CA2062444C - Turbine exhaust arrangement for improved efficiency - Google Patents

Abstract

Steam exhaust outlets of a low pressure steam turbine are fitted with a divider plate to separate exhaust steam into isolated flow paths in fluid communication with a condenser. Separation of the flow paths is maintained through the condenser so that heat rate is improved by lower average back pressure and higher temperature condensate exiting the condenser. In a double flow turbine, a further divider plate separates steam from one exhaust outlet from that of the other exhaust outlet thereby creating four steam flow paths to the condenser.

Description

~002~~~
56,307 TURF3INE EX~iAUST ARRANGEMENT FOR IMPROVED EFFICIENCY
This invention relates to steam turbine power generating system and, more particularly, to a multiple zoned low pressure turbine exhaust.
BACKGROUND OF THE INVENTION
Environmental protection and limited water availability have necessitated the adoption of larger temperature rises in the condensers of utility power plants. There has been increased use of cooling ponds and wet cooling towers (both natural and mechanical draft) and in some instances, dry cooling. An increase in turbine exhaust pressure has accompanied the adoption of these supplementary cooling systems.
This not anly reduces the plant efficiency but also places additional demands on the cooling system.
In the United States, dry cooling has been limited to one utility with an initial application on a 20 MW turbine and a subsequent 360 MW unit. Both of the applications were with air cooled condensers.
South Africa has built six 665 MW units with air cooled condensers, with three more under construction.
In other applications, indirect air cooling is used in which exhaust steam is channeled through a dry tower, usually a natural draft design. South Africa has built six 668 MW units using indirect air cooling. A

56,307 number of smaller size indirect designs were built in England, Germany, Hungary, U.S.S.R., Iran, Brazil, Turkey, and South Africa. These plants employed either spray or surface condensers.
In at least one dry cooling study of a nuclear power plant, it was established that the use of multipressure or zoned condensers improved the plant economics. Moreover, the use of different size last row blades in each low pressure (LP) element (tandem compound six flow exhaust) further improved to the economics. In this instance, the lowest pressure LP element had the largest exhaust annular area with decreasing annulus areas in the higher pressure LP elements. The economic benefit and improvement in turbine performance increases with the number of multipressure levels or zones. Under conventional practice the number of zones correspond to the number of LP
elements. However, U.S. Patent No. 4,557,113 assigned to the assignee of the present invention, discloses a turbine system having separate zones in each half of a double flow LP element with downward exhaust. From the disclosed system, it is possible to obtain two zones with a single LP element, four zones with two double flow LP elements and six zones with three double flow LP elements.
U.S. Patent No. 4,958,985, issued September 25, 1990, assigned to the assignee of the present invention, proposes to vary the gaugings on the last stage (rotating and stationary blades) by reorientating the blade foil while keeping the 20624~~

3 56,307 rotating blade profile the same to optimize the performance in the various zones of the LP turbines and to use different size last row blades in each half of a double flow LP element to achieve more optimum performance if the differences in exhaust pressure were large enough in the various zones. Turbines have been built in which the individual LP turbines of a specific unit have different length last row blades.
With dry finned tubes of air cooled condensers, the temperature of the cooling air rises substantially. The gradient for the transfer of heat is the difference in temperature of the air and the condensing steam. The tubes of the dry finned sections must be comparatively shallow, which means that usually not more than three to six rows of tubes are crossed in succession by the air passing over them. The successive increase in air temperature will produce a successively higher steam condenser pressure in each row, although this is sometimes ameliorated by varying the fin spacing of each row.
The different condensing pressures must equalize in the headers so that: (1) the condensate from all tubes will drain completely; and (2) the air in all tubes will be separated and evacuated. In one exemplary system, the air cooled condenser operates at approximately 15°C lower saturation temperature owing to pressure loss in the steam duct (connecting the turbine exhaust flange and the air cooled condenser) and the condensing elements.
Because of the tendency of the air cooled condenser to produce successively higher steam 206~~~~

4 56,307 condenser pressures in each row of tubes (as the air successively increases in temperature in passing through the air cooled condenser), it is especially suited to multi-pressure or zoning operation. In this case, the lowest pressure zone would occur in the first row of tubes and the highest pressure zone in the last row of tubes.
In May, 1979 the assignee of the present invention was .granted a patent on a zoned or multipressure system for a "Dry Cooling Plant System"
(U.S. Patent No. 4,156,349). In this instance, the LP
steam turbines exhausted to steam condensers-ammonia reboilers. The ammonia evaporated, was ducted to the air cooling tower where it condensed, and returned to the condenser-reboiler. In this instance, the ammonia from one condenser-reboiler went to the cooling tower tubes that received the inlet cooling air. The ammonia from the other condenser-reboiler want to the cooling tower tubes that received hot air leaving the first group of tubes. So, the steam turbine operated with two pressure zones on a dry cooled plant.
It was noted that increasing the number of condensing zones or pressure levels improves cycle performance and economics of indirect air cooled plants because of the large cooling range (large temperature rise) typical of dry cooled systems. In the case of air cooled condensers, there is an inherent tendency for each row of condenser tubes to operate at successively higher pressure as the air passes through the condenser system.

