact that part of the unburnt solids and of the hydrogen contained in the gases continue to burn while passing through the air-heater. (There would be no question of CO, which requires much higher temperatures.) BOILER SETTINGS Sizes of bricks

Certain factories use the standard French brick of 220 X 110 X 60 mm (9 X A\ X 2.3/8 in.). Others prefer the bricks of 300 x 150 x 75 mm (12 x 6 X 3 in.). The larger bricks have several advantages: (a) They are more easily and rapidly laid. (b) They therefore require less labour. (c) They consume less mortar, since they have a smaller area of joints for the same volume. (d) They cost somewhat less per unit volume.

39

CROSS-SECTION OF FLUES

729

They have on the other hand one disadvantage: since their dimensions are greater they are somewhat more brittle, and it is less easy to give walls the desired dimensions. It is necessary to take a multiple of their length and/or of their width. Resistance to temperatures

It is now possible to obtain temperatures of 2,470°F in ordinary furnaces with cold air, 2,650°F in furnaces supplied with pre-heated air. We should choose: (a) In the former case, bricks with at least 20-22% of alumina. (b) In the latter case, bricks of at least 30-33% alumina. The prices of these two types of refractory are about 25% and 50% higher, respectively, at the point of manufacture, than the price of ordinary bricks of 15-18% of alumina. However, when they are purchased at a great distance from the sugar factory, this difference in cost price becomes less important, since expenses of handling, transport, and freight are the same for all types of brick, and consequently independent of the quality. Expansion joints

In all brick walls of the furnace and the boiler setting, it is necessary to allow expansion joints of 3/16 in. every 2 ft. They will not however be placed in the walls of the rhorseshoe or a furnace of hearth type. In the furnace arches, the bricks should be arranged in separate rows. The arch thus becomes a series of independent small arches in juxtaposition, of length equal to that of one arch brick. This has the advantage of avoiding half-bricks. Thus an expansion joint of 3/16 inch can be arranged between these small arches, every three arches. Drying out new brickwork

When construction of furnaces or flues is finished, they should be dried out very carefully, by lighting a small fire, which will be increased progressively during at least three days. Cross-section of flues

(1) Velocity of gases. A cross-section should be adopted for flues such that the velocity of the gases is: (a) In the case of natural draught: 13-16 ft./sec; preferably 15 ft./sec. (b) In the case of forced draught: 16-60 ft./sec; preferably 20-26 ft./sec. (2) Volume of gases. Equation (580) gives us: Vg = 71.3(1 — w)m + 9.17M; + 10.75

With w = 0.45 and m = 1.75: Vg = 71.3 X 0.55 X 1.75 + 9.17 x 0.45 + 10.75 or: Vg= 83.5 cu.ft./lb. bagasse at N.T.P. Or at 482°F: Vgt = 83.5 x

459 4_ 482

= 160 cu.ft./lb. bagasse

730

39

STEAM PRODUCTION

(3) Cross-section of flues. Basing our figures on a gas velocity of 15 ft./sec (for natural draught) and on a volume of 160 cu.ft. of gas/lb. bagasse, the cross-section of the flues will be: s= or:

160

3,600 x 15

= 0.003 sq.ft./lb. bagasse/h ; /

s = *6.6 sq.ft./ton bagasse/h

(636)

With a gas velocity of 23 ft./sec (forced draught), it would be sufficient to allow s = 0.002 sq.ft./lb. bagasse/h or: s = 4.3 sq.ft./ton bagasse/h DRAUGHT

In order to maintain the temperature and the rate of combustion, it is necessary to pass through the furnace and over the fuel bed the required quantity of air. Since the path of the gases is complex, with many resistances to overcome (passing through the fuel bed and between the boiler tubes, sudden changes of direction, etc.) and since it is necessary to give the gases a rather high velocity, as seen in the foregoing figures, maintaining this flow of gas demands a certain expenditure of energy which is normally supplied in the form of a motive pressure. This actuating pressure may be obtained from two sources: (a) from the thermal energy produced by the combustion: this is natural draught, (b) from an external source, fan or some other means; this is forced draught. Natural draught

With natural draught, the suction is created by evacuating the combustion gases by a chimney. These gases being hot, the weight of the gaseous column so formed is lower than that of the same height of ambient air. The body of gas contained in the chimney therefore tends to rise, under pressure from the ambient air, which progressively replaces it and becomes heated in its turn in traversing the furnace. The realisation of a good natural draught is more delicate than that of a forced draught. It demands a knowledge of relationships existing between the following quantities: (1) (2) (3) (4) (5)

Velocity of gases in the flues Draught at the entry to the chimney Velocity of the gases leaving the chimney Cross-section of the chimney Height of the chimney

Vc d V8 S8 H

(1) Velocity of gases in the flues

We have just seen that this velocity should not exceed 16 ft./sec in the case of natural draught. We shall take: Vc = 13-16 ft./sec (637) (2) Draught at the outlet from the flues

The draught or suction is the difference between the outside pressure and the pressure in the

39

731

DRAUGHT AT THE OUTLET FROM THE FLUES

interior of the flue. It is expressed in inches of water, and is easily measured in an existing installation, with the aid of a small U-tube containing water (Fig. 386).

TO i

Fig. 386. Measure of draught.

By application of Bernouilli's theorem Tripier has deduced the following formula: /

V2

d = 0.192ßù^ (1 + Ö) -ã-

d wgm ùá zc zg Vc g Ö

\ (ù á — <»gm) (zc — zg)\

(638)

= draught at outlet from the flues (i.e. entry to the chimney), in inches of water = mean density of the gases between the grate and the entry to the chimney, in lb./cu.ft. = density of the ambient air, in lb./cu.ft. = height of the centre of the entry to the chimney, in ft. = height of the centre of the grate, in ft. = velocity of the gases at outlet from the flues, in ft./sec = 32.16 ft./sec/sec; 2g = 64.32 = numerical coefficient representing, as a ratio to Kc2/2g, the sum of the losses of head from the grate to the outlet from the flues.

Value of Ö. We may estimate Ö by summing the coefficients for the following losses oMiead: (a) Grate: Ö\ = 5-10 according to the nature of the grate, the weight of bagasse and the thickness of the layer of bagasse. (b) Boiler: Ö% = 4-9 according to type. We may use 6 for a boiler with tubes of the Babcock or Steinmuller type, 7 for a Stirling, 9 for a semi-tubular with 3 gas passes. (c) Economiser: Ö$ = 3-6 according to type: 5 for an economiser with gilled tubes of the Steinmuller or Green type. (d) Air-heater: Ö 4 = 3-4. (e) Friction in the flues: Ö5 = 0.3-0.6/10 ft. length of flue, according as the velocity is less than 13 ft./sec or greater than 16 ft./sec. (/) Bends and changes of direction in the flues: Ö 6 = 3-5 according to the number and the angle of the bends. Approximately: 1 per 90° bend. (g) Changes in section in the flue, we may take: Öç = 2 for a normal number of changes in cross-section (2 or 3). As a total, we have generally: 20 < Ö < 33

732

STEAM PRODUCTION

39

and nearly always: 10 < Ö < 50 For an average present-day installation, a value of Ö = 29 is often taken, or: 1 + Ö = 30. Value of ù. The density of air or of gases is obtained by the formula: ù = -£=w = density of the air or the gases of combustion, in lb./cu.ft. p = pressure, in p.s.i. (atmospheric pressure = 14.696 p.s.i.) (0.37 for air \ 0.363 for combustion gases from bagasse T = 459 + f¥ t = temperature of the air or gases under consideration. In other words: 39 7 40 5 · · (Da = - ã -

