hts thermal storage peaking plant
TRANSCRIPT
ORNL/Sub-4188/2 GA-A14160
HTS THERMAL STORAGE PEAKING PLANT
by D. L. VRABLE and R. N. QUADE
Prepared under Subcontract No. 4 1 8 8
for the Oak Ridge National Laboratory U.S. Energy Research and Development Administration
APRIL 1977
19
DISTRIBUTION OF THIS DOCUMENT IS UNLIMl
NOTICE - — This report was prepared as an account of work sponsored by the United States Government .
Neither the United Stales nor the United States Energy Research and Development Administration, nor any of their employees, nor any of their contractors, subcontractors, or their employees, makes any warranty, expiess or implied, or assumes any legal liability o r responsibility fo r the accuracy, completeness or usefulness of any informat ion , appara tus , p roduc t or process disclosed, or represents that its use would not infr inge privately owned rights.
ORNL/Sub-4188/2 GA-A14160
HTS THERMAL STORAGE PEAKING PLANT
by D. L. VRABLE and R. N. QUADE
Ihfc report was prepared an account of Wotk , sponvtred hy the United StatM (juvernnicnl. Netihei the United SuKj n<u Unjied Stales PheW Rcvacth am) Development Administration, m>r any of thru employee, nor any of iheu contractor, subronlrachm, or their employees, make* any i warranty, expieu or implied, or a»umei any lef-al I liability oi responsibility fni the accuracy, uimpletsnet* I or uvfulnew of any information, apparatus,produtf or [ pi««.eji diuiowd, or icpieveftts that if* u<c would m>r [ infringe privately owned tiRhts,
Prepared under Subcontract No. 4 1 8 8
M c r J ° r 0 a k R i d § e N a t l ' °nal Laboratory U.S. Energy Research and Development Administration
ATOMIC PROJECT 3 2 3 7 APRIL 1977
DJS.TRIBU.TJUN Qo IW5 LXiCUJ.' MT !S UNLIMITED** ['> V'
TABLE OF CONTENTS
Section Page 1 ABSTRACT 1 2 INTRODUCTION AND SUMMARY 2
3 STUDY GROUNDRULES AND DESIGN REQUIREMENTS 7 3.1 Study Groundrules 7 3.2 Design Requirements 8 3.2.1 Utility Load Requirements 8 3.2.2 Operation Requirements 11
3.2.3 Structural Requirements 14
4 SYSTEM PARAMETER SELECTION 15 4.1 Salt Temperature Selection 15 4.2 Cycle Parameters for the Reference Design . . . 16 4.2.1 Primary Helium System Parameters 20 4.2.2 Intermediate Loop Parameters 22
4.2.3 Base Load Steam Cycle Parameters 22
4.2.4 HTS Loop and Peaking Plant Cycle Parameters . . 27
5 REFERENCE PLANT DESIGN 32 5.1 System Configuration 32 5.2 Reactor System 34
5.3 Intermediate Helium System . . . 34 5.4 Base Load Steam System 38 5.5 Thermal Energy Storage System 42 5.6 Peaking Load Steam System 46
6 REFERENCE COMPONENT DESCRIPTIONS 49 6.1 PCRV and RCB ' 49 6.1.1 PCRV Structure 49
6.1.2 RCB Structure 53
6.1.3 Thermal Barrier 56 6.2 Helium Circulators . 57
6.2.1 Circulator Design Requirements 57
iii
TABLE OF CONTENTS (Continued)
Section Page
6.2.1.1 Primary Loop Circulator 57
6.2.1.2 Secondary Loop Circulator 60
6.2.2 Circulator Configuration . . . . . . 62
6.2.3 Circulator Installation 66
6.2.3.1 Primary Loop Circulator . . . . . 66
6.2.3.2 Secondary Loop Circulator . . . . 66
6.3 Heat Exchangers . 68
6.3.1 IHX 70
6.3.2 Helium/Steam Heat Exchangers 74
6.3.2.1 Steam Generator Design 75
6.3.2.2 Reheater Design 79
6.3.3 Helium/HTS Heat Exchanger . . . . . 81
6.3.4- H T S / S t e a m Heat Exchangers 85
6.3.4.1 Steam Generator 86
6.3.4.2 Reheater 88
6.4 Salt Loop Components 89
6.4.1 Heat Storage Tank Design 89
6.4.2 HTS Piping System 91
6.4.3 Valves 93
6.4.4 Pumps 96
6.5 HTS Peaking Plant Plot Plan 96
> 6.5.1 Reactor Service Building 98
6.5.2 Control Building 99
6.5.3 Penetration Building 99
6.5.4 CACS Power Supply Building 100
6.5.5 Diesel Generator Building 100
6.5.6 Turbine Buildings 100
6.5.7 Helium Storage Building 100
7 PLANT OPERATION 101
8 PARAMETRIC STUDIES 104
8.1 Alternative Reactor Sizes 104
iv
TABLE OF CONTENTS (Continued)
Section Page 8.2 Variation of Duty Cycle 104
9 ECONOMICS 113 9.1 Reference Plant Cost 113 9.2 Alternate Plant Costs for 850 MW(t) and
3000 MW(t) Plants 117
9.3 Cost Comparisons 117
10 CONCLUSIONS AND RECOMMENDATIONS 123
References 126
v
LIST OF FIGURES
Figure Page 1 Weekly load variation for a large Eastern utility . . . . . 9
