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Studies of the direct input of solar energy to a fossil-fueled central station steam power plant
Solar Energy, Vol. 17, pp. 297-305. Pergamon Press 1975. Printed in Great Britain
STUDIES OF THE DIRECT INPUT OF SOLAR ENERGY TO A FOSSIL-FUELED CENTRAL STATION STEAM POWER PLANT* R. J. ZOSCHAK and S. F. Wu Research Division,Foster Wheeler Corporation, 12Peach Tree Hill Rd., Livingston,NJ 07039, U.S.A.
(Received26June 1974) A~traet--Seven possible methods of absorbing solar energy as direct thermal input to an 800 MW, fossil-fueled, central station steam power plant have been studied. Irrespective of method, the solar heat is first collected by an array of flat mirrors and concentrated on a tower-mountedabsorber where it is transferred into the power cycle. The heat absorbing methods studied were heating of feed-water, evaporation of water, superheatingof steam, combined evaporation and superheating,reheatingof steam, air preheating,and combinedair preheating,and feedwater heating. Factors considered were relative capital cost, energy conversion efficiencyand complexityof design, operation and control. Combinedevaporation and superheatingproved to be the preferred method because of its high utilizationof solar energy, relatively low indicated capital cost and only moderate complexity in design, operation and control. Feedwater heatingalso has very desirablecapitalcost, designand operatingaspects, but suffers from the drawbackthat over 30per cent of the solar energyabsorbed is, in effect, lost becauseof degradationof the steam cycle efficiency.
INTRODUCTION
ASSUMED BASE LINE CONDITIONS
One possible way to use solar energy for the production of commercial electric power would be to use it as a supplemental source of thermal input to a conventional, fossil-fueled, central station steam power plant. This would assure capability of the plant to supply electrical power at all times. During periods of cloud cover and at night, load would be carried solely on fossil fuel. During periods of sunshine, fossil fuel firing rate would be decreased, thus saving fuel and reducing stack emissions. The technical and economic evaluation of a combined solar/fossil-fueled plant of this sort is the subject of a current study sponsored by the National Science Foundation.f The system being considered consists of an array of many flat mirrors arranged and controlled to reflect and concentrate the rays of the sun on a tower-mounted heat exchanger or absorber. The heat thus collected by the absorber would be fed directly into the steam power cycle without any intervening heat transfer fluid, heat storage or topping cycle. The supporting tower might also serve as the stack of the power plant. This is a report on that portion of the study concerned with examining, comparing and evaluating the various possible methods of adding solar heat to a modern, high pressure, high temperature steam cycle. Factors considered in this exercise were relative capital cost as indicated by material requirements, energy conversion efficiency, and complexity of design, operation and control.
To keep the range of the study within reasonable bounds, the following base line conditions have been assumed. (1) The reference fossil-fueled plant is to have a full load electrical output of approximately 800MW and normal full load turbine steam conditions of 2400 psig, 1000F superheat/1000F reheat. Figure 1 is a simplified heat balance diagram of the reference plant. In this diagram the Low Pressure Feedwater Heater (LP FWH) and Intermediate Pressure Feedwater Heater (IP FWH) each represent 3 separate stages of regenerative feedwater heating, so that in the actual plant there are 7 feedwater heaters. (2) Maximum solar thermal input is to be 500 MW. (3) Full load plant electrical output is to remain at 800 MW under combined solar/fossil-fueled operation.
CANDIDATE METHODS OF SOLAR HEAT INPUT
*This paper was presented at the International Solar Energy Society's U.S. Section Meeting held in Fort Collins, Colorado (Aug. 1974). fNSF Grant No. GI-41019. Sheldahl is prime contractor. Subcontractors are Foster Wheeler Corporation and the University of Minnesota. Northern States Power Company is a technical consultant.
Seven candidate methods for absorbing solar heat into the steam cycle have been studied. These are listed below. (1) Feedwater heating--FWH. (2) Evaporation of water--EVAP (3) Superheating of steam--SH (4) Combined evaporation and superheating--EPAV & SH (5) Reheating of steam--RH (6) Air preheating--AH (7) Combined air preheating and feedwater heating-AH & FWH. In the text and figures to follow, an asterisk after an abbreviation denotes the solar heated counterpart of that particular component of the steam cycle. Feedwater heating. The schematic arrangement of incorporating a Feedwater Heater (FWH*) in a fossilfueled steam plant is shown in Fig. 2. During solar input,
297
298
R.J. ZOSCHAKand S. F. wu
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5.5x106#
2535P
IO00F
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561P
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-
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Fig. 2. Simplifiedheat balancediagramshowingsolar-heatedfeedwater heater in reference plant.
extraction steam from the turbine to the high and intermediate pressure FWH's is cut off except for a small amount to keep the heaters at hot standby condition to minimize thermal transients when solar heating ceases. Feedwater is pumped to the FWH* as indicated by the broken line in the diagram. After being heated in the FWH*, the water is fed back to the economizer section of the fired boiler. The fuel firing rate is reduced by the amount equivalent to that supplied by solar input to the FWH*, and the furnace portion is operated at part load. The amount of feedwater circulating to the FWH* can be chosen to produce either the same flow rate at the turbine throttle or the same flow rate at the condenser inlet as that of the reference cycle. The turbine-generator output will be increased for the former condition since no steam will be extracted from the high and intermediate pressure stages of the turbine. For the latter condition, less steam will pass through the higher pressure stages of the turbine and the output will be reduced. Obviously, these two different operating conditions would require two different turbine designs. In order to compare the different schemes on an equal plant output basis in this study, the FWH* flow has been set at a rate which results in the same turbine-generator output as in the reference cycle. Without regenerative feedwater heating by the conventional High Pressure
(HP) and Intermediate Pressure (IP) FWH when the FWH* is in service, the plant thermal efficiency is reduced. Evaporation. In this arrangement, the Evaporation of water (EVAP*) is put in parallel with the fired EVAP as depicted in Fig. 3. When EVAP* is in operation, the heat absorption in the fired steam generator has to be re-distributed between the partially loaded EVAP and other heat transfer surface. A differential firing scheme can be used for this purpose, and adoption of a twin-furnace design is one well-established way of doing it. One furnace includes superheater (SH), reheater (RH) and economizer of pre-evaporator (ECON) sections and the other houses primarily the EVAP section. The amount of feedwater by-passed to EVAP* is proportional to the thermal capacity replaced. The balance of the feedwater circulates through the conventional EVAP and the firing rate of the EVAP furnace is reduced accordingly. During the operation, the firing rate of the other furnace is kept unchanged. A special unconventional arrangement of heat transfer surface is required in the fired steam generator in this scheme. Superheating. In this arrangement, the SH* is put in parallel with the fired SH as shown in Fig. 4. The amount of saturated steam required for SH* is drawn from the boiler drum and the remaining portion is fed to its fired
Directinput of solar energyto a fossil-fueledsteampowerplant
299
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680F
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, '
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Fig.3. Simplifiedheat balancediagramshowingsolar-heatedevaporatorin referenceplant.