OJg = -ø-

(639)

,*M

(640)

The mean density of the gases between grate and entry to the flues may be assumed equal to density corresponding to the mean temperature Tgm: Tgm = ]/TM-TC

(641)

Tgm = mean absolute temperature adopted TM = maximum absolute temperature in the furnace (temperature of combustion) Tc = absolute temperature of the gases at entry to the chimney. (3) Velocity of the gases leaving the chimney

The velocity of the gases leaving the chimney corresponds to a loss of energy. From the point of view of economy in the installation (minimum height of chimney to be constructed, for example), it is of advantage that this velocity Vs should be as low as possible. However, it must be sufficient to avoid disturbances to the functioning of the chimney due to the vertical component of prevailing winds. In tropical countries, it is of advantage to have at least 13 ft./sec, and even 16 ft./sec. It is not necessary to exceed the latter figure, since this would involve building a chimney unnecessarily high and expensive. We shall therefore choose the velocity Vs in such a way that: 13 < Vs < 16 ft./sec

(642)

according to the importance of the prevailing winds. (4) Cross-section of the chimney

The choice of velocity of gases leaving the chimney involves the determination of its cross section at the top. We know the output of gas to be allowed for:

39

ADVANTAGES AND DISADVANTAGES OF NATURAL DRAUGHT

733

Q = volume of gases to be handled by the chimney, in cu.ft./sec B = weight of bagasse burned in the furnaces served, in lb./h Vgt = volume of combustion gases, given by eqn. (580), and converted to the temperature and pressure ruling at the top of the chimney. The temperature of the gases at the top of the chimney may be taken equal to: is = tc — a(tc — to)

ts tc ta a

= = = =

(644)

temperature of gases at the top of the chimney temperature of the gases at the base of the chimney ambient temperature coefficient having a value: 0.05 for brick or stone chimneys 0.09 for chimneys of reinforced concrete 0.15 for steel chimneys

(5) Height of the chimney

The height of the chimney should be determined according to Tripier's second formula:

H = height of the chimney, in ft. d = draught at base of chimney, in inches of water, given for the case of natural draught by eqn. (638) pa = ambient pressure at the level of the base of the chimney, in p.s.i. (1 atm. = 14.696 p.s.i.) Ta = absolute temperature of ambient air = 459 + ta°F Tc = absolute temperature of gases at the base of chimney = 459 + tc°¥ b = numerical coefficient taking into account the cooling of the gases while passing through the chimney, and having the value: 1.05 for brick chimneys 1.08 for chimneys of reinforced concrete 1.12 for steel chimneys The optimum economic height is generally of the order of 130-165 ft. Tromp (p. 298) gives, for ordinary bagasse furnaces, a draught d of 0.6-0.8 in. To ensure obtaining this, he advises building chimneys of 150-200 ft. in height, in order to cope with possible unfavourable conditions: high excess air, or good thermal efficiency. Advantages and disadvantages of n a t u r a l d r a u g h t

Natural draught presents advantages: (a) Long life: a chimney lasts for 100 years. (b) Dependability: there is no risk of breakdown of a fan. (c) Economy in operation: no motor is required, no power is consumed. However, it has some disadvantages: (a) It requires a good foundation, on account of the weight of the chimney. (b) It takes up considerable space, on account of the dimensions of the base of the chimney.

734

STEAM PRODUCTION

39

(c) It can produce only a limited draught, unless an excessive height of chimney is used. (d) It lacksflexibility,when peak loads occur. Mechanical draught

There exist 3 principal systems of mechanical draught: (1) Forced draught. (2) Induced draught. (3) "Injection" induced draught. (1) Forced draught

The air may be blown in under the grate, the ashpit being enclosed. This system has the advantage of permitting introduction of air at atmospheric pressure into the combustion chamber, and consequently avoids any entry of air by leakage, in spite of cracks or leaks in the boiler setting. It is mainly employed in the case where an air-heater is installed. (2) Induced draught

This is the commonest system. Instead of placing the fan ahead of the grate, it is placed at the extremity of the flues, at the base of the chimney. (3) Injection draught

This is another form of induced draught: instead of a fan, the suction is produced by a steam nozzle discharging into the chimney, and producing a Giffard effect on the gases. Alternatively, a fan placed outside the circuit may be used to produce the Giffard effect, by taking from the flues a portion of the gases and returning this gas by means of a nozzle similar to that utilised for steam. Advantages and disadvantages

Mechanical draught requires a fan, hence the possibility of breakdowns and stoppages, necessity for supervision and maintenance; but it is being employed more and more for the following reasons: (a) its first cost of installation is much lower; (b) less space is required; (c) (and most important) it is very flexible: with a variable speed motor the draught may be regulated immediately and as often as desired. The motor may even be fitted with an auto­ matic draught regulator, which assures regular and flexible operation of the factory, and allows the production of steam at any moment to be proportioned to the varying needs of the factory. FANS

Previously large slow-speed fans were used. In modern practice, high-speed fans are preferred, which are much smaller and so more economical in first cost and space required. Power requirements

If a fan has to supply a volume of gas V cu.ft./sec while maintaining a draught of d inches of water the power used will be:

39

POWER REQUIREMENTS

735

à = 5.2 Vd

(646)

T = nett power to be supplied, in ft.lb./sec V = volume of gas to be handled, in cu.ft./sec d = draught produced, in inches of water. The actual power required by the fan will then be: Vd T=5.2

(647) Q

ñ = efficiency of the fan. This efficiency is very low, and varies substantially with the power of the fan. We may use the figures of Table 181. TABLE 181 EFFICIENCY OF FANS

Small fans 0.20-0.50, average: 0.30 Large fans 0.40-0.70, average: 0.50

It must not be forgotten that the power Tis only the power supplied to the shaft of the fan. For the power supplied to the fan motor, it is necessary to allow for the efficiency of the motor, and also that of the belt, where belt drive is used. Influence of gas temperature. It will be noted that the power T indicated above depends only on the volume V of gases handled. Consequently, for a given output in terms of weight of gas, more power is required when the gas is hot than when it is cold. Practical formula for power. It is difficult to choose or to know a priori the efficiency of the fan, which enters into eqn. (647). Furthermore, a small variation would be enough to introduce a serious error, on account of the very low efficiencies under consideration. Shillington has published figures (I.S.J., (1939) p. 261) more particularly applicable to induc­ ed draught fans of the Prat type, from which the following approximate formula for power may be derived: T= T B d /

= = = =

Bd(46

° + 435

t)

(648)

power required at the fan shaft, in h.p. quantity of bagasse consumed by the furnace or furnaces served by the fans, in tons/h draught at suction of fan, in inches of water gas temperature at suction of the fan, in °F.