2 Weekly load variation for a Southwestern utility 10
3 Categorization of peaking, intermediate, and base load 12
4 Daily load variation for a large Eastern utility 13 5 Rate of nitrogen evolution from HTS vs temperature. . . . . 17 6 Tradeoff study of the heat transfer salt temperature vs
differential change in the cost of electricity 19 7 Cycle parameters for the primary helium system 21
8 Cycle parameters for the intermediate helium system . . . . 23
9 Base load plant (English engineering units) (Sheet 1 of 2) 24 Base load plant (SI units) (Sheet 2 of 2) 25
10 Temperature profiles for the salt/steam generator and reheater 28
11 Peaking plant cycle diagram (English engineering units) (Sheet 1 of 2) 30
Peaking plant cycle diagram (SI units) (Sheet 2 of 2) 31
12 HTGR/HTS thermal peaking plant 33
13 Major components and systems of the HTGR NSS 35 14 Flow diagram for the primary and intermediate loop helium
systems 36
15 Layout drawing of PCRV configuration 51 16 Reactor containment building 55
17 Schematic layout of the HTGR/HTS reactor indicating nominal loop temperatures 59
18 Circulator conceptual design for 4000 M(t) SC-HTGR (4 loop) plant ClOOO MW(t)3 circulator rating 63
vi
LIST OF FIGURES (Continued)
Figure Page
19 View showing primary circulator installation in PCRV . . . 67 20 View showing installation of secondary loop circulator
in steel pressure vessel 69
21 Conceptual design of the intermediate helium heat exchanger 73
22 Schematic of the helium/steam generator 77 23 Conceptual helium/steam reheater design . . 80 24 Conceptual design of helium/salt heat exchanger 83 25 HTGR/HTS thermal storage peaking plant plot plan 97 26 Primary and secondary helium cycle parameters for a
3000 MW(t) reactor system 105
27 Base load plant with 3000 MW(t) reactor system (English engineering units) (Sheet I of 2) 106 Base load plant with 3000 MW(t) reactor system (SI units) (Sheet 2 of 2) 107
23 Primary and secondary helium cycle parameters for an 850 MW(t) reactor system 108
29 The peaking plant power available for variations in the daily duty cycle 110
30 Base load power variation with daily duty cycle with a constant peaking plant output of 1009 MW(e) Ill
31 The relationship between the salt inventory and variations of the duty cycle 112
32 Peaking power cost for variation of the base load power credit 118
33 Power cost comparison, HTGR-HTS and gas turbine 122
vii
LIST OF TABLES
Table Page 1 Preliminary cost tradeoff study for hot salt
storage temperature 18
2 Adjusted auxiliary power requirements for the base load and peaking plants 26
3 Intermediate loop He piping summary 39 4 Base load steam system components 43 5 Peaking load steam system components 47 6 PCRV design data 52 7 Liner and concrete temperature limits 58 8 Helium circulator design conditions 61 9 Helium circulator conceptual design details for
1000 MW(t) rating (HTGR unit) 64 10 IHX design summary 71 11 He Hum-to-steam heat exchanger design summary 76 12 Peaking plant hiilium-to-HTS heat exchanger 82 13 Salt/steam heat exchanger design summary 87 14 Heat storage tanks 90 15 Heat-transport loop materials '92 16 Compatibility of materials with anticipated chemical
environments 94 17 Piping and fittings 95 18 Shutoff valves 96 19 Pump performance 98 20 Cost breakdown for the 2000 MW(t) peaking plant study . . . 114 21 Reference 2000 MW(t) peaking plant cost of electricity . . 115 22 Assumptions for estimating power costs 116 23 Cost comparison of 850 MW(t), 2000 MW(t), and 3000 MW(t)
peaking plants 119 24 Capital cost of energy storage system 121
viii
1. ABSTRACT
At the request of Oak Ridge National Laboratory, a study was conducted to evaluate the use of the General Atomic Company (GA) high-temperature gas-cooled reactor (HTGR) to generate intermediate load and peak power by using heat-transfer salt (HTS) thermal storage. The study investigated the technical and economic aspects of the conceptual design to confirm the system's viability for this application. Using a reference thermal power level of 2000 MW(t), preliminary system analysis was performed and a selec-tion of the system configuration was completed.