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counterpart. The fuel firing rate is reduced in the furnace to account for the reduced heat absorption of the SH section. To re-distribute the heat absorption among the surfaces in the furnace, the convective SH and RH sections are arranged in parallel passes and the flue gas crossing each section is apportioned by control dampers. To maintain the reheat steam temperature constant during operation, control dampers are adjested so that more flue gas is directed to the RH section. Superheater outlet steam temperature is controlled by a spray type desuperheater wherein feedwater is evaporated to cool the steam. Flow of feedwater to the boiler is consequently reduced by the amount by-passed to the superheat control device. The operation of this scheme is relatively complicated due to the dual control involved. Combined evaporation and superheat!ng. Figure 5 depicts this combined arrangement. The total solarassisted capacity (500MWt) is distributed between EVAP* and SH* in the same proportions as the heat absorption in their counterparts in the fired unit. For this scheme, no special surface arrangement is needed in the fired steam generator. Superheated steam temperature
can be controlled by water spray and the absorption of the RH in the furnace is maintained at its proper rate by biased-flow damper control as described previously. Since the percentage of the combined evaporation and superheating capacity replaced by solar heating is much lower than the percentage of capacity replaced when solar heating is applied to each component alone, fewer operation and control problems are expected for this combined type of arrangement. Reheating. This arrangement is shown in Fig. 6. During operation, the steam to be reheated is by-passed to RH* except for a small amount which is passed through the fired RH to keep it from overheating. In the furnace, the RH is arranged in parallel passes with the SH so that flue gas flow can be controlled by dampers to by-pass RH completely when RH* is in service. The feedwater flow to the boiler remains at the full-load rate, but the fuel firing rate is reduced to the level needed to supply the heat input requirements of the SH, EVAP and ECON. A relatively large volume of high temperature steam has to be transported back and forth between the power plant and solar receiver for this scheme, which makes the piping
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Fig. 6. Simplifiedheatbalancediagramshowingsolar-heatedreheaterin referenceplant.
line expensive. Also, circulating steam through the idling RH to prevent overheating would further reduce the already limited capacity of the RH*. Air preheating. In this arrangement, the Air Preheating (AH*) is put in parallel with the air preheater section of the steam generator as shown in Fig. 7. To maximize the capacity of the AH*, the air required for combustion is preheated to 1000°F during the period of solar input, which is a considerably higher temperature than normally used in the conventional mode of operation. With the conventional AH not in service, the energy in the high temperature flue gas discharged directly to the stack is not recovered and the efficiency of the steam generator is reduced. Consequently, not all of the solar input is useful to the cycle. Fuel firing rate is cut by an amount equal to the net useful thermal capacity gained by the AH*, while the other components in the power plant are not affected by this operation. When the air supply fan is located upstream of the AH* as shown in the figure, the fluid side heat transfer surface has to be enclosed. If the fan is located downstream of the AH*, it must be designed to
withstand high air temperature and possible rapid temperature changes. Another conceivable alternative is to locate the fan downstream of the conventional AH where the gas temperature is always relatively low. However, this would require extra reinforcement of the steam generator casing to withstand the high negative pressure differential, and a more costly conventional AH of tubular design to prevent air leakage. In any event, high temperature air ducting is required and the burners must be specially designed for high temperature. Combined air preheating and .feedwater heating. This combined arrangement is a modification of the previous approach of rising the AH* alone. As shown in Fig. 8, an additional heat exchanger is installed in the flue gas pass downstream of the ECON so that stack gas can be used to heat feedwater when the conventional AH is taken out of service. This FWH is arranged in parallel with the conventional intermediate pressure feedwater heaters and takes over part of their thermal duty during solar input periods. Except that it is not solar-heated, its mode of operation is similar to that of the FWH* as described in
Direct input of solar energy to a fossil-fueledsteam power plant
301
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the first scheme. To maintain the same turbine-generator output, the feedwater flow rate through the stack FWH must be reduced below what would normally flow through the regenerative IP FWH because of the reduced steam extraction from the turbine. The turbine cycle efficiency is reduced when the regenerative IP FWH is out of service, but the heat recovery from the stack gas greatly improves the boiler efficiency. The net result is an increase in the net utilization of solar input as compared to using AH* alone. CO~IP~'~ATIVESTUDY~ 0 ~SULTS The 7 candidate schemes were studied to identify their design and operating conditions, to estimate their physical dimensions, and to evaluate the trade-offs among them. The pertinent parameters selected for this study were divided into four groups as presented in Tables 1-4. Table 1 lists the thermal capacity and operating conditions of each arrangement. The capacity of each com-
ponent is based on operation of the reference plant cycle at 5 per cent over-pressure condition as shown in Fig. 1. At this condition the thermal input and electric output of this cycle are 1885 MWtt and 815 MWe, respectively. For the combined arrangements such as EVAP* & SH* and AH* & FWH, The results for each component are shown, separated by a slant, wherever appropriate. The net useful solar capacity of each scheme was calculated by subtracting from its solar-heated capacity an amount that accounts for any consequent reduction in thermal efficiency of the plant. From the table we see that only three schemes, EVAP*, SH* and EVAP* & SH*, are capable of absorbing the maximum solar input of 500 MWt and also providing a net useful solar capacity of ?Note that the 1885MWt is the input to the steam turbine cycle or, in other words, the output of the steam generator. The total thermal input to the plant is 1885MW divided by the efficiencyof the steam generator.