Example. We require the power necessary to drive a general fan serving all the boilers in a factory crushing 59 t.c.h. and producing 560 lb. bagasse/t.c. Draught required: d = 2 in. at suction of fan. Temperature of gases entering the fan: / = 392°F. We have: B=

59 x 560 ,_ , , = 14.75 tons bagasse/h 2,240 '

736 and:

STEAM PRODUCTION

39

i 4^> L 2_(46^ L 392)_ = =

435

Shillington recommends installing a fan of ample capacity, since, if there is need to force the boilers, the efficiency falls very rapidly when normal output is exceeded, whereas it drops only slowly when output is decreased below that corresponding to the optimum efficiency. In other words, a large fan, having a large margin of capacity, is preferable to a smaller one working at full capacity. BOILER FEED WATER R e t u r n e d condensates

The sugar factory has a large number of condensates available from the various heat exchangers: multiple effects, juice heaters, vacuum pans, etc. These condensates are generally pure, since they have been boiled and condensed as distilled water. It is necessary, however, to classify them according to their origin: (a) Water derived from condensation of the steam coming directly (heat exchangers heated by live steam) or indirectly (heat exchangers using exhaust steam) from the boilers. (b) Condensates originating from vapour derived from the juice: condensate returned from the second and subsequent vessels of the multiple effect, from juice heaters or other heat exchangers heated by vapour bled from the multiple effects. Condensates of the first type involve little risk of being contaminated. Since they originate from steam under pressure, even if a tube develops a leak in the first effect, for example, it will be steam that will leak into the juice, rather than the opposite. The greatest risk originates in the juice heaters, where the possibility of a split tube may admit juice under pressure into the steam space. Condensates of the second type are more dangerous. They involve first the same risk of direct contamination by juice, aggravated by the fact that the juice vapours are most often under vacuum and consequently more subject to entry of juice under pressure through leaks. However, the main risk is that these condensates may contain sugar originating in entrainment in the evaporators. Even if this is present only in imperceptible traces, these traces will be returned to the boilers and will end up by accumulating on the boiler tubes a harmful and dangerous carbonaceous deposit. It is important then to separate carefully the returned condensates: (a) from direct or exhaust steam, (b) from vapour derived from juice. Utilisation of condensates. Condensates of the first group only should be sent to the feed water tank. Those of the second group may be utilised for imbibition, washing of filter cakes, dilution of molasses, etc. If there is an excess of them, it is preferable to discard the excess rather than send it to the boilers. Make-up water

The steam in the sugar factory describes a closed circuit: evaporated in the boilers, condensed in calandrias, it returns to the boilers by way of the feed water tank.

39

CAPACITY OF FEED WATER TANK

737

It would therefore be possible to use the steam in this closed cycle, without addition of water, if it were not for the following losses, which must be compensated: (a) Steam lost to atmosphere: Leakages at joints and stuffing-boxes Operation of safety valves Operation of soot blowers Engines exhausting to atmosphere: steam derrick, etc. Steaming out of filter presses Cleaning of mills and cush-cush plant with steam Steam used at centrifugals. (b) Water lost to the drain: Washing out of piping Blowing down of boilers. (c) Steam partially lost in the molasses: Washing out of pans Dilution of molasses. The total of these losses of steam or condensates represents, according to circumstances, from 10-20% of the steam produced at the boilers. In order to replace this quantity, it is necessary to return to the feed water tank either: (a) cold water, or (b) portion of the condensates of the second group above. These condensates in effect originate from the juice, that is they enter into the factory with the cane. They thus consist of water from an outside source. However, on account of the risks involved, their use is reduced to a minimum, and they are taken from the point which presents least danger; that is the second vessel, the condensates from which present the minimum risk of containing sugar due to entrainment. In order to keep the contribution of condensate from the second effect to the minimum strictly necessary, this condensate should be discharged into a tank alongside the feed water tank and arranged so as to deliver into the latter only by a float-operated valve which opens when the level in the feed tank has dropped below a certain limit. The addition of second effect condensate should normally be sufficient to supply the make-up necessary. Capacity of feed w a t e r tank

To avoid the necessity of supplementing it with cold water, it is necessary to allow in the feed water tank a reserve capacity of water sufficient to cope with the sudden demands of the boiling house. Tromp (p. 308) estimates that the fluctuations in instantaneous steam consumption in a sugar factory amount to 25% above and below the mean. These fluctuations generally do not last for more than one hour and, if the peak represents ±25% on normal consumption, the mean increase or decrease in consumption during the hour in which it occurs represents only ±15% of the average consumption for the day or for the week. Assuming a high consumption of 1,456 lb. steam/t.c, we see then that the difference between maximum consumption and minimum return of condensates represents approximately:

738

STEAM PRODUCTION

39

1,456 (1.15 — 0.85) - 437 lb. water/t.c. However, the time elapsing between the steam leaving the boiler and the return of the corre­ sponding condensate to the feed tank represents only £-£ hour. It will be seen then that it will be amply sufficient to provide a feed tank of capacity equal to 225 lb./t.c.h., in order to cope with fluctuations in steam consumption without having recourse to an avoidable addition of cold water: c = 22C

(649)

c = capacity of the feed water tank, in imperial gallons C = crushing rate of the factory, in t.c.h. A factory of 60 t.c.h. should therefore have a feed tank of 1,320 gallons capacity. We may comment in passing that this reserve capacity does not depend on the steam economy of the factory: it should be as large for a factory using 1,120 lb. steam/t.c. as for one using 1,456 lb./t.c, since the variations in consumption, due predominantly to the pans, are the same in the two factories if they are expressed in weight of steam per ton of cane. It is of advantage to provide the feed tank with two float-operated valves: the first controlling the addition of water from the second vessel, the second the addition of cold water. It will then be necessary to allow a sufficient margin between the different levels, say approximately 1 /3 of the height of the tank: (a) between the overflow level and the opening of the float-operated valve controlling condensate from the second vessel (b) between the opening of this condensate valve and the point of opening of the floatoperated valve for cold water (c) between the opening of the cold water valve and the bottom of the tank. Feed water pump

French law requires that there should be at least two feed water pumps. Generally, the pump in normal use is motor-driven and the stand-by pump is a duplex steam pump, of Burton type. At the delivery side of the pump, a non-return valve should be provided with the object of avoiding the risk, when the pump is stopped, that water can run from the boiler back towards the feed tank if the delivery valve does not close properly. The feed pump should be capable of an output of at least 50% greater than the mean steam production of the boilers which it supplies. It should be capable of pumping against a pressure head at least 25% greater than the working pressure of the boilers. On account of the temperature of the feed water (194-203°F), it is of advantage to place the feed tank at least 3-7 ft. above the level of the pump; preferably 8 ft. (cf. p. 823). Feed water pipes

The diameter of the piping should be chosen so as to obtain the following velocities: (a) Suction pipe, from the feed tank to the pump: v = 3 ft./sec (b) Delivery pipe, leading from the pump to the boiler: v = 6 ft./sec. Influence of feed temperature on fuel consumption

The last column of Table 168 shows that an increase of 11°F in the feed water temperature corresponds approximately to a saving in fuel of 1%.