The approach to the study consisted of establishing a reference point by selecting a peaking capacity, duty cycle, and the thermal power split between the base load and peaking plant.
The reference point design served as a basis for evaluating the tech-nical and economic aspects of the system. Using the same salt storage loop and conversion system, the evaluation was extended to larger and smaller reactor sizes to assess economic attractiveness.
Conceptual design data are provided for the major plant components. Preliminary cost estimates for the plant design and power costs were performed for the selected reference design.
This report was prepared for ORNL in fulfillment of Supplemental
Agreement No. 4 to Subcontract 4188.
1 1
2. INTRODUCTION AND SUMMARY
It is becoming increasingly important in the electric utility industry to relieve the dependency on petroleum-based fuel by shifting to coal and nuclear fuel. Rising capital costs also make it important to obtain higher utilization of installed capacity. Both of these needs can be met by utilizing thermal energy storage for intermediate and peaking power coupled to a base-loaded nuclear power plant. A concept for coupling an HTGR with a thermal storage scheme for efficient, low-cost intermediate and peaking power is discussed in this report.
The load demand on an electric utility grid undergoes significant variations on a daily and weekly basis. These demands are broadly cate-gorized as base load, operating an average of 12 to 24 hr/day; intermediate load, averaging about 4 to 12 hr/day; and peaking load, operating up to about 4 hr/day. Since the available generating capacity must meet the peak demands, a significant portion of that capacity is idle during off-peak periods.
The intermediate and peak power requirements are generally met by
using smaller increments of capacity representing lower capital costs and
higher operating costs, such as steam plants fired by natural gas, oil,
and coal (particularly older units), and in recent years, gas turbines.
The base-load demands are generally met by the newer, larger fossil-fueled
plants and by nuclear plants, both of which represent relatively large
capital investments and have lower operating costs. Sharply increased
fossil fuel prices havs placed a heavy economic penalty on the intermediate
and peaking load plants and have created an incentive for increased use
of baseload plants. This implies that a means for energy storage must be
available for storing energy when the base-load capacity exceeds the demand
and supplying this energy when the demand exceeds the base-load capacity.
To date, only pumped hydrostorage has been available for this use.
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Recently, a number of different concepts utilizing thermal energy storage (TES) have been identified and conceptual plant design studies initiated to prove their feasibility and technical merit. One such con-cept is to store off-peak power in pressurized hot water in an underground cavity (Ref. 1). This stored hot water can be flashed to steam and directed to a peaking turbine facility for generation of additional peaking power. References 2 and 3 discuss a TES system using hot oil in which to store off-peak thermal energy. Storage is achieved by increasing the steam extraction from the turbine during off-peak periods, resulting in a drop In power output. The steam heats oil through heat exchangers as the oil passes from cold tanks to hot tanks. During the peak demand periods, the oil flow is reversed and the stored heat is used to preheat the boiler feed-water, releasing the extraction steam to do useful expansion work in the turbine, and thus raising the power production. While these electrical output changes are taking place, the power plant's boiler runs under con-stant conditions. A similar scheme is also being studied in France by Electricite de France in conjunction with several companies (Ref. 4).
Compressed air storage systems have also been considered to meet utilities' peak energy demands. References 5 through 13 discuss the various studies utilizing this method. In general, the compressed air energy storage (CAES) plant operates in a utility network by utilizing low-cost electrical power (during low-demand periods) to power large electric motors driving air compressors. After cooling and dehumidification, the compressed air is injected into an underground storage chamber. When the load demand is high, this stored compressed air is used in a turbine expander coupled to an electric generator for additional peak power. In this concept, the air must be heated before expansion in the turbine, by either burning fuel or using stored heat.
Another study (Ref. 14) investigated the concept of retrofitting feed-
water heat storage to steam electric power stations. Results of this study
indicated, however, that because of the high capital cost and long constru-
ction downtime, the retrofitting concept is not economically feasible.
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The various concepts of TES presented thus far all have relatively low value of thermal storage efficiency. Thermal storage efficiency is defined as the ratio of the efficiency of the conversion from heat source to bus bar via the storage system to that which can be obtained without the storage system. And many of the concepts required certain geological considerations, which Impact the siting of the TES system.
In the future, it is expected that an increasing portion of electri-cal demand will be met by nuclear plants, taking advantage of the lower fuel costs and the large energy resource represented by domestic supplies of nuclear fuel. The operating characteristics of available nuclear plants make them particularly well suited to largely base-loaded operation.
Past studies of the use of nuclear power for the intermediate and peak load categories have not shown them to be particularly attractive. The economics (i.e., high capital costs and low fuel cycle costs) and operating characteristics for part-time operation were not advantageous, and storage schemes either had unacceptably low efficiency or required significant technological development.