56,307 Moreover, many wet cooling systems with conventional steam condensers have large temperature rises and are especially suited to multi-pressure or zoned condenser applications. As noted earlier, 5 increasing the number of pressure zones improves performance on both indirect air cooled and wet cooling tower plants. The problem is that the number of zones is limited to the number of turbine exhaust flows. The aforementioned U.S. Patent No. 4,557,113 discloses a system in which two zones are obtainable on a double flow LP element, i.e., a condenser is divided into two zones with exhaust from one end of the turbine coupled to one of the zones and exhaust from the other end of the turbine coupled to the other of the zones. The advantages of this two zone system suggest that more zones might provide additional improvement. However, it has been believed that the number of zones is limited to the number of available turbine exhausts.
If it .were possible to obtain more than two exhaust pressure zones on a double flow LP element or multiple pressure zones on a single flow LP element, additional improvements could be obtained. Table I
illustrates the pressure levels and increase in available energy from use of a low pressure zone in a two zone single flow LP element over single pressure operation, both designs having a 20.o°C temperature rise of the cooling water. To is the incoming cooling water temperature. TZ is the cooling water outlet from the second zone of a mufti-pressure, two zone 6 56,307 condenser. PZ and P~ are the saturation (condensing) pressures corresponding to Tz and T~, respectively.
The portion of the exhaust steam (approximately half) that exhausts to the low pressure zone has between 15.5 and 16.4 Kcal/Kg more available energy than the steam in the single pressure design. The increase in available energy is dependent upon the initial condenser temperature which was varied between 30°C and 56.7°C,. corresponding to a range of water temperatures leaving a cooling tower.
In Table II, a single pressure and a four pressure Eoned condenser are compared. In this case, To is the initial cooling water temperature with T4 being the water temperature leaving the last zone. T~, TZ, and T3 are the water temperatures leaving the other zones. P~, PZ, P3, and P4 are the condensing pressures in the various zones. P4 is also the condensing pressure of an.unzoned or single pressure design.
There are corresponding increases in available energy of the steam expanding in the various zones above the available energy of the single pressure design.
Tables III and IV relate to comparisons between one zone and two zone and one zone and faur zone designs, respectively, for a temperature rise of 13.3°C. The temperature rises in dry cooling systems would probably approach the 20.0°C level while the 13.3°C to 20.0°C range would be more typical of natural draft wet cooling towers. Fossil units with natural draft wet cooling towers would tend to be in the lower half of the 13.3°C to 20.0°C range while nuclear units would be in the upper half of this range. Fossil 7 56,307 applications with wet type mechanical draft cooling towers generally have temperature rises between 8.3°C
and 13.9°C while nuclear plants with mechanical draft towers would usually have temperature rises between 13.3°C and 16.7°C. In areas with low humidity, mechanical draft wet towers have been built with temperature rises of 16.7°C to 20.0°C.
Tables V and VI identify the steam temperatures and pressures in. the various zones for single, two, and four zone combinations with 13.3°C and 20.0°C
temperature rises and given conditions in the maximum pressure zone.
Calculations were made with the standard hood loss on the turbine configuration utilized to evaluate zoning as well as with 0.56, 1.11, and 1.67 Kcal/Kg hood loss increases. Table VII compares single or unzoned performance with two zone performance, and 13.3°C temperature rises. The two zone performance is presented with 0, 0.56, 1.11, and 1.67 Kcal/Kg increases in hood loss. Table VIII presents comparable data but with a 20.0°C temperature rise.
Both of these comparisons relate to a single flow LP section. Even with a 1.67 Kcal/Kg increase in hood loss, there is still an output increase with two zones. The increase in output is larger with a 20.0°C
rise than with a 13.3°C rise.
If the turbine had a double flow LF element, it could be built with two zones as shown in the aforementioned U.S. Patent No. 4,577,113. For that design, there would be no increase in hood loss for a given exhaust volumetric flow.