39

PRINCIPLE OF STEAM ACCUMULATORS

739

Reaction of feed water

The corrosive properties of feed water depend on its pH. It is at a minimum, not for a pH of 7, but for a slightly alkaline value. A certain alkalinity is maintained in the feed water to prevent possible corrosion, and this is done by introducing hydroxyl (OH) ions into the feed water. When a pH of 9 is reached, a protective film of ferrous hydrate is formed on the metal. On the basis of this fact, Brola (p. 54) recommends that a pH of 9.5 be maintained. An English recommendation (I.S.J., (1945) p. 188) is 8.3 pH as the lower limit, and it is stated that it is mainly in the economiser that corrosion is manifest. Leggett, in India (I.S.J., (1944) p. 39) recommends that a pH of 7.6 be maintained in the feed tank, which corresponds to 8.6-9.0 in the boilers. Gregory, in Cuba (I.S.J., (1947) p. 66) recommends a pH of 10.5-11. The present author would recommend maintaining a pH of 8 in the feed tank. Alkalizing agents. In order to introduce hydroxyl ions into the feed tank, we may employ: caustic soda NaOH sodium carbonate Na2C03 trisodium phosphate. Neumann, in Java (F.A.S., (Sept. 1940) p. 30) recommends trisodium phosphate, which offers certain advantages over the others; but it is not suitable for high pH values, which would require excessive quantities. STEAM ACCUMULATORS

The continual variations in momentary steam consumption, and the consequent fluctuations in steam pressure, have led certain factories to install accumulators. These are vessels serving to regulate the output of steam. Contrary to the impression which the name would give, it is not actually steam which they store, but water; but it is indeed steam which they deliver. Principle

A steam accumulator is a pressure vessel, a large cylindrical drum resembling a boiler drum, filled with water, and communicating on one side with the boilers, on the other side with the low pressure or exhaust steam piping. The connections are made by means of valves and regulators located in such a way as to permit circulation of steam only in the sense which we shall describe. We shall designate by: P = normal boiler pressure P' = pressure slightly lower than P, to which it is arranged that steam may drop in the boilers p = normal pressure at which steam is used for manufacture; this is generally the exhaust steam pressure p' = pressure slightly higher than p p" = pressure slightly lower than p. The regulators and valves are arranged so that: (a) When the h.p. steam is at a pressure between P and P\ the accumulator allows steam to enter from the boilers, but does not deliver any steam. It thus becomes charged.

740

39

STEAM PRODUCTION

(b) When the pressure falls below P' the connection with the boilers is closed. (c) When the exhaust steam pressure falls below /?", the accumulator delivers steam to the low-pressure line, and thus discharges. (d) When the pressure in the exhaust system rises above p\ the connection between it and the accumulator is closed. The two pressures p' and p" are chosen relative to p in such a way as to avoid too frequent opening and closing of the valves and to provide a margin of stable operation. The principle of the accumulator is simple. It stores water at the temperature of saturated steam corresponding to the boiler pressure. If the boilers produce superheated steam, it will become saturated when it is forced into the water in the accumulator. In order to obtain an effective and silent mixing of the steam with the water, the steam is introduced into the vessel by conical or bell-shaped nozzles. When a drop in pressure occurs in the accumulator, a portion of the water evaporates practically instantaneously: it is the sensible heat of the whole mass of stored water which is transformed into latent heat and so furnishes the heat of vaporization of the portion evaporated. The accumulator should obviously be very carefully insulated. Calculations for a steam accumulator

Let: P = normal boiler pressure P' = pressure slightly lower than P to which the boiler pressure is allowed to drop p = back pressure, or exhaust pressure (in the calculation, we shall assume for simplification:

P=P'=P") R' r T' t x q

= latent heat of steam at the pressure P' — latent heat of steam at pressure p = temperature of saturated vapour at pressure P' = temperature of saturated vapour at the pressure p = any temperature between T' and / = the quantity of steam furnished by one pound of water, when its temperature drops from T to t Q = the quantity of steam furnished by the accumulator, when its temperature drops from T to / V = the volume of water contained by the accumulator V = gross total interior volume of the accumulator. When the accumulator delivers steam, and when the temperature of the water which it contains drops from x to {x — ax), each pound of water will supply a quantity of steam aq such that: aq [1,093 — 0.7(JC — 32)] = — 1.013d* (650) Whence: \Q\2> d g

=-

(651)

1,115-0.7* " ^

For a temperature drop from T to t, each pound of water will give:

'Ã,ðð™--^ ■*<'.'..-«*

1.013|

t

T

or:

1,115

7TZ~=

~~

Lin v i , i i 3 —

*= L 4 5 1 n w^' = 3 - 3 3 l 0 4

í./÷\ô

(652

>

39

CALCULATIONS FOR A STEAM ACCUMULATOR

741

However, during this time, the accumulator has delivered Q pounds of water, and the total quantity of steam which it can furnish at the moment it passes through temperature x has a value: d ß = F(l — q)aq

(653)

Whence: Q= Kj^,(l--<7)d<7= v(q

^-)

or: ß= ^(l—y)

(654)

Numerical example. In a factory of 59 t.c.h., the steam consumption of which is 1,120 pounds per ton cane, it is desired to allow for periods of half an hour during which steam consumption exceeds the mean consumption by 20% = 224 lb./t.c. We assume that the boilers are sufficient to supply the required steam for the remainder of the time, without fall in pressure. Other conditions are as follows: P' = 313 p.s.i. p = 7 p.s.i. P = 356 p.s.i. r = 426°F t = 232°F R! = 801 B.Th.U./lb. r = 957 B.Th.U./lb. Solution. The mean steam consumption amounts to: 59 x 1,120= 66,080 lb./h and the excess expected during the half hour of peak demand to: 59 x 224 x i= 6,608 lb./h The quantity of steam which each pound of water stored in the accumulator can furnish in dropping from 7" to / will be: q = 3.33 log - ^ = 3.33 (log 957 — log 801) 801 q = 3.33 (2.9809 — 2.9036) = 0.257 We require then an accumulator the contents of which V will be given by: 6,608 = Vx 0.257 (l — °'2*7 ) = 0.224 V V=

6,608 ' , = 29,500 lb. water 0.224

At 426°F, the specific volume of water is 0.01904 cu.ft./lb., hence: V= 29,500 x 0.01904= 562 cu.ft. The water contained by the accumulator represents only about 80X of its total volume, on

742

STEAM PRODUCTION

39

account of the necessity of allowing a margin to avoid entrainment due to possible priming. The total volume of the vessel will then be: 0.8

' '

say a cylinder of approximately 6 ft. in diameter and 25 ft. in length designed for a pressure of 356 p.s.i. STEAM-REDUCING VALVES

In a well balanced sugar factory, the quantity of steam necessary for manufacture is much greater than that required for production of power. Since all the steam is generally produced at high pressure, there is then normally a large excess of high-pressure steam to be passed into the low-pressure system. This operation is effected by means of pressure regulators, which are actuated by the lowpressure-steam system: as soon as the pressure, which is generally the exhaust or back pressure, falls below a fixed value, the regulator opens and admits steam (Fig. 387). These regulators generally consist of a diaphragm the underside of which is subjected to the pressure to be regulated. The pressure exerted by the steam on this diaphragm is balanced

Pressure line

Pressure regulator

Fig. 387. Pressure reducer (Quint & Flamant),

39

743

TEMPERATURE OF THE REDUCED STEAM

by an adjustable counterweight. When this counterweight overcomes the steam pressure, the lever descends under the action of the weight and thus actuates a needle valve. This admits water under pressure to an actuating piston, which controls the opening of the valve admitting high-pressure steam. In order that this valve shall work at a normal opening, it is essential that it should be of a diameter not only much smaller than that of the downstream pipe for the expanded vapour, but also much smaller than that of the upstream high-pressure-steam line. Otherwise it would operate in an almost closed position, and would rapidly wear both the valve and the seat. Temperature of the reduced steam

The final state of the reduced steam is given by the MoUier diagram (Fig. 367). The horizontal lines in this diagram correspond to changes without external work, at constant total heat, and this is the case with pressure regulators of reducing valve type. If for example superheated steam at 342 p.s.i.g. and 662°F is to be reduced to 7 p.s.i.g., the point is taken on the diagram corresponding to 357 p.s.i.a. and 662°F, and the horizontal line through this point will be followed to the isobar 22 p.s.i.a. The isotherm through this point is 626°F, the temperature sought for the expanded vapour. DE-SUPERHEATING

The high-pressure steam is superheated in all electrified factories possessing one or more turbo-alternator sets. Temperature regulator

if isersjj^jj

Desuperheated I. p. steam

L^^^

Desuperheating water

Desuperheating -w . .chamber .