R. J. ZOSCHAK and S. F. Wu
302
Tablel. Thermalcapacityandoperatingconditionsofsolarheatabsorbingschemes Scheme
EWE*
EVAP*
EVAP* & SH*
SH*
RH*
AH & FWH*
AH*
Compound capacity in the reference fossil plant (Mwt) 425
745
630
1375-
305
200
200/170
Capacity replaceable by solar heating (MWt)
425
500
500
270/230
305
400
400
Net useful solar c a p a c i t y (MWt)
290
500
500
305
200
305
Fluid temperature Fluid p r e s s u r e
(°F)
240-525
(psi)
680
3000
500
680-1000
680-1000
2500
2750
2750
640-1000
80-1000
600
80-1000/240-370
15
15/500
Table 2. Thermal/hydraulic characteristics of solar heat absorbing schemes FWH*
EVAP*
SH*
EVAP* & SH*
RH*
AH*
4.75
14.8
4.37
8.0/2.0
5.02
6.0
6.0/4.53
Fluid mass velocity (1061bZhr-ft)
1.5
1.0
1.0
1.0
0.3
0.02
0.02/1.0
Average fluid velocity (ft/sec)
7.7
ii
56
11/56
102
134
134/5
Fluid heat transfer cosff. (Btu/hr-°F-ft 2 )
2,500
9,000
11000
9,000/i,000
260
15
15/1,700
300,000
180,000
i00,000
44,200
12,000
12,000~8,500
120
20
i00
170
600
4,840
9,500
17,000
5,130/7,820
6
50
6~50
Component Flow rate
(1061b/hr)
Heat flux (Btu/hr-ft 2 ) Fluid film ~ T
(OF}
Heat transfer surface (ft 2 ) A p p r o x . ~ P through the component (psi)
i0
180,000/100,000 20z100
AH* & FWH
800/5
23,600 114,000
25
114,000/68,300
0.6
0.6/12
Table 3. Materials and weights for solar heat absorbing schemes Component
FWH*
Design temperature Design p r e s s u r e
(°F)
800
Weight of surface
RB*
AH*
AH* & FWH
1200
800/1200
1200
1500
1500/800
3000
3000
800
30
30/800
T22
T22
T22/TP304
T22/TP304
T22/TP304
TP304
TP304/T22
0.2
0.2
0.41
0.2/0.41
0.14
0.034
0.034~0.06
275
165
190
165/190
40
5
5/5
45,400
90,000
392,000
229,000
160,000
160,000z645,000
(OF)
(ib)
EVAP* & SH*
800
3000
Tube material
across tube wall
SH*
3000
(psi)
Minimum wall thickness (in) ~T
EVAP*
149,000
Table 4. Approximate piping dimensions and pressure losses of solar heat absorbing schemes EVAP *
Component Pipe flow rate
(106 ibZhr)
FWH*
EVAP*
SH*
& SH*
4.75
3.69
4.37
2.0
20x2
20x2
32x3.4
RH*
AH*
5.02
6.0
Supply line: Pipe O.D. x thickness
(in)
NO. of parallel lines Velocity ~P
(psi)
1
(ft/sec) frictional/gravitational
1
1 39
16xi.6 1
40xi.3 2
16
17
14
87
38/413
31t307
28162
26/307
14/8
20x2
24x2.5
34x5.8
30x5.1
40x2.4
Return line: Pipe O.D. x thickness
(in)
No. of parallel lines Velocity ~P
(ft/sec)
(psi) frictional/gravitational
Total pipe line ~ P (psi) frictional/gravitational Total piping w e i g h t
(103 ib)
1
1
2
1
20
58
62
78
47/-338
71t-62
30/-25
50~-25
85/75
102/245
58/37
950
i,ii0
5,600
ll0x½
3
4
104
243
13/-5
1.2 /0.33
76/282
27/3
1.2/0.33
2,000
4,900
3,000
Direct input of solarenergyto a fossil-fueledsteam powerplant 500 MWt. RH* is limited to a solar input of 305 MWt. FWH*, AH* and AH* & FWH are not only limited to less than 500 MWt input, but also have respective net useful solar inputs which are only about 68, 50 and 76 per cent of their total solar inputs. Table 2 shows the thermal and hydraulic characteristics of the components of the candidate schemes. The fluid mass velocity in each component was chosen so that the fluid velocity and pressure drop through the unit would be within the range normally used in the design of each type of component for conventional power plants. In the evaporator, a recirculation ratio of 4:1 was assumed, which is equivalent to an exit steam quality of 25 per cent by weight. Fluid side heat transfer coefficients were determined from applicable fluid properties and correlations for the respective fluids and flow regimes. The heat flux given for each component except the evaporator is the limiting heat flux at which the calculated mean tube wall temperature closely approaches the design temperature of the tube material. For the evaporator, the heat flux was selected to be conservatively below the critical heat flux for Departure from Nucleate Boiling (DNB) at the assumed pressure, mass velocity and steam quality in the tubes. The heat flux and heat transfer surface given in the table are both based on the inside surface of the tubes. The pressure drop shown is the drop through the heat transfer tubes only and does not include pressure losses in distribution piping and headers. Factors concerning material selection and weights are covered in Table 3. Design preseure is based on the estimated maximum operating pressure to be expected in the component plus a design margin. Design temperature is based on practical upper limits for the materials used under the given operating conditions. To simplify the selection of tube material only two typical materials were considered, one ferritic (ASTM SA-213 T22) for moderate temperatures and the other austenitic (ASTM SA-213 TP304) for high temperatures. In the SH* and RH*, both materials were required for different temperature ranges. Two-inch O.D. tubes were used throughout since experience has shown that this tube size is near optimum for all components. Tube wall thicknesses shown in the table are as calculated and are not adjusted to the next largest standard wall thickness. A maximum design temperature of 1500°F was assumed for the AH*, although the air outlet temperature and air-to-tube AT were taken as 1000°F and 800°F, respectively. In an actual design, this would require a reduced heat flux at the hot end of the exchanger and a higher heat flux at the cool inlet end. Table 4 presents the pipe line dimensions and approximate piping pressure losses for each scheme. Listing for the AH* & FWH has been omitted because it is the same as for the AH*. (The relatively small amount of additional piping required for the stack FWH has been neglected.) Figures for the supply line from power plant to tower-mounted component and for the return line to the power plant are listed separately. Pipe friction loss was estimated by assuming a 1200-ft line length in each direction plus an additional loss of 10 velocity heads (5 for the AH*) to account for losses in valves, bends, entrance and exit. Static pressure change due to gravity head was
303
calculated by assuming a tower height of 1000 ft. Piping associated with the SH* and RH* was sized to give pressure losses roughly comparable to corresponding line losses in the reference plant. Reasonable pressure drops were assumed for sizing the piping for other components. Because of the low allowable pressure drop for RH* and AH* piping, we see that very large pipes are required. It is even conceivable that a tubular supporting tower could serve as a single large duct for the AH*. Comparing Tables 3 and 4 we see that the weight of piping exceeds the weight of the heat transfer surface by more than a factor of 10 for every solar-heated component except the EVAP* & SH*. It might be possible to reduce piping weight by putting the supply pipe within the return pipe so that the inner pipe need withstand only the differential pressure. However, this would introduce problems of insulating the inner pipe to minimize regenerative heat transfer and supporting the inner pipe while accomodating the relative thermal expansion of the pipes. SELECTION OF PREFERRED METHOD