20~~~4~

8 56,307 It is obvious that there is a significant increase in available energy with multi-pressure. For the case of two versus one zone, the increase is between 7.72 and 8.22 Kcal/Kg for a 20.0°C rise and 5.33 to 5.61 Kcal/Kg for a 13.3°C rise, based on the total exhaust flow (half of value shown on Tables I , and III). In the case of four versus one zone, the increase is between 11.6 and 12.3 Kcal/Kg for a 20.0°C
rise and between 8.06 and 8.39 Kcal/Kg for a 13.3°C
rise, based on the total exhaust flow (half of value shown on Tables II and IV).
Tables V and VI identify the pressures associated with the various zoning canfigurations for various maximum condensing temperatures and condenser temperature rises of 13.3°C and 20.0°C.
SZJN~IARY OF THE INVENTION
The above described advantages of a multi-zone turbine system are attained in one form of the present inventian by placing a divider plate along the vertical axis (axi~xl orientation) of a turbine exhaust to create two pressure zones in each end of a downflow or upflow exhaust. In the case of side exhausts in both cover and base halves of a turbine, the divider plate may be placed along either the horizontal or vertical center line but maintaining an axial orientation. With an axial exhaust, the divider plate may also be placed along either the vertical or horizontal center line depending upon the condenser orientation.

~~6244~

9 56,307 Because of the differences in exhaust pressure an each side of the divider plate, there would be incidence at the leading edge of the divider plate at the last rotating blade exit annular. The inlet edge of the plate would be placed far enough downstream so that the last row blades do not make contact because of differential movement during speed and load changes.
.. BRIEF DESCRIPTION OF THE DRAWINGS
For a better understanding of the present invention, reference may be made to the following detailed description taken in conjunction with the accompanying drawings in which:
FIG. 1 is a simplified, partial cross-sectional ~ view of a double flow steam turbine in which a flow divider of the present invention is shown in the left hand exhaust outlet; and FIG. 2 is a simplified, partial cross-sectional view taken through the right-hand end of FIG. 1 to illustrate how it would appear with a flow-divider plate of the present invention.
DETAILED DESCRIPTION OF THE INVENTION
Referring to FIG. 1, there is shown a low pressure double flow steam turbine element 1 and a zoned or multi-pressure condenser 3 incorporating the teaching of the present invention.
The condenser 3 comprises a shell portion 5 which encloses a plurality of horizontally disposed straight tubes 7 with water boxes or headers 9 and 11 disposed ~o~~~~~
56,307 on opposite ends of the shell 5 and tubes 7. An inlet cooling water nozzle 13 is disposed in fluid communication with one of the headers 9 and an outlet cooling water nozzle 15 is disposed in fluid 5 communication with the other header 11 so that influent cooling water enters the right-hand end of the tubes 7 and effluent cooling water is discharged from the left-hand end of the tube 7 as shown in fIG.
1.

10 The turbine comprises a casing or housing 17 which is disposed in fluid communication with the shell 5 of the condenser 3. Rotatably disposed within the housing 17 is a rotor 19 and a plurality of stationary and rotatable interdigitated blade rows 21 and 23, respectively, forming two steam flow paths which originate at the central portion of the housing 17 and extend axially in opposite directions to the axial ends of the turbine 1. A steam inlet nozzle 25 is disposed in the center portion of the housing 17 to supply steam to the blade rows in each flow path.
A partition plate or baffle 27, which may include more than one plate, is disposed within the shell 5 and housing 17 so as to form two separate chambers 29 and 31 within the shell 5 and housing 17. The chamber 29 has tubes with influent cooling water flowing therethrough and the chamber 31 has tubes with effluent cooling water flowing therein so that the back pressure in the chamber 31 which are, respectively, called low and high pressure chambers 29 and 31. The partition plate 27 may be attached to the condenser or turbine housing by welding on one side 11 56,307 and provided with a tongue-and-groove arrangement as shown generally at 33 wherever necessary to allow for thermal expansion of the partition plate 27.
The last row of rotatable blades 23A on the right-hand end of the steam flow path which discharge into the low pressure chamber 29 may be longer 'than the last row of rotatable blades 23B on the left-hand side of the steam flow path which discharges into the high pressure chamber 31, and may include corresponding changes in the last row of stationary blades 21A and 21B. The gauging of the last row of stationary blades 21A or rotating blades 23A may be greater than the gauging in the last row of stationary blades 21B or rotating blades 23B in the flow path.
The zoned or multi-pressure condenser and turbine combination of FIG. 1 as thus far described will have up to 0.7~ better thermal performance than units without multiple pressure or zoned condensers. As previously discussed, Applicants believe that further performance improvement can be attained if the turbine exhaust can be divided into additional zones.
The left-hand half of FIG. 1 illustrates one embodiment of the present invention. A pair of vertical divider plates 35A, 35B are attached to outer flow guide 37 and to inner flow guide 39, which define . an exhaust outlet 47A, and extend therebetween to effectively divide the steam exiting the turbine into a left half and a right half portion 47A', 47A°' when viewed from the exhaust end. Division of the steam into two separate portions is completed by another pair of vertical divider plates 41A, 41B attached to 2~~~,4~~