2D ^£

¿Ú

I (Upstream temperature II Actuating water " I pressure

Fig. 388. Pressure reducer - de-superheater (Quint & Flamant).

744

STEAM PRODUCTION

39

We have just seen that expansion through a valve reduces the steam temperature only slightly, since the isothermal lines on the Mollier diagram are almost horizontal. The reduced steam is therefore superheated. If the quantity of make-up steam necessary for manufacture is relatively small, this superheat will have no great disadvantage: it will serve mainly to reduce the moisture content of the exhaust steam, or to give it a slight superheat. We have already seen (p. 349) that this superheat presents no great disadvantage, and would even be advantageous, provided it does not exceed 50-90°F. However, the make-up is generally much too great for the superheat to remain as low as this, and it is thus necessary to de-superheat the reduced steam. Hence the necessity, in this case, to follow the pressure regulator by a de-superheater (Fig. 388). Calculations for de-superheater

Suppose it is required to de-superheat the steam furnished by the regulator above, leaving only 90°F of superheat. At the entry to the de-superheater, the steam at 7 p.s.i.g. and 626°F contains 1,346 B.Th.U./lb., according to the diagram or the table, (7 p.s.i.g. = approximately 21.7 p.s.i.a.). At this pressure, the temperature of saturated steam is approximately 232°F. If we wish to leave 90°F of superheat, it will then be necessary to reduce the steam to 322°F. The diagram shows that vapour at 21.7 p.s.i.a. and 322°F contains 1,203 B.Th.U./lb. Per pound of vapour passing through the de-superheater, it is then necessary to remove: 1,346 — 1,203 = 143 B.Th.U. For this purpose, we introduce into the de-superheater water from the boiler, which is at 341 p.s.i.g. and 433°F, and possesses a latent heat of vaporization: r = 794 B.Th.U./lb. As it flashes into vapour at the reduced pressure, it will absorb heat. It will be necessary then to introduce: 143

794

= 0.1801b. water

per pound of steam to be de-superheated. If the quantity of vapour expanded per hour is 11,000 lb., this will require: 11,000 x 0.180 = 2,000 lb. water/h and we shall have 13,000 lb./h of steam which is partially de-superheated. The de-superheating water is finely atomized in the de-superheater, the successive opening of the several atomizers in the battery being actuated by a long thermostat element immersed in the outlet pipe carrying the de-superheated steam. TYPES OF BOILER

There exist in the sugar factory four principal types of steam boiler: (a) The "elephant" fire-tube boiler (Fig. 389). (b) The semi-tubular or "multi-tubular" boiler (Fig. 390). (c) The water-tube boiler with headers and straight tubes of moderate slope (Fig. 385).

39

745

TYPES OF BOILER

fe^ft

ίTδ

i3¥h

rni

Fig. 389. "Elephant" boiler with fire-tubes.

Fig. 390. Semi-tubular boiler.

(d) The water-tube boiler with bent tubes, steeply inclined and connected directly to the boiler drums (Fig. 378). The "elephant" boiler is now seldom found. It is moreover very similar to the semitubular, but has an appreciably greater water capacity. Again, the two modern water-tube types do not differ greatly. The type with straight tubes presents the advantage of being able to arrange the tubes in staggered formation (Babcock and Wilcox or Cail-Steinmuller), hence a better mixing of the combustion gases, better utili­ zation of the heating surface of the tubes, and less deposit of soot on the front and back faces of the tubes. The type with bent tubes, on the other hand, promotes more rapid circulation of the water, on account of the steeper inclination of the tubes (50-80°, in place of about 15°). Finally, a similar evaporation rate is obtained for both types (cf. Table 172). A disadvantage of the bent-tube type is the replacement of tubes, which is difficult. This difficulty is avoided by arranging the tubes with longitudinal intervals alternatively wide and narrow, so that the tube to be replaced can be passed through the wide space. To offset this, the type with headers and straight tubes, while it allows of easy replacement of tubes, requires for this operation that the front of the boiler should have space available for a distance equal to the length of the tubes, and bagasse conveyors and chutes must be arranged accordingly.

746

STEAM PRODUCTION

39

Water-storage capacity

Certain technicians, especially in Java, attribute some importance to the volume of water contained in the boiler. This volume is at the boiling point under the pressure concerned, or very close to it, and it is considered that in case of a drop in pressure, due to a sudden demand for steam for manufacture, the body of water thus stored would constitute a reserve which would act as a steam accumulator and would permit the boiler to cope with part of this peak demand. It is known that a drop in pressure causes the instantaneous vaporization of the water which is present at a temperature above that of saturated steam corresponding to the reduced pressure. However, the principle of conservation of energy shows that the quantity of heat absorbed by the weight of water vaporized is equal to the sensible heat lost by the body of water. Now, the latent heat is considerably higher, and the reserve steam thus stored is consequently insignificant. We shall take for example a boiler of 2,690 sq.ft. containing 26,460 lb. water, the evaporation rate of which is 2.5 lb./sq.ft./h and the normal pressure 114 p.s.i.g. (t = 347°F). If its pressure falls to 100 p.s.i.g. (t = 338°F), the sensible heat liberated will be: 26,460(347 — 338) = 238,140 B.Th.U. and this will evaporate:

238,140 ——-— = w271 1U lb. steam 877

(877 B.Th.U./lb. = average latent heat between 100 and 114 p.s.L). Now, this boiler normally produces: 2,690 x 2.5 = 6,725 lb./h = 112 lb./min of steam, and the 271 lb. obtained therefore correspond only to that produced in: 271 · 25 ,c sec —— = 2. min 112 of normal operation, and this for a drop in pressure of 14 p.s.i.! Furthermore, it is very difficult to heat and evaporate this large quantity of water, and this renders the multi-tubular boiler sluggish and slow to adapt itself to sudden demands for steam, whereas a boiler with rapid evaporation responds more readily to regulation of the draught. Provided of course that the heating surface installed is sufficient, we see no disadvantage in having all the boilers of the factory in the form of water tube units of high evaporation rate; especially as the necessary heating surface in this case will be substantially lower than that which would be required with semi-tubular boilers.

Series of b o i l e r sizes

We give in Tables 182 and 183 the series of boilers fabricated by the two principal French manufacturers of sugar mill equipment. Boiler tubes. We give in Table 184 the standard dimensions of boiler tubes (drawn weldless tubes).

39

747

JOINTS IN BOILER DRUMS

TABLE 182 CAIL-STEINMULLER BOILERS WITH HEADERS. PRESSURES UP TO

Heating surface (m 2 )

Number of drums

Step-grates Number and width (mm)

1 1 1 1 1 1 1 1 2 2 2 2 2 2

160 180 200 224 250 280 315 355 400 450 500 560

630 710

32 kg/cm 2

1 1 2 2 2 2 2 2 2

2

2 2 2 2

x x x x x x x x x x x x x x

Horseshoe furnaces Number and width (mm)

1,300 1,500 1,000 1,100 1,100 1,200 1,200 1,300 1,500 1,600 1,800 2,000 2,300 2,500

1

x 1,000

1

x 1,200

1

x 1,500

1

x 2,000

2

x 1,200

2

x 1,500

2 3

x 2,000 x 1,500

TABLE 183 FIVES-STIRLING BOILERS WITH 3 STEAM DRUMS, TYPE VS.