It was recently proposed by ORNL that an HTGR coupled with a thermal
storage system utilizing an 11TS was an attractive way to generate intermediate/
peak power using a base-loaded nuclear plant. This approach was an outgrowth
of a study done by GA on a nuclear process heat plant for petroleum refinery
applications (Ref. 15) in which HTS was found to be the most advantageous
means of transporting heat over long distances. Further study by ORNL (Ref.
16) and GA (Ref. 17) has indicated that the HTS TES concept, using existing
technology, indeed offers a means of supplying peak power at a cost lower
than other alternatives.
The approach used the HTGR design developed for steam-cycle electric
power production, which is an excellent match to the desired salt con-
ditions. A portion of the thermal output is used directly to generate
superheated steam for continuous operation of a conventional turbine
U 1
generator to produce base-load power. The remaining thermal output is used constantly to heat a conventional HTS, such as DuPont HITEC, which is stored in a high-temperature reservoir [538° to 593°C (1000° to 1100°F)]. During peak demand periods, the salt is circulated from the high-temperature to a low-temperature [204° to 316°C (400° to 600°F)] reservoir through steam generators that power a conventional steam-cycle plant. The period of oper-ation can vary, but may typically be the equivalent of about 4 to 8 full-power hr each day. The system can be tailored to the utilities' demand conditions by varying the base-load level and period of operation of the peak-load system. For example, a 2000 MW(t) HTGR could generate approxi-mately 400 ;tW(e) base-load and an additional 1000 MW(e) peaking power for an 8-hr period.
The concept has two basic advantages: 1) the overall efficiency of
the storage and conversion systems is high, and 2) the reactor operates
continuously at full-load conditions, taking full advantage of the lower
fuel costs. The reactor core outlet temperature for this application is
the same as that for the standard steam-cycle design eliminating the need
for new reactor core technology development.
Preliminary cost evaluations have indicated the unit cost of peaking power for 8 hr/day operation would be approximately 45 mills/kW-hr including capital charges (based on July 1976 dollars). This appears clearly com-petitive with fossil-supplied power, which is in the range of 45 to 60 mills/kW-hr.
The present study results in completion of a conceptual plant design
and layout of the HTGR plant with a salt storage-peaking power system. The
conceptual design is based on a reference point power level of 2000 MW(t),
of which 850 MW(t) is utilized continuously by the storage system and the
remainder by the base-load portion of the plant. Major plant component
conceptual designs have been carried out and are presented in this report.
5 1
Results of this study indicate that the HTGR is a practical way to
supply peaking power, because, it is economically attractive and encompasses
" state-of-the-art technology.
The use of an intermediate loop to remove heat from the reactor system is practical; however, further investigation of the graphite/salt reaction may result in the elimination of the intermediate loop. The HTS is a practical low-cost heat transport fluid, however recent surveys of other salt mixtures, have indicated that draw salt may result in lower HTS re-placement costs (Ref. 18). Also techniques to retard the thermal decom-position of the HTS mixture should be investigated.
Although the temperature, configuration, and primary coolant fluid of the HTGR provide a particularly good match, the salt storage concept application is not limited to the HTGR. It can be applied to other heat sources such as coal-fired fossil or the liquid metal fast breeder reactor (LMFBR) and gas-cooled fast reactor (GCFR) nuclear plants, as well as solar systems having similar high-temperature capability. General Atomic has also studied HTS for solar energy storage in conjunction with its fixed-mirror solar concentrator system program.
6 1
3. STUDY GROUNDRULES AND DESIGN REQUIREMENTS
3.1 STUDY GROUNDRULES
The basic objective of the HTGR-HTS thermal storage system was to carry out sufficient conceptual design definition and plant layout to evaluate the technical viability and to permit preliminary cost estimates. Optimization of the system's performance and cost are not within the scope of this study. The following groundrules were established to minimize the parametric studies and maximize the design definition of a selected reference plant.
J.. The reference heat source will be a 2000 MW(t) steam-cycle HTGR.
Existing standard HTGR technology will be assumed, including a
reactor core outlet temperature of 760°C (1400°F).
2. The reference peaking power capacity and salt loop size will be based on continuous use of 850 MW(t) from the reactor.
3. An intermediate helium loop arrangement will be used. The
helium/helium heat exchanger will replace the steam generator
in the PCRV. The salt storage loop and power-conversion
equipment will be located outside the PCRV.
4. Where'T->- possible, component designs from previous studies (i.e., process heat design studies} should be utilized to mini-mize component design efforts.
5. The molten salt to be used will be the KN0 3/NaN0 2/NaN0 3 eutectic mixture (Ref. 19).
7 1
6. A nitrogen cover gas should be assumed for the HTS to prevent undesirable reactions of the salt with contaminants in the air and to reduce salt decomposition.