12 56,307 the outer cylinder wall or housing 17. The plates 41A, 41B are coupled to respective ones of the plates 35A, 35B by tongue and groove or other form of resilient joint, such as joint 33, which joint both facilitates assembly and accommodates any differential thermal expansion of the coupled plates. The plates 41A, 41B may also be welded or otherwise joined to abutting surfaces of the outer flow guide 37, inner cylinder housing 43, and plate 27. As with plate 33, the plate 41B extends through the condenser 3 further dividing the left-hand half of condenser 3 into a front and rear section 3A, 3B as viewed in FIG. 1.
While only one exhaust end of the double flow turbine of FIG. 1 has been shown as incorporating a flow-divider in accordance with the present invention, it will be appreciated that a similar flow-divider could be used on the other exhaust end, with the condenser 3 being further divided into two zones on its right half side. Assuming that the left-hand half of the turbine of FIG. 1 represents a single flow exhaust turbine, a substantial increase in output, i.e., a decrease in heat rate, can be realized.
Furthermore, while a vertically oriented divider plate is shown for the axially aligned exhaust annuli 47A, 47B of FIG. 1, a horizontal divider plate along the horizontal axis or a vertical plate perpendicular to the axis may be used in side exhaust turbines. Other arrangements of divider plates adapted for a particular exhaust will be apparent.
Referring to FIG. 2, there is shown an end view of the turbine of FIG. 1 which, for purposes of 13 56,307 description, will be assumed to be the right-hand end and will be further assumed to incorporate flow-divider plates 41, 35 in accordance with the above description of the left-hand end of FIG. 1. Since each end is essentially a mirror image of the other, the same reference numbers are used on both ends except that the exhaust annulus is designated 47B on the right-hand end. The two plates 41 and 35, further divided into A and B segments, separated the exhaust flow into two fluid paths, one designated 47B' and the other 47B". Each fluid path is coupled to separate sections 3A', 3B' of the condenser 3.
While the improvement is considerably lower on a double flow exhaust such as that of FIG. 1 in which the teachings of U.S. Patent 4,557,113 have been incorporated, the improvement can reasonably be expected to be between 0.25% and 0.7% depending upon the condenser rise. If the heat rate improvement comparison is made with an unzoned double flow exhaust, the improvement would be in excess of 1%. If the turbine has side exhausts, the increase in hood loss is minimal with the proposed arrangement.
Angled slots 45 may be formed in the divider plates 35A, 35B to transfer flow between a high pressure zone and a lower pressure zone resulting from the swirl that occurs at higher exhaust pressures and thereby reduce flow separation in the hood.
The incorporation of the divider plates 35, ~41 at the turbine blading exhaust results in substantial reduction in heat rate. The maximum improvement occurs when it is applied on a single flow exhaust 2~~2~~!~

14 56,307 with output increases of about 1%, in spite of increased hood loss. With side exhaust turbines, there is a potential increase of still greater magnitude. When comparing a four zone arrangement (left and right-hand ends of FIG. 1 being divided) with a two zone arrangement as shown in U.S. Patent No. 4,557,113, an improvement of 0.250 and 0.5o is feasible. Although the blading experiences shock loading as it moves from one zone to another, the clearance between the blade exit plane and the divider inlet allows this transition to be reduced in severity.
While there is an anticipated exhaust pressure differential across the divider plates 35, incidence occurs along the leading edge of the plates. This incidence would result in poorer hood performance than would occur with single pressure operation without the divider. Table VII (13.3°C rise) and Table VIIT
(20.0°C rise) compare a single or unzoned design with a two zone design with 0, 0.56, 1.11, and 1.67 Kcal/Kg increases in hood loss. Table IX (13.3°C rise) and Table X (20.0°C rise) compared the two zone design (with no increase in hood loss) with the four zone design with 0, 0.56, 1.11, and 1.67 Kcal/Kg increases in hood loss. The reason for the negative improvement at low exhaust steam temperature is two-fold. First, the low pressure zones are choked and cannot utilize all of the improvement in exhaust pressure. See 42.2°C
case on Table IX. Second, the performance in the highest pressure zone is degraded because of the increased hood loss.