Rows of tubes

Heating surface (mV

14 17 19 21 24 29 33 38 43 48

294 357 399 441 504 609 693 798 903 1,008

Step-grates "Stella" or Width between walls "Quartier frangais" type Horseshoe furnaces Area in m2 (m) Number and width (mm) 2 2 2 2 2 2 2 3 3 3

2.190 2.700 2.970 3.310 3.710 4.530 5.140 5.850 6.630 7.370

X X X X X X X X X X

1 1 2 2 2 2 2 2 3 3

1,000 1,200 1,400 1,500 1,700 2,000 2,300 1,800 2,000 2,300

= 2.22 = 2.50 = 2.76 = 3.12 = 3.60 = 4.06 = 4.44 = 5.94 = 6.09 = 6.66

TABLE 184 DIMENSIONS OF BOILER TUBES

Outside diameter mm

in.

70 76 83 89 95 102 108

2* 3 3i

H 3f 4 4*

Weight

Inside diameter (mm)

Thickness (mm)

kg/m

Ib.lft.

64 70 76.5 82.5 88 94.5 100.5

3 3 3.25 3.25 3.50 3.75 3.75

4.960 5.400 6.390 6.870 7.900 9.090 9.640

3.333 3.629 4.294 4.617 5.309 6.108 6.478

CONSTRUCTION OF BOILERS Joints in boiler drums Tromp (p. 283) rightly draws attention to the comparison to be made between the longitudinal and transverse stresses involved in the plates and joints of boiler drums under pressure.

748

39

STEAM PRODUCTION

(a) Circumferential joints. Consider a vertical section of the drum (Fig. 391). The section is subject to a bursting force of: Fc=

(655)

nRZp

Fc = bursting force, in pounds R = radius of cylinder, in inches p = pressure inside the drum, in p.s.i.

I Fig. 391. Circumferential joints.

This force acts on an area Sc of the metal concerned and the stress in the latter is: fc =

nRP-p Sc

2nR · e

Rp 2e

(656)

e = thickness of the metal, in inches.

Fig. 392. Longitudinal joints.

(b) Longitudinal joints. Imagine the drum cut by an axial plane (Fig. 392). The section is then subjected to a force: F , = IRL-p

(657)

Fi = force acting on a longitudinal section, in lb. L = length of the cylinder, in inches. The stress is then: ft

Fi Ä

IRLp 2Le

Rp

(658)

neglecting the additional strength due to the tube plates (or the ends) which is negligible in the median portion of the cylinder. Conclusion. We see then that the stress acting on the longitudinal joints is double that on the transverse joints:

39

STEAM BALANCE

fi = 2/c

749 (659)

The boiler plates should then be calculated as a function of fi and the longitudinal joints should have a double row of rivets if the circumferential joints have a single row. Plate and tubes nearest the fire. Tromp (p. 289) also recommends the practice of making the bottom plate of fire-tube boilers, which is the first surface encountered by the hot gases, of thicker plate. In the same way, in water-tube boilers, the tubes of the first row (called "coup-de-feu" tubes) could advantageously be made of thicker metal than the other tubes. Location of joints. The circumferential joints are generally made as lap joints: the joint should be so located that the metal plate forming the outer plate of the lap joint faces away from the furnace, rather than facing the flames. Otherwise corrosion at these joints would be very marked. MAINTENANCE OF BOILERS

Slack season. The boilers often deteriorate as much during the slack season as during the crushing season. To avoid this it is necessary, at the end of the season, to empty the boilers carefully, and to check that no pockets of water remain. The interior will be painted with a rust-preventive paint, and the boilers will be closed up again on a very dry day, after having introduced a tray of quicklime or calcium carbide, which will absorb the remaining humidity. Do not forget to remove the tray before the next crushing season! Inspection. Check whether the tubes connecting headers and drums (straight-tube boilers) or the ordinary tubes (bent-tube boilers) are pitted. If pits are found, and if they are deep, the tubes should be replaced. These pits are due to attack of the metal by dissolved oxygen, and are particularly to be feared with acid water. Hence, in this case, check the pH of the water. Check whether the refractory baffles separating the gas passes are in good condition. A short circuit in the path of the hot gases can diminish the efficiency very markedly. Soot blowers. Before using the soot blowers it should be checked that the piping to the blowers has been well drained. Otherwise, the water blown among the tubes will provoke rapid oxida­ tion in all parts which it reaches. STEAM BALANCE

There is frequently need to establish the steam balance of the factory. We shall proceed to establish one but, since, if we sought to work out general formulae, we should end up with very long and complicated equations, we propose to take a concrete case and reduce all figures to a crushing rate of 1 ton per hour. It will be easy to repeat the calculation in any practical case whatever which may arise, by replacing the values adopted by the appropriate figures, and calculating the quantities of steam for the hourly rate concerned. We shall make the calculation for the case of thermo-compression. It will be easy to reduce this to the case of evaporation without thermo-compression, by omitting the portion which deals with the thermo-compressor.

750

39

STEAM PRODUCTION

Steam balance

Numerical calculation. We shall assume the following conditions: Crushing rate of the factory (a) Steam

59 t.c.h.

production

Fibre in cane Weight of bagasse Moisture in bagasse Excess air in the furnaces Temperature of gases at entry to the chimney Boiler pressure Temperature of the superheated live steam Temperature of water in feed tank Coefficients characterising the efficiency of the combustion (c/. p. 691) a = ß = V=

13% 560 lb./t.c. 45% 50% 356°F 356 p.s.i.g. 662°F 194°F 0.99 0.95 0.95

Factory completely electrified. Compound clarification. Quadruple effect evaporation with thermo-compressor. (b) Steam consumption Weight of mixed juice per ton of cane 2,240 lb./t.c. 12 Brix of clarified juice 60 Brix of syrup 2,464 lb./t.c. Weight of primary juice 1,232 lb./t.c. Weight of secondary juice 122°F Temperature of primary juice after mixing 122°F Temperature of secondary juice after mixing Vapour bled from third effect 45 lb./t.c. Pz = Vapour bled from second effect 67 lb./t.c. P2 = Vapour bled from first effect 90 lb./t.c. Pi = Quantity of vapour aspirated by the thermo-compressor y = 448 lb./t.c. Temperature of primary and secondary juices at the outlet from vapour heating 200°F Temperature of heating of primary and secondary juices by exhaust steam 221°F Back-pressure 8 p.s.i.g. Vacuum 25 in. Calculation Input. The N.C.V. of the bagasse will be (eqn. 570): N.C.V. = 7,650 — (8,730 x 0.45) = 3,722 B.Th.U./lb. The loss of sensible heat in the flue gases will be (eqn. 594):
39

751

CALCULATION

1,343 — 162 = 1,181 B.Th.U./lb. One pound of bagasse will then give: 2,865 = 2.42 lb. steam at 356 p.s.i. and 662°F 1,181 and a ton of cane can supply, if necessary: 560 x 2.42 = 1,355 lb. live steam Output. We shall first calculate the steam consumption of the engines, then we shall consider the steam consumed in manufacture. (a) Engines. The power consumption of the engines in a sugar factory is of the order given in Table 185. TABLE 185 MEAN TOTAL POWER CONSUMPTION (i.h.p./t.C.h.)

(a) Mills (b) Other plant: Electrified factories Non-electrified factories Total: Electrified factories Non-electrified factories, approx.