7. The use of separate storage reservoirs will be assumed for this
study. The use of thermocline tanks or rock storage beds will not
be considered now, even though future studies may prove the merit
of such approaches.
8. The base-load turbine inlet conditions shall be consistent with
current HTGR design conditions of 510°C/16.5 MPa (950°F/2400
psia).
3.2 DESIGN REQUIREMENTS
3.2.1 Utility Load Requirements
The initial step in the HTGR/HTS thermal storage system was to define
a representative duty cycle on which to base the reference design. The
load demand on an electric utility grid undergoes significant variations
on a daily, weekly, and seasonal basis. These load variations also depend
on the geographical location of the utility, the local meteorological con-
ditions, and the region's industrialization. Figures 1 and 2 illustrate
the daily fluctuations for two different power utilities. Figure 1 repre-
sents typical load variations for a large eastern utility, and Fig. 2
represents load data for a smaller southwestern utility. The large differ-
ences in load variations between the different utilities indicate that the
thermal storage peaking system must be capable of being tailored to the
specific needs of the utility. In addition, the storage system peaking
plant must be adjustable to the weekly and seasonal load variations.
The approach in this study is to select representative utility grid
demand characteristics to serve as a model for determining the operating
characteristics and comparisons with fossil-fuel alternatives. The larger
eastern utility load characteristics were used as the base.
8 1
ID
IX) o 3000
2000
1000
I 48 72 96
TIME (HR) 120 144 168
GA-A14160
Fig. 1. Weekly load variation for a large Eastern utility
2000
1800
1600
1400
1200
u £ 7000 < o
| 800 < cc UJ % BOO u
400
200
0 J_ X _L 24 48 120 72 96
TIME (HR)
Fig. 2. Weekly load variation for a Southwestern utility
144 168 GA-A14160
The electric demand levels can be broadly categorized as base load, intermediate load, and peaking load. Figure 3 illustrates the relative portions of these demand levels. The load demand level with which the HTGR/HTS thermal storage concept will compete is the peaking and inter-mediate power levels. The thermal storage peaking system should be designed, therefore, to operate in a duty cycle range of 4 to 12 hr. In selecting the appropriate duty cycle, two considerations must be made: the total amount of energy (MW(e)-hr) available for peaking, and the shape of the load demand curve. The MW(e)-hr of peaking energy depends on the amount of ther-mal power input to the storage system and the efficiency of the thermal stor-age system and power-conversion systems. Using the groundrule of 850 MW(t) thermal input to the storage system and estimating the efficiencies of storage and conversion equipment results in approximately 8000 MW(e)-hr of available peaking energy. This means a daily duty cycle of roaghly 667 MW(e) of power for a 12-hr period, or 2000 MW(e) of power for a 4-hr period. For this study a full power duty cycle of 8 hr was selected. This selection does not limit the actual time the system can be in operation, but it does limit the maximum output to approximately 1000 MW(e) of peaking power. Longer periods of operation cari be accommodated by a reduction in peaking power output.
3.2.2 Operation Requirements
The operational requirements of the HTGR/HTS thermal storage peaking system are to maintain a continuous base load power output and cycle the peaking capacity power during the peaking demand periods. The peaking plant is designed as a load-following plant, capable of following the daily load demand variations. Figure 4 shows a representative daily load demand, with the cross-hatched area indicacing the 8000 MW(e)-hr of power capacity supplied by the peaking plant. It is expected that the short-term load variations (1 to 2 hr) will be followed by the fast-acting gas turbine peaking units on the electrical grid.
1 1
ELECTRIC 7 0
D E M A N D
U ) 60
20 40 60 TIME (%)
80 100 GA-A14160
Fig. 3. Categorization of peaking, intermediate, and base load
12 1
11000
U)
1000C
9000
8000
7000
t 6000 o < a. < u
8000 MW(e)-HR PEAKING CAPACITY
(3 Z
ce LU z UJ C5
5000
4000
3000
2000
1000
_L J_ J_ _L I _L 8 16 18 10 12 14
TIME (HR)
Fig. 4. Daily load variation for a large Eastern utility
20 22 24 GA-A14160
The HTGR/llTS thermal storage peaking system is expected to have a fast startup time. During off-peak periods, the critical components will be thermally conditioned to minimize the startup time.
The operation of the peaking can accommodate varying load demands. The peak peaking power output is limited to approximately 1000 MW(e); however, extended operation at lower power levels is available.
The base load is isolated from the peaking steam systems. This allows the base load portion of the plant to operate while the peaking portion of the. plant is down for maintenance.
3.2.3 Structural Requirements
The HTGR/HTS thermal storage peaking system is based on the current HTGR technology for the reactor system. The reactor system is designed with the same structural requirements as the HTGR reactor.