To facilitate comparison of the candidate schemes, the results of the comparative study and other important factors are summarized in Table 5. The first 6 items in the table are significant ratios derived from pertinent values given in Tables 1--4. A low ratio of solar-heated component duty to reference plant component duty is generally indicative of better ability to maintain high solar input at partial plant load and, for the EVAP*, SH*, EVAP* & SH* and RH*, it also indicates fewer problems in designing and controlling the fired steam generator to accomodate the full range of solar input. A higher ratio of useful solar capacity to reference plant capacity means a greater saving in fossil fuel and, therefore, a greater justified investment in solar heat collecting and absorbing equipment. A ratio of net useful solar capacity to gross solar input below 100 per cent indicates that the overall thermal efficiency of the plant is degraded by adding solar input (but not in direct proportion to this ratio) and that, in effect, some of the solar heat input and investment in collecting and absorbing equipment is wasted. A high ratio of heat transfer surface to minimum aperture area means not only that the absorber is less compact and has more heating surface, but also that more problems will probably arise in designing the absorber to attenuate the high incident solar flux. To derive this ratio, an incident solar flux of 700,000 Btu/hr-ft 2 at the absorber aperture was assumed. The ratio of weight of surface and piping weight to useful solar capacity are both indicative of the relative capital investment required per unit of solar-derived power generated. Items 7 through 10 in the table are qualitative comparisons of the design and operating characteristics of the candidate schemes based on engineering judgment and experience. When the solar/fossil-fueled plant is operated at part load, it would be economically preferable to reduce the fuel consumption as much as possible by maintaining the maximum possible solar input. The limitation on this capability is indicated by item 7 in the table. Analysis of thermal transients is a complex problem which is beyond the scope of this study. However, for a
R. J. ZOSCHAKand S. F. Wu
304
Table 5. Summaryof candidatesolarheatabsorbingschemes SH*
EVAP* & SH*
RH*
AH*
AM* & FWH
67
79
36
i00
200
108
15
27
27
27
16
ii
17
68
i00
i00
i00
i00
50
76
2.3
4
7
5.3
16
59
59
160
180
780
460
490
800
800z6,140
3,300
2,200
9,800
Scheme
FWH*
Solar-heated component duty (%) Reference plant component duty
i00
Useful solar capacity Ref. plant capacity (1885 MWt) Net useful solar capacity Gross solar input
(%)
(%)
Heat transfer surface M i n i m u m aperture area Weight of surface Useful solar capacity
EVA/'*
(ib/~t)
Piping weight (ib/MWt) Useful solar capacity Utilization of solar input at p a r t - l o a d
ll,O00
Less Limited Limited
Less Limited
4,000
16,000
15,000
Least Limited
Limited
Highly Limited
Limited
Relative thermal transient severity
Low
Low
High
Moderate
High
High
Relative design complexity
Low
High
High
Moderate
High
Low
Moderate
Relative operation and control difficulty
Low
High
High
Moderate
High
Low
Moderate
rough comparison the relative severity of thermal transients shown in the table for each candidate component has been estimated based on the operating temperature, the ratio of weight of surface to solar input and the thermal capacitance of the contained fluid. The final two items in the table were judged based on the design, operating and control aspects discussed previously in the section on candidate methods of solar input. Using the comparative statistics in Table 5, a selection of the preferred scheme can be made by a process of elimination. The SH*, despite its high useful solar capacity, is unquestionably eliminated because of its high surface and piping weights, severity of thermal transients, relative design complexity and difficulty of operation and control. The RH* suffers from all the disadvantages of the SH* plus a low useful solar capacity and limited part load solar input, and is therefore easily eliminated. The AH* is relatively attractive from the standpoint of design, operation and control, but it can be eliminated on the basis of its very low useful solar capacity, highly limited part load solar input and high weights of surface and piping. Although the AH* & FWH has a higher useful solar capacity and somewhat less limited part load solar input than the AH*, these gains were obtained at the expense of extra equipment and more design, operating and control problems. It still has essentially the same drawbacks as the AH* and is therefore eliminated. Despite the high useful solar capacity and very low weight of surface and piping, the EVAP* can be eliminated on the basis of the complex design and difficult operation and control of the system. This narrows the field down to two remaining alternatives, the FWH* and the EVAP* & SH*. The FWH* appears very desirable because of its low surface and piping weights, low thermal transient severity, simplicity of design and ease of operation and control. However, it has a low useful solar capacity and limited part load solar
High
input. The EVAP* & SH*, while somewhat less desirable than the FWH* in terms of surface and piping weights, thermal transients, design, operational and control, has the highest part load solar input of all schemes studied and nearly twice the useful solar capacity of the FWH*. Moreover, the EVAP* & SH* can have a considerably higher useful solar capacity than indicated in the table,
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Justified incremental investment in solar heat collecting and absorbing equipment as a function of useful solar capacity and fuel cost. F i g . 9.
Direct input of solar energy to a fossil-fueled steam power plant since the value shown in the table is fixed by the limit of 500 WMt solar input assumed as a base line condition for this study. Since justified capital investment is directly proportional to useful solar capacity, a much larger investment is permitted for the EVAP* & SH*. This is graphically illustrated in Fig. 9 for a range of present-day fuel costs assuming 85 per cent overall plant load factor, 15 per cent fixed charge rate, 88 per cent fossil-fueled steam
305
generator efficiency and an admittedly optimistic full solar input 50 per cent of the time (12 hr/day) over the life of the plant. The much larger investment justified for the EVAP* & SH* tips the balance in its favor as the preferred scheme.
Acknowledgement--The authors gratefully acknowledge the help of the General Electric Company in furnishing heat balance diagrams for the 800 MW steam turbine cycle upon which this study was based.
STUDIES OF THE DIRECT INPUT OF SOLAR ENERGY TO A FOSSIL-FUELED CENTRAL STATION STEAM POWER PLANT* R. J. ZOSCHAK and S. F. Wu Research Division,Foster Wheeler Corporation, 12Peach Tree Hill Rd., Livingston,NJ 07039, U.S.A.