15 56,307 In reality, the hood loss increase should be close to zero at the low steam temperatures because the turbine exhaust flow is close to axial and there would be low incidence on the divider between the two halves at a given flow. At the high exhaust temperatures, the increase in hood loss would be closer to the 1.67 Kcal/Kg value.
While the principles of the invention have now been made clear in an illustrative embodiment, it will become apparent to those skilled in the art that many modifications of the structures, arrangements, and components presented in the above illustrations may be made in the practice of the invention in order to develop alternate embodiments suitable to specific operating requirements without departing from the spirit and scope of the invention as set forth in the claims which follow.

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16 56,307 TABLE I
TWO ZONE
VS SINGLE
ZONE
(UNZONED) PERFORMANCE

20.0C Temperature Rise Isentropic Increased Sat. Sat. Press.Moisture,Enthalpy Heat Drop Temp.

C Kcal/sqcm %, at Kcal/Kg Kcal/Kg T2=76.7 P2=.4213 3.00 h2=613.5 0.0 Tl=66.7 P1=.2747 hl=598.0 15.5 TO=56.7 T2=72.2 P2=.3496 3.70 h2=607.8 0.0 T1=62.2 P1=.2250 hl=592.1 15.7 TO=52.2 ' T2=66.7 P2=.2747 4.51 h2=600.8 0.0 T1=56.7 P1=.1738 hl=585.1 15.7 TO=46.7 T2=61.1 P2=.2138 5.50 h2=592.8 0.0 T1=51.1 P1=.1329 h1=576.7 16.1 TO=41.1 T2=55.6 P2=.1648 6.44 h2=585.0 0.0 T1=45.6 P1=.1005 h1=568.7 16.3 TO=35.6 T2=50.0 P2=.1258 7.56 h2=576.1 0.0 T1=40.0 P1=.0752 hl=559.7 16.4 TO=30.0 ~osz~4~

17 56,307 TABLE II
FOUR ZONE PERFORMANCE
ZONE
VS SINGLE
(UNZONED) 20.0C Temperature Rise IsentropicIncreased Sat. Sat. Press.Moisture, Enthalpy Heat Temp. Drop C Kcal/sqcm %, at P2 Kcal/Kg Kcal/Kg T4=76.7 P4=.4213 3.00 h4=613.5 0.0 T3=71.7 P3=.3414 h3=605.8 7.7 T2=66.7 P2=.2747 h2=598.0 15.5 Tl=61.7 P1=.2193 h1=590.2 23.3 TO=56.7 T4=72.2 P4=.3496 3.70 h4=607.8 0.0 T3=67.2 P3=.2815 h3=599.9 7.9 T2=62.2 P2=.2250 h2=592.1 15.7 T1=57.2 P1=.1784 hl=584.2 23.6 TO=52.2 T4=66.7 P4=.2747 4.51 h4=600.8 0.0 T3=61.7 P3=.2193 h3=593.0 7.8 T2=56.7 P2=.1738 h2=585.1 15.7 T1=51.7 P1=.1366 h1=577.0 23.8 TO=46.7 T4=61.1 P4=.2138 5.50 h4=592.8 0.0 T3=56.1 P3=.1693 h3=584.8 8.0 T2=51.1 P2=.1329 _ h2=576.7 16.1 T1=46.1 Pl=.1034 h1=568.6 24.2 T0=41.1 T4=55.6 P4=.1648 6.44 h4=585.0 0.0 T3=50.6 P3=.1293 h3=576.9 8.1 T2=45.6 P2=.1005 h2=568.7 16.3 T1=40.6 P1=.0775 h1=560.5 24.5 TO=35.6 T4=50.0 P4=.1258 7.56 h4=576.1 0.0 T3=45.0 P3=.0977 h3=567.9 8.2 T2=40.0 P2=.0752 h2=559.7 16.4 Tl=35.0 P1=.0573 hl=551.3 24.8 TO=30.0 206~4~~

18 56,307 TABLE III
TWO ZONE VS SINGLE ZONE PERFORMANCE
13.3°C Temperature Rise Isentropic Increased Sat. Sat. Press.Moisture, Enthalpy Heat Drop Temp.