The design life of the HTGR/llTS thermal storage peaking system is
40 yr. Any components for which the design life is not reasonably obtain-
able are located so they can be easily removed.
14 1
4. SYSTEM PARAMETER SELECTION
4.1 SALT TEMPERATURE SELECTION
The initial step in the thermal storage study was to select the appropriate hot salt storage temperature. This parameter is important because it directly affects the peaking power conversion efficiency, salt loop component sizing and costs, and the salt replacement (from thermal decomposition) cost. Increased storage temperature increases the peaking plant's thermal efficiency and reduces the salt inventory; however, the thermal decomposition rate increases and results in additional salt replace-ment costs. The approach used to select the proper salt storage temperature was determined by preliminary cost tradeoff studies for various salt tem-peratures .
The salt inventory was estimated for various hot salt storage tem-peratures . It was assumed in this analysis that the cold reservoir temper-ature remained constant. The salt inventory is inversely proportional to the hot storage temperature, with the assumption of constant thermal heat input.
Information pertaining to the HTS thermal stability was evaluated on an economic basis over the range 510°C (950°F) to 593°C (1100°F). In this range the salt undergoes a slow thermal decomposition, which is largely a thermal breakdown of the nitrite to yield nitrate, alkali metal oxide, and nitrogen:
5 NaN0 2 3 NaNO^ + Na 20 + N 2
This reaction is evidenced by a slow evolution of nitrogen gas. Oak Ridge National Laboratory assembled data from various sources and from these
15 1
plotted a curve of nitrogen evolution rate as a function of salt temper-ature, which is shown In Fig. 5 (Ref. 20). Using these data as our refer-ence, the annual salt replacement for the various temperatures were determined.
The influence of salt temperature on the steam plant power conversion efficiency was somewhat more involved, and required a number of simplifying assumptions to reduce the investigative effort. (At the time of the study, the plant configuration and cycle parameters had not been determined. The approach was to use the performance of an existing steam cycle as the refer-ence and adjust the cycle efficiency in proportion to the Carnot cycle efficiency based on the average temperature of heat addition. A constant temperature differential of 56 UC (100 F) was assumed between the hot salt and maximum steam temperature.
Table 1 summarizes the results of this preliminary cost tradeoff
study. The cost tradeoff was made by adjusting the individual effects to
mills/UW-hr. The preliminary cost estimates scaled from the estimated
costs are given in Ref. 17.
The results from Table 1 are illustrated graphically in Fig. 6. From this plot the minimum impact on the cost of electricity occurs at a salt temperature of approximately 543°C (1010°F). Of all the parameters con-sidered, the salt thermal decomposition rate is the most uncertain. The effect of increasing and decreasing the thermal decomposition rate by 50% was analyzed to determine its overall effect on the salt storage temperature. The results indicate only a marginal shift from 535°C (995°F) to 565°C (1050°F) as shown in Fig. 6. The resulting salt storage temperature was taken as 543°C (1010°F), on which to base the remainder of the plant design.
4.2 CYCLE PARAMETERS FOR THE REFERENCE DESIGN •
The cycle parameters were selected based on reasonable engineering
judgments consistent with conventional practice and the study groundrules.
16 1
NITROGEN EVOLUTION (CC/MOLE NaM02 /d)
727 593 500 °C
GA-A14160 Fig. 5. Rate of nitrogen evolution from HTS vs temperature
17 1
TABLE 1 PRELIMINARY COST TRADEOFF STUDY FOR EOT SALT STORAGE TEMPERATURE
Item or Parameter Case 1 Case 2 Case 3 Case 4
Hot salt temperature, °C (°F) 510 (950) 538 (1000) 566 (1050) 593 (1100)
Maximum steam temperature, °C (°F) 454 (850) 482 (900) 510 (950) 538 (1000)
Steam plant cycle efficiency 36.2 37.2 38.1 39.0
HTS inventory, kg x 10 6 (lbm x 10 ) 128 (282) 117 (258) 109 (240) 100 (220) HTS inventory cost, $ x 106 37.5 34.3 31.9 29.3
£
Tank cost, $ x 10 19.2 18.0 17.0 16.0 Annual salt replacement cost, $ x 10^ 3.7 4.9 8.4 12.9 A. Salt replacement cost in
mills/kW(e)-hr 1.6 2.1 3.9 6.5
B. Annual capital charge for HTS inventory and storage tanks (Case 4 as reference base)
0.8 0.5 0.2 0
C. Peaking power cost delta (assume base efficiency = 39% mills/kW(e)-hr
3.1 2.0 0.95 0
Cost of electricity tradeoff E (A + B + C) mills/kW(e)-hr
5.5 4.6 5.05 6.5
7.0 510°
6.0
DIFFERENTIAL CHANGE IN COST OF ELECTRICITY (MILLS/kW-HR)
5.0
4.0
3.0
2.0
540°C T
COST OF ELECTRICITY WITH 50% MORE THERMAL DECOMPOSITION
\ \ s v . .