(Received26June 1974) A~traet--Seven possible methods of absorbing solar energy as direct thermal input to an 800 MW, fossil-fueled, central station steam power plant have been studied. Irrespective of method, the solar heat is first collected by an array of flat mirrors and concentrated on a tower-mountedabsorber where it is transferred into the power cycle. The heat absorbing methods studied were heating of feed-water, evaporation of water, superheatingof steam, combined evaporation and superheating,reheatingof steam, air preheating,and combinedair preheating,and feedwater heating. Factors considered were relative capital cost, energy conversion efficiencyand complexityof design, operation and control. Combinedevaporation and superheatingproved to be the preferred method because of its high utilizationof solar energy, relatively low indicated capital cost and only moderate complexity in design, operation and control. Feedwater heatingalso has very desirablecapitalcost, designand operatingaspects, but suffers from the drawbackthat over 30per cent of the solar energyabsorbed is, in effect, lost becauseof degradationof the steam cycle efficiency.
INTRODUCTION
ASSUMED BASE LINE CONDITIONS
One possible way to use solar energy for the production of commercial electric power would be to use it as a supplemental source of thermal input to a conventional, fossil-fueled, central station steam power plant. This would assure capability of the plant to supply electrical power at all times. During periods of cloud cover and at night, load would be carried solely on fossil fuel. During periods of sunshine, fossil fuel firing rate would be decreased, thus saving fuel and reducing stack emissions. The technical and economic evaluation of a combined solar/fossil-fueled plant of this sort is the subject of a current study sponsored by the National Science Foundation.f The system being considered consists of an array of many flat mirrors arranged and controlled to reflect and concentrate the rays of the sun on a tower-mounted heat exchanger or absorber. The heat thus collected by the absorber would be fed directly into the steam power cycle without any intervening heat transfer fluid, heat storage or topping cycle. The supporting tower might also serve as the stack of the power plant. This is a report on that portion of the study concerned with examining, comparing and evaluating the various possible methods of adding solar heat to a modern, high pressure, high temperature steam cycle. Factors considered in this exercise were relative capital cost as indicated by material requirements, energy conversion efficiency, and complexity of design, operation and control.
To keep the range of the study within reasonable bounds, the following base line conditions have been assumed. (1) The reference fossil-fueled plant is to have a full load electrical output of approximately 800MW and normal full load turbine steam conditions of 2400 psig, 1000F superheat/1000F reheat. Figure 1 is a simplified heat balance diagram of the reference plant. In this diagram the Low Pressure Feedwater Heater (LP FWH) and Intermediate Pressure Feedwater Heater (IP FWH) each represent 3 separate stages of regenerative feedwater heating, so that in the actual plant there are 7 feedwater heaters. (2) Maximum solar thermal input is to be 500 MW. (3) Full load plant electrical output is to remain at 800 MW under combined solar/fossil-fueled operation.
CANDIDATE METHODS OF SOLAR HEAT INPUT
*This paper was presented at the International Solar Energy Society's U.S. Section Meeting held in Fort Collins, Colorado (Aug. 1974). fNSF Grant No. GI-41019. Sheldahl is prime contractor. Subcontractors are Foster Wheeler Corporation and the University of Minnesota. Northern States Power Company is a technical consultant.
Seven candidate methods for absorbing solar heat into the steam cycle have been studied. These are listed below. (1) Feedwater heating--FWH. (2) Evaporation of water--EVAP (3) Superheating of steam--SH (4) Combined evaporation and superheating--EPAV & SH (5) Reheating of steam--RH (6) Air preheating--AH (7) Combined air preheating and feedwater heating-AH & FWH. In the text and figures to follow, an asterisk after an abbreviation denotes the solar heated counterpart of that particular component of the steam cycle. Feedwater heating. The schematic arrangement of incorporating a Feedwater Heater (FWH*) in a fossilfueled steam plant is shown in Fig. 2. During solar input,
297
298
R.J. ZOSCHAKand S. F. wu
~
5.5x106#
2535P
IO00F
"
561P
5,02xi06# 815~wE
® 635F
-
553~" 3150P 242F
!
Fig. |. Simplifiedheat balance diagram of reference steam power plant. Figures shown are for operation at 5% overpressure at turbine throttle.ff, pressure, psia; F, temperature,°F; #,flow, lb/hr.
FUEL
STACK~ AIk
~ m I ._~---___.-~
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Fig. 2. Simplifiedheat balancediagramshowingsolar-heatedfeedwater heater in reference plant.
extraction steam from the turbine to the high and intermediate pressure FWH's is cut off except for a small amount to keep the heaters at hot standby condition to minimize thermal transients when solar heating ceases. Feedwater is pumped to the FWH* as indicated by the broken line in the diagram. After being heated in the FWH*, the water is fed back to the economizer section of the fired boiler. The fuel firing rate is reduced by the amount equivalent to that supplied by solar input to the FWH*, and the furnace portion is operated at part load. The amount of feedwater circulating to the FWH* can be chosen to produce either the same flow rate at the turbine throttle or the same flow rate at the condenser inlet as that of the reference cycle. The turbine-generator output will be increased for the former condition since no steam will be extracted from the high and intermediate pressure stages of the turbine. For the latter condition, less steam will pass through the higher pressure stages of the turbine and the output will be reduced. Obviously, these two different operating conditions would require two different turbine designs. In order to compare the different schemes on an equal plant output basis in this study, the FWH* flow has been set at a rate which results in the same turbine-generator output as in the reference cycle. Without regenerative feedwater heating by the conventional High Pressure
(HP) and Intermediate Pressure (IP) FWH when the FWH* is in service, the plant thermal efficiency is reduced. Evaporation. In this arrangement, the Evaporation of water (EVAP*) is put in parallel with the fired EVAP as depicted in Fig. 3. When EVAP* is in operation, the heat absorption in the fired steam generator has to be re-distributed between the partially loaded EVAP and other heat transfer surface. A differential firing scheme can be used for this purpose, and adoption of a twin-furnace design is one well-established way of doing it. One furnace includes superheater (SH), reheater (RH) and economizer of pre-evaporator (ECON) sections and the other houses primarily the EVAP section. The amount of feedwater by-passed to EVAP* is proportional to the thermal capacity replaced. The balance of the feedwater circulates through the conventional EVAP and the firing rate of the EVAP furnace is reduced accordingly. During the operation, the firing rate of the other furnace is kept unchanged. A special unconventional arrangement of heat transfer surface is required in the fired steam generator in this scheme. Superheating. In this arrangement, the SH* is put in parallel with the fired SH as shown in Fig. 4. The amount of saturated steam required for SH* is drawn from the boiler drum and the remaining portion is fed to its fired
Directinput of solar energyto a fossil-fueledsteampowerplant
299
3.69xI06~
680F
I I I
, '
LSJ: FUEL~I
STACK
~. T~_4 . AIR
Fig.3. Simplifiedheat balancediagramshowingsolar-heatedevaporatorin referenceplant.