C Kcal/sqcm %, at P2 Kcal/Kg Kcal/Kg T2=70.0 P2=.3178 4.42 h2=602.8 0.0 T1=63.3 P1=.2366 h1=592.4 10.4 TO=56.7 T2=64.4 P2=.2488 5.28 h2=595.6 0.0 T1=57.8 P1=.1831 hl=585.0 10.6 T0=51.1 T2=58.9 P2=.1929 6.12 h2=588.3 0.0 T1=52.2 P1=.1403 hl=577.6 10.7 TO=45.6 T2=53.3 P2=.1481 6.95 h2=581.1 0.0 T1=46.7 P1=.1064 h1=570.3 10.8 TG=40.0 T2=47.8 P2=.1126 7.86 h2=573.3 0.0 T1=41.1 P2=.0798 h1=562.3 11.0 TO=34.4 T2=42.2 P2=.0846 8.87 h2=566.1 0.0 T1=35.6 P1=.0591 hl=554.9 11.2 TO=28.9 19 56,307 TABLE IV
FOUR ZONE VS SINGLE ZONE PERFORMANCE
13.3°C Temperature Rise Isentropic Increased Sat. Sat. Press.Moisture,Enthalpy Heat Drop Temp.

C Kca1/sqcm %, at Kcal/Kg Kca1/Kg T4=70.0 P4=.3178 4.42 h4=602.8 0.0 T3=66.7 P3=.2746 h3=597.6 5.2 T2=63.3 P2=.2366 h2=592.4 10.4 T1=60.0 P1=.2031 hl=587.1 15.7 TO=56.7 T4=64.4 P4=.2488 5.28 h4=595.6 0.0 T3=61.1 P3=.2138 h3=590.3 5.3 T2=57.8 P2=.1831 h2=585.0 10.6 T1=54.4 P1=.1563 h1=579.7 15.9 TO=51.1 T4=58.9 P4=.1929 6.12 h4=588.3 0.0 T3=55.6 P3=.1648 h3=583.0 5.3 T2=52.2 P2=.1403 h2=577.6 10.7 T1=48.9 P1=.1190 h1=572.2 16.1 TO=45.6 T4=53:3 P4=.1481 6.95 h4=581.1 0.0 T3=50.0 P3=.1258 h3=575.7 5.4 T2=46.7 P2=.1064 h2=570.3 10.8 T1=43.3 P1=.0896 hl=564.8 16.3 TO=40.0 T4=47.8 P4=.1126 7.86 h4=573.3 0.0 T3=44.4 P3=.0949 h3=567.8 5.5 T2=41.1 P2=.0798 h2=562.3 11.0 T1=37.8 P1=.0668 h1=556.8 16.5 TO=34.4 T4=42.2 P4=.0846 8.8? h4=566.1 0.0 T3=38.9 P3=.0709 h3=560.6 5.5 T2=35.6 P2=.0591 h2=554.9 11.2 T1=32.2 P1=.0491 hl=549.3 16.8 TO=28.9 20 56,307 TABhE V
STEAM PRESSURE AND TEMPERATURE IN SINGLE AND TWO ZONE
CONDENSERS
Cond. Zone 1 Zone 2 Rise Temp.* Press.* Temp.Press.

C C Kg/sqcm C Kg/sqcm 13.3 42.2 .0846 35.6 .0591 13.3 47.8 .1126 41.1 .0798 13.3 53.3 .1481 46.7 .1064 13.3 58.9 .1929 52.2 .1403 13.3 64.4 .2488 57.8 .1831 13.3 70.0 .3178 63.3 .2366 20.0 50.0 .1258 40.0 .0752 , 20.0 55.6 .1648 45.6 .1005 20.0 61.1 .2138 51.1 .1329 20.0 66.7 .2747 56.7 .1738 20.0 72.2 .3496 62.2 .2250 20.0 76.7 .4213 66.7 .2747 * Operating condition with an unzoned or single pressure condenser 21 56,307 TABLE VI
STEAMPRESSURE AND TEMPERATURE TWO
WITH AND
FOUR
ZONE

CONDENSERS

Cored.Zone 1 Zone 2 RiseTemp.* Press.* Temp.Press.