COST OF ELECTRICITY WITH CALCULATED THERMAL DECOMPOSITION
COST OF ELECTRICITY — — WITH 50% LESS THERMAL DECOMPOSITION I 950 1000 1050 1100
HEAT TRANSFER SALT TEMPERATURE (°F) GA-A14160
Fig. 6. Tradeoff study of the heat transfer salt temperature vs differential change in the cost of electricity
19
The system design parameters can be divided into the following areas: primary helium, secondary or intermediate loop helium, base load steam, HTS loop, and peak load steam systems. The system parameters for the primary helium system were selected in order to remain consistent with the HTGR design conditions. The major system parameters were then calculated for the intermediate helium loop, base load steam, HTS loop, and peaking plant steam systems. These system parameters were then utilized in the initial component design studies. If unreasonable component designs resulted because of the imposed design conditions, the system parameters were adjusted accordingly. The number of such iterations had to be mini-mized due to the constraints of time; however, a few iterations were required to properly trade off the design conditions for the various heat exchangers.
The final selection of system parameters does not represent an optimi-
zed set of parameters. They do represent a selection based on sound engineer-
ing judgment consistent with the time and scope of this study. The parameters
selected for the various systems are discussed below.
4.2.1 Primary Helium System Parameters
The starting point in the selection of the system parameters was the
primary helium conditions. Consistency with the existing HTGR technology
was assumed as the basic ground rule in the selection procedure.
A reactor outlet temperature of 760°C (1400°F) was selected. Heat losses to the PCRV from the core cavity, to the core auxiliary heat exchangers (CAHE) and cavity, and to the intermediate heat exchangers (IHX) cavities are taken the same as in the reference 2000 MW(t) steam cycle HTGR (SC-HTGR). System pressure drops and helium compressor efficiencies similar to the 2000 MW(t) SC-HTGR were used in calculating compressor work and thermal heat load imposed on the IHX. Figure 7 illustrates the primary helium system and the selected cycle parameters.
20 1
HELIUM CIRCULATOR EFFICIENCY -•= 79.4%
352.7°C (666.8°F) 5 MPa (725 PSIA)
REACTOR CORE
2000 MW(t)
K>
HELIUM FLOW 943.7 Kg/SEC (7.481 X 10® LB/HR)
SHAFT WORK 44.4 MW(t)
343.6°C (650.5°F) 4.8 MPa (704.3PSIA)
3.5 MW(t) PCRV
3.4 MW(t) PCRV
Z0Z7.3 MW(t)
10.2 MW(t) CAHE
INTERMEDIATE HEAT EXCHANGER
760°C (1400°F) 4.9 MPa (713 PSIA)
757.8°C (1396°F) 4.9 MPa (712 PSIA) HEAT BALANCE
HEAT GENERATED BY THE CORE 2000 MW(t) HEAT ADDED BY THE CIRCULATORS 44.4 MW(t) HEAT LOST TO PCRV FROM
CORE CAVITY 3.5 MW(t) HEAT LOST TO CAHE AND CAVITY 10.2 MW(t) HEAT LOST TO PRCRV FROM IHX CAVITY 3.4 MW(t) NET HEAT TRANSFERRED TO IHX 2027.3 MW(t)
GA-A14160
Fig. 7. Cycle parameters for the primary helium system
4.2.2 Intermediate Loop Parameters
The interface between the primary and intermediate loop occurs in the IHX. A hot end temperature difference (between the inlet primary helium and outlet secondary helium) of approximately 55.5°C (100°F) and a cold end temperature difference (between the outlet primary helium and inlet secondary helium) of 27.7°C (50°F) resulted after iterating through the entire system calculation and trading off cycle performance with com-ponent design conditions. The thermal heat loss from the piping and heat exchanger cavities was approximated as 6% of the total thermal load. The secondary loop cycle conditions calculated are given in Fig. 8, which illustrates schematically the intermediate loop. The heat transferred from the intermediate loop is 2060 MW(t), of which 1210 MW(t) is utilized by the base load steam plant and the remaining 850 MW(t) is utilized by the HTS loop.
4.2.3 Base Load Steam Cycle Parameters
The base load steam parameters were modeled after the cycle conditions
used in the reference 2000 MW(t) SC-HTGR. Some modifications were necessary,
however, to account for the additional circulator power requirements.
Figure 9 illustrates the final cycle flow diagram and gives the appropriate
cycle parameters.