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Fig.4. Simplifiedheat balancediagramshowingsolar-heatedsuperheaterin referenceplant.
counterpart. The fuel firing rate is reduced in the furnace to account for the reduced heat absorption of the SH section. To re-distribute the heat absorption among the surfaces in the furnace, the convective SH and RH sections are arranged in parallel passes and the flue gas crossing each section is apportioned by control dampers. To maintain the reheat steam temperature constant during operation, control dampers are adjested so that more flue gas is directed to the RH section. Superheater outlet steam temperature is controlled by a spray type desuperheater wherein feedwater is evaporated to cool the steam. Flow of feedwater to the boiler is consequently reduced by the amount by-passed to the superheat control device. The operation of this scheme is relatively complicated due to the dual control involved. Combined evaporation and superheat!ng. Figure 5 depicts this combined arrangement. The total solarassisted capacity (500MWt) is distributed between EVAP* and SH* in the same proportions as the heat absorption in their counterparts in the fired unit. For this scheme, no special surface arrangement is needed in the fired steam generator. Superheated steam temperature
can be controlled by water spray and the absorption of the RH in the furnace is maintained at its proper rate by biased-flow damper control as described previously. Since the percentage of the combined evaporation and superheating capacity replaced by solar heating is much lower than the percentage of capacity replaced when solar heating is applied to each component alone, fewer operation and control problems are expected for this combined type of arrangement. Reheating. This arrangement is shown in Fig. 6. During operation, the steam to be reheated is by-passed to RH* except for a small amount which is passed through the fired RH to keep it from overheating. In the furnace, the RH is arranged in parallel passes with the SH so that flue gas flow can be controlled by dampers to by-pass RH completely when RH* is in service. The feedwater flow to the boiler remains at the full-load rate, but the fuel firing rate is reduced to the level needed to supply the heat input requirements of the SH, EVAP and ECON. A relatively large volume of high temperature steam has to be transported back and forth between the power plant and solar receiver for this scheme, which makes the piping
300
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Fig. 6. Simplifiedheatbalancediagramshowingsolar-heatedreheaterin referenceplant.
line expensive. Also, circulating steam through the idling RH to prevent overheating would further reduce the already limited capacity of the RH*. Air preheating. In this arrangement, the Air Preheating (AH*) is put in parallel with the air preheater section of the steam generator as shown in Fig. 7. To maximize the capacity of the AH*, the air required for combustion is preheated to 1000°F during the period of solar input, which is a considerably higher temperature than normally used in the conventional mode of operation. With the conventional AH not in service, the energy in the high temperature flue gas discharged directly to the stack is not recovered and the efficiency of the steam generator is reduced. Consequently, not all of the solar input is useful to the cycle. Fuel firing rate is cut by an amount equal to the net useful thermal capacity gained by the AH*, while the other components in the power plant are not affected by this operation. When the air supply fan is located upstream of the AH* as shown in the figure, the fluid side heat transfer surface has to be enclosed. If the fan is located downstream of the AH*, it must be designed to
withstand high air temperature and possible rapid temperature changes. Another conceivable alternative is to locate the fan downstream of the conventional AH where the gas temperature is always relatively low. However, this would require extra reinforcement of the steam generator casing to withstand the high negative pressure differential, and a more costly conventional AH of tubular design to prevent air leakage. In any event, high temperature air ducting is required and the burners must be specially designed for high temperature. Combined air preheating and .feedwater heating. This combined arrangement is a modification of the previous approach of rising the AH* alone. As shown in Fig. 8, an additional heat exchanger is installed in the flue gas pass downstream of the ECON so that stack gas can be used to heat feedwater when the conventional AH is taken out of service. This FWH is arranged in parallel with the conventional intermediate pressure feedwater heaters and takes over part of their thermal duty during solar input periods. Except that it is not solar-heated, its mode of operation is similar to that of the FWH* as described in
Direct input of solar energy to a fossil-fueledsteam power plant
301
AIR
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the first scheme. To maintain the same turbine-generator output, the feedwater flow rate through the stack FWH must be reduced below what would normally flow through the regenerative IP FWH because of the reduced steam extraction from the turbine. The turbine cycle efficiency is reduced when the regenerative IP FWH is out of service, but the heat recovery from the stack gas greatly improves the boiler efficiency. The net result is an increase in the net utilization of solar input as compared to using AH* alone. CO~IP~'~ATIVESTUDY~ 0 ~SULTS The 7 candidate schemes were studied to identify their design and operating conditions, to estimate their physical dimensions, and to evaluate the trade-offs among them. The pertinent parameters selected for this study were divided into four groups as presented in Tables 1-4. Table 1 lists the thermal capacity and operating conditions of each arrangement. The capacity of each com-
ponent is based on operation of the reference plant cycle at 5 per cent over-pressure condition as shown in Fig. 1. At this condition the thermal input and electric output of this cycle are 1885 MWtt and 815 MWe, respectively. For the combined arrangements such as EVAP* & SH* and AH* & FWH, The results for each component are shown, separated by a slant, wherever appropriate. The net useful solar capacity of each scheme was calculated by subtracting from its solar-heated capacity an amount that accounts for any consequent reduction in thermal efficiency of the plant. From the table we see that only three schemes, EVAP*, SH* and EVAP* & SH*, are capable of absorbing the maximum solar input of 500 MWt and also providing a net useful solar capacity of ?Note that the 1885MWt is the input to the steam turbine cycle or, in other words, the output of the steam generator. The total thermal input to the plant is 1885MW divided by the efficiencyof the steam generator.