C C Kg/sqcm C Kg/sqcm 13.342.2 .0856 38.9 .0709 13.347.8 .1126 44.4 .0949 13.353.3 .1481 50.0 .1258 13.358.9 .1929 55.6 .1648 13.364.4 .2488 61.1 .2138 13.370.0 .3178 66.7 .2747 20.0~ 50.0 .1258 45.0 .0977 20.055.6 .1648 50.6 .1293 20.061.1 .2138 56.1 .1693 20.066.7 .2747 61.7 .2193 20.072.2 .3496 67.2 .2815 20.076.7 .4213 71.7 .3414 Cored.Zone 3 Zone 4 RiseTemp.* Press.* Temp.Press.

C C Kg/sqcm C Kg/sqcm 13.335.6 .0591 32.2 .0491 13.341.1 .0798 37.8 .0668 13.346.7 .1064 43.3 .0896 13.352.:d .1403 48.9 .1190 13.357.8 .1831 54.4 .1563 13.363.3 .2366 60.0 .2031 20.040.0 .0752 35.0 .0573 20.045.6 .1005 40.6 .0775 20.051.1 .1329 46.1 .1034 20.056.7 .1738 51.7 .1366 20.062.2 .2250 57.2 .1784 20.066.7 .2747 61.7 .2193 * Operating conditions with a two zone condenser 22 56,307 TABLE III
INCREASE IN OUTPUT FROM ZONED CONDENSER 13.3°C
CONDENSER RISE SINGLE FLOW LP SECTION
TWO ZONE VS ONE ZONE CONFIGURATION
(EFFECT OF HOOD LOSS INCREASE, AHL, ON TWO ZONE CONFIGURATION) Steam Temp.
Top 1 Zone Two Zone Output, KW
Zone, Output, ~ AHL=0* aHL=0.68* flHL=1.1* ~HIr-1.7*
°C KW KW KW KW KW
42.2 432,725 432,787 432,766 432,735 432,690 47.8 429,689 431,184 431,076 430,883 430,729 53.3 423,476 427,021 426,545 426,207 425,873 58.9 414,776 419,772 419,299 418,809 418,294 64.4 405,368 410,845 410,272 409,698 409,133 ?0.0 395,559 401,258 400,640 400,495 399,936 Steam Temp. Two Zone Increase Top 1 Zone In output, KW
Zone, Output, ~HL=O* NHL=0.68* NHL=1.1* ~F~=1.7*
°C KW KW KW KW KW
42.2 432,725 62 41 10 -35 47.8 429,689 1495 1387 1194 1040 53.3 423,476 3545 3069 2731 2397 58.9 414,776 4996 4523 4033 3518 64.4 405,368 5487 4904 4330 3765 70.0 395,559 5699 5081 4936 4377 * OHL is given Kcal/Kg 23 56,307 TABLE VIII
INCREASE IN OUTPUT FROM ZONED CONDENSER 20.0°C
CONDENSER RISE SINGLE FLOW LP SECTION
TWO ZONE VS ONE ZONE CONFIGURATION
(EFFECT OF HOOD LOSS INCREASE, NHL, ON TWO ZONE CONFIGURATION) steam Temp.
Top 1 Zone Two Zone Output, KW
Zone, Output, OHL=0* L1HL=0.68* ~HL=1.1* ~HIl-1.7*
°C KW KW KW KW KW
50.0 427,568 430,078 429,741 429,761 429,577 55.6 420,009 425,442 425,111 424,784 424,434 61.1 411,040 418,523 418,052 417,574 417,096 66.7 401,615 409,790 409,221 408,582 '408,010 72.2 392,153 400,423 399,735 399,038 398,338 76.7 382,232 391,628 390,899 390,177 389,391 Steam Temp. Two Zone Increase Top 1 Zone In Output, KW

Zone,Output, AHL=0* ~HL=0.68*NHL=1.1* AFiL=1.7*

C KW KW KW KW KW

50.0 427,568 2510 2373 2193 2009 55.6 420,009 5433 5002 4775 4425 61.1 411,040 7483 7012 6534 6056 66.7 401,615 8175 7606 6967 6395 72.2 392,153 8270 7582 6885 6185 76.7 382,232 9396 8667 7945 7159 * NHL is given Kcal/Kg 20~~~~~

24 56,307 TABLE TX
INCREASE IN OUTPUT FROM ZONED CONDENSER 13.3°C
CONDENSER RISE DOUBLE FLOW LP SECTION
FOUR ZONE VS TWO ZONE CONFIGURATION
(EFFECT OF HOOD LOSS INCREASE, dHL, ON TWO ZONE CONFIGURATION) Steam Temp.
Top 2 Zone Four Zone Output, KW
Zone, Output, AHL=0* NHL=0.68* AHL=1.1* ~1.7*
°C KW KW KW KW KW
42.2 432,787 432,805 432,709 432,697 432,677 47.8 431,184 431,613 431,503 431,407 431,289 53.3 427,021 428,303 428,037 427,754 427,475 58.9 419,772 421,913 421,475 421,030 420,523 64.4 410,845 413,474 413,138 412,386 411,884 70.0 401,258 403,819 403,336 402,423 402,172 Steam Temp. Four ZoneIncrease Top 2 Zone In Output, KW