12-16 8-16 10-16 approx. 20-32, average 25 22-32, average 27

If the factory is provided with turbo-alternators operating between 327 p.s.i. (662°F) and 8.5 p.s.i., for example, and if they are operating between 3/4 and full load, they will consume approximately 22-24 lb. steam/kWh, or say 17-18 lb./h.p.h. Suppose the factory consumes a total of 20 i.h.p./t.c.h. At the switchboard of the power house this corresponds to: 20 —— —— = 25 h.p./t.c.h. 0.85 X 0.97 '

0.85 = average efficiency of the electric motors 0.97 = average efficiency of transmission lines. With a steam consumption of 18 lb./h.p.h, this will require at the turbines: 25 x 18 = 450 lb. steam/t.c. However, this steam is not lost: it is recovered almost entirely in the form of exhaust steam. We shall take only: Loss at turbines 5% or: 450 x 0.05 = 22 lb. Leaks from the piping 11 lb./t.c. or: 11 lb. 33 1b. and we shall recover in the exhaust: M = 450 —33 = 4171b./t.c. (b) Manufacture. (1) Multiple effects. The quantity of water evaporated by the multiple effects is:

752

39

STEAM PRODUCTION

2 , 2 4 è ( À - ^ ) = 1,792 lb./t.c. and the distribution of this evaporation between the vessels will be, designating by x the evaporation in the last vessel: 4th 3rd 2nd 1st

vessel vessel vessel vessel

x x + 45 x + 45 + 67 x + 45 + 6 7 + 9 0 + 448 Ax + 135 + 134 + 90 + 448= 1,792

hence: 1,792 — 807

♦= 246

The steam consumption of the first vessel will then be: (a) Steam 327 p.s.i.: 448 q = — - = 224 lb. 2 since we have already seen that the value of//, in these conditions, was of the order of 2 (see p. 406).

x»P2»^

q+y

q'

Fig. 393. Diagram of operation of first effect.

(b) Steam at 7.5 p.s.i. (Fig. 393): q' = x + 45 + 67 + 90 + 448 — 448 — 224 = 224 lb. Or a total of: q + q' = 224 + 224 = 448 lb./t.c. (2) Vacuum pans. The weight of syrup is: S = 2,240 — 1,792 = 448 lb./t.c.

39

753

CALCULATION

and the steam consumption at the pans (cf. p. 497): 60\ C = 448 (\ — — \ 1.5 = 252 lb./t.c assuming that the massecuites are heavied up to 96° brix at least. (3) Juice heaters. The finishing heaters heated with exhaust steam, consume: ^v =

2,464 x 0.9(221 — 200) ——; —— 955 x 0.95

h

1,232 x 0.9(221 — 200) 955 x 0.95

or: R = 77 lb./t.c. The total steam consumption for manufacture is therefore: Evaporation q + q' 448 lb./t.c. Pans C 252 lb./t.c. Heaters R 77 lb./t.c. 777~lb./tc7 It is necessary to add the following figures for steam used: Steaming at the centrifugals = 20 lb./100 lb. sugar = approx. 54 lb./t.c. 54 lb./t.c. Michaelis, soot blowers, washing atfilters,heating of molasses, etc. 56 lb./t.c. Losses by condensation, leakage, steam traps, valves, etc. 90 lb./t.c. 977 lb./t.c. Since we have only 417 lb. of exhaust steam, we shall have to utilize steam between 60 and 85 p.s.i. for centrifugals, michaelis, etc., or to expand to the exhaust system: 977 — 417 = 560 lb. high-pressure steam The total quantity of live steam to be produced will then be: 450 + 560= 1,010 lb. steam A factory equipped and operating in the foregoing conditions will thus consume about 1,000 lb. of steam/t.c. If it had: quintuple effects instead of quadruple, or an evaporator working under pressure, or vacuum pans boiled by bled vapour, it would be possible to reduce consumption further. This factory could then save: 1,355-1,000 1,355

26%

of its bagasse. This economy, as we have just indicated, could easily be surpassed. Many large Cuban factories utilize in their boilers only 60% of the bagasse which they produce (F.A.S., (Jan. 1939) p. 47). Furthermore, Jenkins estimates in Queensland (F.A.S., (Oct. 1939) p. 31) that a typical

754

39

STEAM PRODUCTION

average Australian factory of that date, crushing about 60 t.c.h., and fully electrified, with high-pressure boilers, should be able to produce 3,300 kW over and above the requirements of the factory. This would be surplus power, made available by the use of high-pressure boilers of high efficiency, and cannot readily be expressed in terms of an equivalent saving in bagasse. However, it shows that great savings are possible with improved efficiency and improved heat balance. Steam consumption of prime movers

(a) Electrified factories. The steam consumption of turbines may be calculated as shown on p. 786. Example. A turbo-alternator taking steam at 313 p.s.i.g. and 662°F, and exhausting at 8.5 p.s.i.g. will have a steam consumption such as in Table 186. TABLE 186 STEAM CONSUMPTION OF A TURBO-SET (313 - 8.5 p.S.i.g.)

Ib./kWh lb./h.p.h.

At \ load

At I load

At full load

25.6 19.1

23.8 17.8

22.5 16.8

(b) Non-electrified factories. The steam consumption of steam engines varies greatly accord­ ing to the condition and pressure of the steam. It may be calculated from the graph of Fig. 409. Examples are given in Table 187. TABLE 187 EXAMPLES OF STEAM CONSUMPTION OF ENGINES ( l b . / h . p . h . )

Type of engine

Admission

Exhaust

Consumption

Corliss Drop-valve

114 p.s.i.g. satd. 213 p.s.i.g. 482°F

1 p.s.i.g. 7 p.s.i.g.

31.3 21.2

The highest steam consumptions are shown by duplex direct-action steam pumps, which require at least 80 and often as much as 100 and 110 lb./h.p.h. In practice, the calculation of the steam consumption of the engines should be made by considering the engines one by one, each with its appropriate steam consumption. Losses. As indicated in the numerical calculation above, two types of loss have to be taken into account: (a) Losses in the engines by condensation, leakages by valves or safety valves, stuffing boxes, etc. We may take the figures given in Table 188. TABLE 188 STEAM LOSSES AT ENGINES

Steam turbines Engines with slight superheat Engines using saturated steam

1/20 of consumption 1/10 of consumption 1/6 of consumption

This loss occurs between admission and exhaust of steam.