R. J. ZOSCHAK and S. F. Wu
302
Tablel. Thermalcapacityandoperatingconditionsofsolarheatabsorbingschemes Scheme
EWE*
EVAP*
EVAP* & SH*
SH*
RH*
AH & FWH*
AH*
Compound capacity in the reference fossil plant (Mwt) 425
745
630
1375-
305
200
200/170
Capacity replaceable by solar heating (MWt)
425
500
500
270/230
305
400
400
Net useful solar c a p a c i t y (MWt)
290
500
500
305
200
305
Fluid temperature Fluid p r e s s u r e
(°F)
240-525
(psi)
680
3000
500
680-1000
680-1000
2500
2750
2750
640-1000
80-1000
600
80-1000/240-370
15
15/500
Table 2. Thermal/hydraulic characteristics of solar heat absorbing schemes FWH*
EVAP*
SH*
EVAP* & SH*
RH*
AH*
4.75
14.8
4.37
8.0/2.0
5.02
6.0
6.0/4.53
Fluid mass velocity (1061bZhr-ft)
1.5
1.0
1.0
1.0
0.3
0.02
0.02/1.0
Average fluid velocity (ft/sec)
7.7
ii
56
11/56
102
134
134/5
Fluid heat transfer cosff. (Btu/hr-°F-ft 2 )
2,500
9,000
11000
9,000/i,000
260
15
15/1,700
300,000
180,000
i00,000
44,200
12,000
12,000~8,500
120
20
i00
170
600
4,840
9,500
17,000
5,130/7,820
6
50
6~50
Component Flow rate
(1061b/hr)
Heat flux (Btu/hr-ft 2 ) Fluid film ~ T
(OF}
Heat transfer surface (ft 2 ) A p p r o x . ~ P through the component (psi)
i0
180,000/100,000 20z100
AH* & FWH
800/5
23,600 114,000
25
114,000/68,300
0.6
0.6/12
Table 3. Materials and weights for solar heat absorbing schemes Component
FWH*
Design temperature Design p r e s s u r e
(°F)
800
Weight of surface
RB*
AH*
AH* & FWH
1200
800/1200
1200
1500
1500/800
3000
3000
800
30
30/800
T22
T22
T22/TP304
T22/TP304
T22/TP304
TP304
TP304/T22
0.2
0.2
0.41
0.2/0.41
0.14
0.034
0.034~0.06
275
165
190
165/190
40
5
5/5
45,400
90,000
392,000
229,000
160,000
160,000z645,000
(OF)
(ib)
EVAP* & SH*
800
3000
Tube material
across tube wall
SH*
3000
(psi)
Minimum wall thickness (in) ~T
EVAP*
149,000
Table 4. Approximate piping dimensions and pressure losses of solar heat absorbing schemes EVAP *
Component Pipe flow rate
(106 ibZhr)
FWH*
EVAP*
SH*
& SH*
4.75
3.69
4.37
2.0
20x2
20x2
32x3.4
RH*
AH*
5.02
6.0
Supply line: Pipe O.D. x thickness
(in)
NO. of parallel lines Velocity ~P
(psi)
1
(ft/sec) frictional/gravitational
1
1 39
16xi.6 1
40xi.3 2
16
17
14
87
38/413
31t307
28162
26/307
14/8
20x2
24x2.5
34x5.8
30x5.1
40x2.4
Return line: Pipe O.D. x thickness
(in)
No. of parallel lines Velocity ~P
(ft/sec)
(psi) frictional/gravitational
Total pipe line ~ P (psi) frictional/gravitational Total piping w e i g h t
(103 ib)
1
1
2
1
20
58
62
78
47/-338
71t-62
30/-25
50~-25
85/75
102/245
58/37
950
i,ii0
5,600
ll0x½
3
4
104
243
13/-5
1.2 /0.33
76/282
27/3
1.2/0.33
2,000
4,900
3,000
Direct input of solarenergyto a fossil-fueledsteam powerplant 500 MWt. RH* is limited to a solar input of 305 MWt. FWH*, AH* and AH* & FWH are not only limited to less than 500 MWt input, but also have respective net useful solar inputs which are only about 68, 50 and 76 per cent of their total solar inputs. Table 2 shows the thermal and hydraulic characteristics of the components of the candidate schemes. The fluid mass velocity in each component was chosen so that the fluid velocity and pressure drop through the unit would be within the range normally used in the design of each type of component for conventional power plants. In the evaporator, a recirculation ratio of 4:1 was assumed, which is equivalent to an exit steam quality of 25 per cent by weight. Fluid side heat transfer coefficients were determined from applicable fluid properties and correlations for the respective fluids and flow regimes. The heat flux given for each component except the evaporator is the limiting heat flux at which the calculated mean tube wall temperature closely approaches the design temperature of the tube material. For the evaporator, the heat flux was selected to be conservatively below the critical heat flux for Departure from Nucleate Boiling (DNB) at the assumed pressure, mass velocity and steam quality in the tubes. The heat flux and heat transfer surface given in the table are both based on the inside surface of the tubes. The pressure drop shown is the drop through the heat transfer tubes only and does not include pressure losses in distribution piping and headers. Factors concerning material selection and weights are covered in Table 3. Design preseure is based on the estimated maximum operating pressure to be expected in the component plus a design margin. Design temperature is based on practical upper limits for the materials used under the given operating conditions. To simplify the selection of tube material only two typical materials were considered, one ferritic (ASTM SA-213 T22) for moderate temperatures and the other austenitic (ASTM SA-213 TP304) for high temperatures. In the SH* and RH*, both materials were required for different temperature ranges. Two-inch O.D. tubes were used throughout since experience has shown that this tube size is near optimum for all components. Tube wall thicknesses shown in the table are as calculated and are not adjusted to the next largest standard wall thickness. A maximum design temperature of 1500°F was assumed for the AH*, although the air outlet temperature and air-to-tube AT were taken as 1000°F and 800°F, respectively. In an actual design, this would require a reduced heat flux at the hot end of the exchanger and a higher heat flux at the cool inlet end. Table 4 presents the pipe line dimensions and approximate piping pressure losses for each scheme. Listing for the AH* & FWH has been omitted because it is the same as for the AH*. (The relatively small amount of additional piping required for the stack FWH has been neglected.) Figures for the supply line from power plant to tower-mounted component and for the return line to the power plant are listed separately. Pipe friction loss was estimated by assuming a 1200-ft line length in each direction plus an additional loss of 10 velocity heads (5 for the AH*) to account for losses in valves, bends, entrance and exit. Static pressure change due to gravity head was
303
calculated by assuming a tower height of 1000 ft. Piping associated with the SH* and RH* was sized to give pressure losses roughly comparable to corresponding line losses in the reference plant. Reasonable pressure drops were assumed for sizing the piping for other components. Because of the low allowable pressure drop for RH* and AH* piping, we see that very large pipes are required. It is even conceivable that a tubular supporting tower could serve as a single large duct for the AH*. Comparing Tables 3 and 4 we see that the weight of piping exceeds the weight of the heat transfer surface by more than a factor of 10 for every solar-heated component except the EVAP* & SH*. It might be possible to reduce piping weight by putting the supply pipe within the return pipe so that the inner pipe need withstand only the differential pressure. However, this would introduce problems of insulating the inner pipe to minimize regenerative heat transfer and supporting the inner pipe while accomodating the relative thermal expansion of the pipes. SELECTION OF PREFERRED METHOD