Zone,Output, NHL=0* NHL=0.68*NHL=1.1* ~HIr-1.7*

C KW KW KW KW KW

42.2 432,787 18 -78 -90 -110 47.8 431,184 429 319 223 105 53.3 427,021 1282 1016 733 454 58.9 419,772 2141 1703 1258 751 64.4 410,845 2629 2293 1541 1039 70.0 401,258 2561 2078 1165 914 * AHL is give Kcal/Kg 25 56,307 TABLE X
INCREASE IN OUTPUT FROM ZONED CONDENSER 20.0°C
CONDENSER RISE DOUBLE FLOW LP SECTION
FOUR ZONE VS TWO ZONE CONFIGURATION
(EFFECT OF HOOD LOSS INCREASE, AHL, ON TWO ZONE CONFIGURATION) Steam Temp.
Top 2 Zone Four Zone Output, KW
Zone, Output, AHL=0* NHL=0.68* NHL=1.1* ~I~-1.7*
°C KW KW KW KW KW
50.0 430,078 431,958 431,083 430,737 430,609 55.6 425,442 427,443 427,192 426,936 426,665 61.1 418,523 421,601 421,195 420,787 420,257 66.7 409,790 413,684 413,148 412,601 412,029 72.2 400,433 403,910 403,181 402,418 401,648 76.7 391,628 394,653 393,275 392,396 391,802 steam Temp. Four ZoneIncrease Tap 2 Zone In Output, KW

Zone,Output, flHL=0*~HIr-0.68*NHL=1.1* ~HI~-1.7*

C KW KW KW KW KW

50.0 430,078 1880 1005 659 531 55.6 425,442 2001 1750 1494 1223 61.1 418,523 3078 2672 2264 1734 66.7 409,790 3894 3358 2811 2239 72.2 400,433 3477 2748 1985 1215 76.7 391,628 3025 1647 768 174 * pHL is given Kcal/Kg

Claims (6)

1. A low pressure steam turbine and condenser combination having multiple pressure zones in a single exhaust flow comprising:
a condenser divided into multiple sectors;
a turbine housing in fluid communication with said condenser for passing exhaust steam from the turbine into the condenser;
at least one exhaust outlet coupled to the turbine and positioned to exhaust steam into said housing;
at least one divider plate positioned in said exhaust outlet and extending into said housing for dividing exhaust steam into at least two separated flow paths, each flow path being coupled to a respective one of the multiple sectors of said condenser and;
a plurality of slots in said at least one divider plate adjacent said at least one exhaust outlet for controlling flow separation related to swirl in the stream at relatively high exhaust pressure.

2. The low pressure steam turbine of claim 1 wherein the turbine comprises a double flow turbine having a second exhaust outlet positioned to exhaust steam into said housing and further comprising;
a second divider plate positioned in said second outlet and extending into said housing for dividing exhaust steam from said second outlet into at least two separate second flow paths, each of said second flow paths being coupled to a respective one of the multiple sectors of said condenser.

3. The low pressure steam turbine of claim 2, including a third divider plate extending through said housing generally transverse to the orientation of said at least one divider plate and said second divider plate for separating exhaust flow from each exhaust outlet into two substantially isolated portions of said housing.

4. A low pressure steam turbine and condenser combination, the turbine having at least one exhaust annulus for exhaust steam, the improvement comprising:
means for dividing the exhaust steam into at least two substantially isolated flows, and means directing each of the two flows into respective sections of the condenser, said dividing means comprising a divider plate positioned in said annulus and dividing said exhaust steam into two substantially equal parts, and including vents in said divider plate for permitting flow from one side of said plate to another for controlling flow separation when swirl is present in the steam flow.

5. The low pressure turbine of claim 4 wherein the turbine comprises a double flow turbine and wherein each exhaust thereof is divided into at least two flows, each of said flows being directed through respective isolated sections of the condenser.

6. The low pressure steam turbine of claim 2 wherein the condenser comprises a shell and tube condenser and including baffling disposed in the condenser for isolating the two flows therethrough.