39

STEAM CONSUMPTION OF PRIME MOVERS

755

(b) Piping. The steam pipes lose steam by leaks at the joints, by condensation and at steam traps, etc. Noel Deerr (p. 338) estimates that a factory has about 56 sq.ft. of pipe surface per t.c.h., in high-pressure steam piping, and as much in exhaust steam piping. Say a total of 112sq.ft./t.c.h. If these pipes are bare, they would lose: Exhaust steam: 450 B.Th.U./sq.ft./h = 22 lb.steam/t.c. Live steam: 1,350 B.Th.U./sq.ft./h = 66 lb.steam/t.c. Properly lagged, these pipes would lose on an average 110 B.Th.U./sq.ft./h, say about 12,000 B.Th.U./t.c, or approximately 11 lb. steam/t.c. (c) Multiple effect. In estimating that an evaporator of n effects evaporates n pounds of water per pound of steam admitted to the first effect, we have taken into account its losses by radiation and convection. However, this is in the case of a multiple effect which is correctly lagged. If it is badly lagged, we have seen elsewhere (cf. p. 392) the loss which must then be calculated. Steam consumption per ton cane. The steam consumption per t.c. varies finally in rather large proportions according as the factory is or is not electrified, according to the degree of steam economy realised in the manufacture, i.e. according as operation is in triple, quadruple, or quintuple effect; according to the extent of vapour bleeding used; and also whether or not thermo-compression is employed, etc. The highest steam consumptions are reached in the case where white sugar is made, with evaporation in triple effect, and extensive use of direct acting steam pumps. Steam consumption may then reach 70-80% on cane. Factories manufacturing raw sugar and not utilizing electricity, even for pumps, will show a steam consumption between 60 and 70% on cane. Modern factories with an electric generating set supplying all small units, and particularly with motor-driven pumps, and with careful attention to use of steam, will reduce steam con­ sumption to between 50 and 60 %on cane. With quintuple effect and a full sequence of vapour bleeding, with thermo-compression or evaporation under pressure, and with superheated steam at high pressure, consumption may go as low as 40-50% on cane. In Hawaii (I.S.J., (1933) p. 201), for a factory of 80 t.c.h., modern and with steam turbines, electrified except for the mills, with a fibre content of 12.5% on cane, a sugar yield of 12.66%, and boiler efficiency of 66.7%, Mott-Smith estimates: Electric power required for the machines: 10 kWh/t.c. = 13.4 h.p./tx.h. Steam consumption for manufacture: 1,177 lb./t.c. = 52.5% on cane. In Cuba (I.S.J., (1944) p. 236), Jose L. Plana estimates as average consumption: 54.7% on cane. Fluctuations in steam consumption. The variations in momentary requirements of steam in the sugar factory are due to many factors, only one of which is unavoidable in practice: the fluctuations in steam consumption at the vacuum pans. We have seen (p. 496) the tremendous

756

STEAM PRODUCTION

39

variations in evaporation rate from the beginning to the end of a strike. Since the number of pans is generally between 4 and 6, these individual variations have a very marked effect on the general steam consumption of the factory. For the pans, we may estimate approximately a total variation of 50% (in other words, from 50 to 150, if the mean consumption is 100). According to the brix value adopted for the syrup, the quantities of dilution and washing water, etc. added at the pan station, the total steam consumption of this station will vary from 5 to 10% on cane (112-224 lb./t.c.) We may comment that it is preferable to express this variation as a percentage on cane rather than to reckon it as a percentage on the total steam consumption, since, in practice, only the variation in steam consumption at the pans is concerned. As a percentage of the total steam consumption, this variation will generally represent between ± 1 0 and ±20%. Total power. Similarly the average total power consumed in a sugar factory is rather variable. Noel Deerr (p. 330) estimates: 18 i.h.p./t.c.h. Tromp (I.S.J., (1938) p. 177): 25 i.h.p./t.c.h. but the latter author comments that it is necessary to calculate this value in each case, since it varies greatly in different installations. The present author has given (Table 185): 20-32 i.h.p./t.c.h. Influence of electrification. It is appropriate to comment that electrification itself effects no economy in steam. It involves a triple transformation of energy (thermal energy into mechanical, mechanical energy into electric, electric energy into mechanical energy) which affects the steam consumption in each case, and thus loses the benefit of the high thermodynamic efficiency of the steam turbine. For example, the calculation of steam balance (p. 750) carried out for a factory with steamdriven mills, steam engine for central vacuum system, and a generating set for pumps and small isolated units only, where the steam engines were drop-valve engines working at 213 p.s.i., with steam slightly superheated to 482°F, gave a total steam consumption hardly higher than for the electrified factory: 1,053 lb./t.c. The true steam economy effected by electrification lies in the replacement of direct-acting pumps and small isolated units. This is the most important improvement and the first one which should be carried out. General arrangement of t h e steam circuit

The simplest steam cycle, in a modern factory, consists of producing the steam in high-pressure boilers and expanding it to the pressure necessary for manufacture in turbo-alternators (T.A.) which produce all the electric power necessary for the factory (Fig. 394). Since the quantity of exhaust steam supplied by these turbo sets is insufficient, the highpressure circuit is connected to the low-pressure circuit through a steam reducer-desuperheater (Desup.) which supplies the make-up steam necessary. Finally the high-pressure circuit will be completed by a connection to the thermo-compressor (T.C.). In the case where the mills are not electrified, there is a medium-pressure circuit supplying the steam engines (Fig. 395). We thus have a choice between two arrangements for the turbo sets: they may be installed to work on the pressure drop h.p./m.p., or alternatively the drop h.p./l.p. (as in Fig. 395). These two solutions are practically equivalent from the thermodynamic point of view, but the

39

757

CHOICE OF STEAM PRESSURE

second is by far preferable, since it allows for the possibility of eventual total electrification. A reducer-desuperheater should be installed between the h.p. and m.p. circuits, but a simple reducer will be sufficient between the m.p. and l.p. circuits. H.R B o i l e r s H.R Steam

rih

yi

.od

1ÄÁ

Desup.

mi

b_JX^"

Condensates Fig. 394. Steam circuit No.|l.

M.R Steam

L.R Steam

, π—

ßã| i 1 Lfl

ft r^A.

Ik

r—^LL·^ <\—T\ »

t-ta O I I

ruh

Condensates

Fig. 395. Steam circuit No. 2. Choice of s t e a m pressure

In the sugar factory we have three principal steam pressures: (a) the h.p. steam for the turbo sets

1 BITX.

758

39

STEAM PRODUCTION

(b) the m.p. steam for reciprocating engines (c) the l.p. steam for manufacture. (a) Choice of high pressure. To achieve a suitably economical operation of the steam turbines, it is necessary to have a high pressure between 225 and 575 p.s.i. Below 225 p.s.i., the steam consumption of the turbines increases rapidly. Also the domain of the reciprocating engine readily extends to 210 p.s.i. If this limit were not exceeded, there would be every advantage, as much from the point of view of first cost of the installation as from that of steam consumption, in retaining the drop-valve engine, operating at 210 p.s.i., with a slight superheat to 482°F. Above 575 p.s.i., the expense of the installation, due to the cost of the boilers, which increases rapidly with pressure, quickly reduces the value of the steam economy obtained. Conditions will generally be kept within the ranges shown in Table 189. TABLE 189 OPTIMUM CONDITIONS FOR H.P. STEAM IN THE SUGAR FACTORY

p.s.i.

°F

250 285 350 425

575-675 600-700 650-750 700-800

Comparing these four values, the advantages and disadvantages compensate each other and, in our opinion, there is not much to choose between them. (b) Choice of medium pressure. The medium pressure is more especially of interest with regard to reciprocating engines. In order to obtain economical operation with these, it is necessary to use a pressure between 85 and 200 p.s.i. The highest pressures (170-200 p.s.i.) are of interest for factories having large steam engines, especially in cases where the mills are not electrically driven. Particularly economical conditions of operation will then be obtained. A pressure lower than 85 p.s.i. should not be used, on account of the rapid increase in steam consumption of reciprocating engines below this figure. (c) Choice of low pressure. The low pressure is generally at the same time the upper limit of pressure for manufacture as well as the back-pressure for the engines. To avoid development of colour in the juices and to avoid destruction of sucrose, we have seen (p. 357) that it is advisable not to exceed 20 p.s.i.g. in an ordinary multiple effect, and 28 p.s.i.g. in a Kestner. If pressure evaporation is to be used, it is beneficial to approach these limits. Otherwise, pressure should be kept between 7 and 14 p.s.i.g. A value of 9-11 p.s.i.g. is often chosen. Sometimes the low-pressure system is divided into two: a higher pressure for coil pans (20 p.s.i.g., for example), and a lower pressure for juice heaters and multiple effects (7 p.s.i.g., for example). This has the disadvantage of complicating the piping system for steam distribution.