To facilitate comparison of the candidate schemes, the results of the comparative study and other important factors are summarized in Table 5. The first 6 items in the table are significant ratios derived from pertinent values given in Tables 1--4. A low ratio of solar-heated component duty to reference plant component duty is generally indicative of better ability to maintain high solar input at partial plant load and, for the EVAP*, SH*, EVAP* & SH* and RH*, it also indicates fewer problems in designing and controlling the fired steam generator to accomodate the full range of solar input. A higher ratio of useful solar capacity to reference plant capacity means a greater saving in fossil fuel and, therefore, a greater justified investment in solar heat collecting and absorbing equipment. A ratio of net useful solar capacity to gross solar input below 100 per cent indicates that the overall thermal efficiency of the plant is degraded by adding solar input (but not in direct proportion to this ratio) and that, in effect, some of the solar heat input and investment in collecting and absorbing equipment is wasted. A high ratio of heat transfer surface to minimum aperture area means not only that the absorber is less compact and has more heating surface, but also that more problems will probably arise in designing the absorber to attenuate the high incident solar flux. To derive this ratio, an incident solar flux of 700,000 Btu/hr-ft 2 at the absorber aperture was assumed. The ratio of weight of surface and piping weight to useful solar capacity are both indicative of the relative capital investment required per unit of solar-derived power generated. Items 7 through 10 in the table are qualitative comparisons of the design and operating characteristics of the candidate schemes based on engineering judgment and experience. When the solar/fossil-fueled plant is operated at part load, it would be economically preferable to reduce the fuel consumption as much as possible by maintaining the maximum possible solar input. The limitation on this capability is indicated by item 7 in the table. Analysis of thermal transients is a complex problem which is beyond the scope of this study. However, for a
R. J. ZOSCHAKand S. F. Wu
304
Table 5. Summaryof candidatesolarheatabsorbingschemes SH*
EVAP* & SH*
RH*
AH*
AM* & FWH
67
79
36
i00
200
108
15
27
27
27
16
ii
17
68
i00
i00
i00
i00
50
76
2.3
4
7
5.3
16
59
59
160
180
780
460
490
800
800z6,140
3,300
2,200
9,800
Scheme
FWH*
Solar-heated component duty (%) Reference plant component duty
i00
Useful solar capacity Ref. plant capacity (1885 MWt) Net useful solar capacity Gross solar input
(%)
(%)
Heat transfer surface M i n i m u m aperture area Weight of surface Useful solar capacity
EVA/'*
(ib/~t)
Piping weight (ib/MWt) Useful solar capacity Utilization of solar input at p a r t - l o a d
ll,O00
Less Limited Limited
Less Limited
4,000
16,000
15,000
Least Limited
Limited
Highly Limited
Limited
Relative thermal transient severity
Low
Low
High
Moderate
High
High
Relative design complexity
Low
High
High
Moderate
High
Low
Moderate
Relative operation and control difficulty
Low
High
High
Moderate
High
Low
Moderate
rough comparison the relative severity of thermal transients shown in the table for each candidate component has been estimated based on the operating temperature, the ratio of weight of surface to solar input and the thermal capacitance of the contained fluid. The final two items in the table were judged based on the design, operating and control aspects discussed previously in the section on candidate methods of solar input. Using the comparative statistics in Table 5, a selection of the preferred scheme can be made by a process of elimination. The SH*, despite its high useful solar capacity, is unquestionably eliminated because of its high surface and piping weights, severity of thermal transients, relative design complexity and difficulty of operation and control. The RH* suffers from all the disadvantages of the SH* plus a low useful solar capacity and limited part load solar input, and is therefore easily eliminated. The AH* is relatively attractive from the standpoint of design, operation and control, but it can be eliminated on the basis of its very low useful solar capacity, highly limited part load solar input and high weights of surface and piping. Although the AH* & FWH has a higher useful solar capacity and somewhat less limited part load solar input than the AH*, these gains were obtained at the expense of extra equipment and more design, operating and control problems. It still has essentially the same drawbacks as the AH* and is therefore eliminated. Despite the high useful solar capacity and very low weight of surface and piping, the EVAP* can be eliminated on the basis of the complex design and difficult operation and control of the system. This narrows the field down to two remaining alternatives, the FWH* and the EVAP* & SH*. The FWH* appears very desirable because of its low surface and piping weights, low thermal transient severity, simplicity of design and ease of operation and control. However, it has a low useful solar capacity and limited part load solar
High
input. The EVAP* & SH*, while somewhat less desirable than the FWH* in terms of surface and piping weights, thermal transients, design, operational and control, has the highest part load solar input of all schemes studied and nearly twice the useful solar capacity of the FWH*. Moreover, the EVAP* & SH* can have a considerably higher useful solar capacity than indicated in the table,
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Justified incremental investment in solar heat collecting and absorbing equipment as a function of useful solar capacity and fuel cost. F i g . 9.
Direct input of solar energy to a fossil-fueled steam power plant since the value shown in the table is fixed by the limit of 500 WMt solar input assumed as a base line condition for this study. Since justified capital investment is directly proportional to useful solar capacity, a much larger investment is permitted for the EVAP* & SH*. This is graphically illustrated in Fig. 9 for a range of present-day fuel costs assuming 85 per cent overall plant load factor, 15 per cent fixed charge rate, 88 per cent fossil-fueled steam
305
generator efficiency and an admittedly optimistic full solar input 50 per cent of the time (12 hr/day) over the life of the plant. The much larger investment justified for the EVAP* & SH* tips the balance in its favor as the preferred scheme.
Acknowledgement--The authors gratefully acknowledge the help of the General Electric Company in furnishing heat balance diagrams for the 800 MW steam turbine cycle upon which this study was based.