The gross electrical power output from the base load plant is 402.9 MW(e), with an estimated auxiliary power load of 7.6 MW(e), resulting in a base plant net efficiency of 32.7%. The base plant, however, supplies all of the circulator power for the primary system and intermediate loop. Proportioning the circulator power to both the base load and peaking plant would result in an adjusted base load plant efficiency of 37.4%, which is equivalent to the 2000 MW SC-HTGR plant efficient with an intermediate loop. Table 2 gives a breakdown of the base plant electric power output, auxiliary loads, and adjusted circulator power requirements.
22 1
317°C (602.7°F)
2027.3 MW(t) 1
HELIUM FLOW 1008 Kg/SEC (7.99 X 106 LB/HR)
315.8°C (600.5°F) 5 MPa (725 PSIA)
INTERMEDIATE LOOP HELIUM CIRCULATOR
. / I
6.4 MW(t) RECUPERATED FROM PIPING LOSS
3.2 MW(t) HEAT EXCHANGER CAVITY HEAT LOSS
9.1 MW(t) PIPING HEAT LOSS
SHAFT WORK 1 45 MW(t)
307.4°C (585.4°F) 4.8 MPa (704 PSIA)
6.4 MW(t) HOT SALT PIPING HEAT LOSS
- 4 704.4°C (1300°F) 4.9 MPa (712 PSIA)
702.8°C (1297°F 4.9 MPa (711 PSIA)
856.4 MW(t) I * SALT LOOP"
2066.4 MW(t) —I
I — * 1210 MW(t) BASE LOAD FLANT
/ 850 MW(t) THERMAL STORAGE
HEAT BALANCE HEAT TRANSFER FROM THE IHX 2027.3 MW(t) HEAT ADDED BY THE CIRCULATORS 45.0 MW(t) HEAT LOSS FROM PIPING 2.7 MW(t) HEAT LOSS FROM HEAT EXCHANGER
CAVITIES 3.2 MW(t) NET HEAT TRANSFER 2066.4 MW(t)
GA-A14160
Fig. 8. Cycle parameters for the intermediate helium system
1518.1 H
HELIUM
1297 F 712P
SH EVAP ECON
2513P 955 F
1210 MW(t)
2965P 369'F
4.67 X 106 W
585 F 7MP
HELIUM
130S.4H 643P 634 F
3.20 X 106 W
U27.3H
STM GEN THERMAL POWER HELI'JM CIRC. POWER NSS THERMAL POWER GROSS PLANT ELEC OUTPUT CIRC. WATER PUMP POWER OTHER STATION AUX NET PUNT ELECTRICAL OUTPUT
1210 MWIt) 89.4 MW|t)
1120.6 MW(t| 402.9 MW|e)
6.3 MWIel 1.3 MW[e)
395.3 MWte)
586P 1002 F
REHEAT
PRIMARY HE CIRC TURBINE DRIVE
C 44.4MW
1 INT. LOOP HE CIRC
LEGEND H ENTHALPY (BTU/LBI W LB/HR P PSIA
346.4H
BFP
GA-A14160
Fig. 9. Base load plant (English engineering units) (Sheet 1 of 2)
3531.1H STM GEN THERMAL POWER HELIUM CIRC. POWER NSS THERMAL POWER GROSS PLANT ELEC OU"?UT CIRC. WATER PUMP POWER OTHER STATION AUX NET PLANT ELECTRICAL OUTPUT
1210 MWU) 89.4 MW|!)
1120.6 MW(t| 402.9 MW|e)
6.3 MW(e) 1.3 M W I e l
395.3 MW|e)
to cn
H ENTHALPY KJ/Kg W Kg/SEC P MPa
GA-A14160
Fig. 9. Base load plant (SI units) (Sheet 2 of 2)
TABLE 2 ADJUSTED AUXILIARY POWER REQUIREMENTS FOR
THE BASE LOAD AMD PEAKING PLANTS
Base Load Peak Plant
Item Actual Adjusted Actual AdJ usted
Steam Generator Thermal Power, MW(t) 1210 1210 2550 2550
Helium Circulator Power, MW(t) Primary Loop 44.4 25.5 0 56.7 Intermediate Loop 45 25.9 0 57.3
Nuclear Steam System Power, MW(t) 1120.6 1158.6 2550 2436
Gross Electrical Power Output, MW(e) 402.9 440.9 1026.7 1026.7-114
Plant Auxiliary Power, MW(e) 7.6 7.6 17.2 17.2
Net Plant Power, MW(e) 395.3 .433.3 1009.5 895.5
Plant Efficiency 35.3 37.4 39.6 35.1
GA-A14160
26 1
4.2.4 HTS Loop and Peaking Plant Cycle Parameters
One of the most critical areas in the selection of the system para-meters occurs in the interface between the salt loop and the peaking steam plant. The peaking plant requires a steam reheat temperature of approxi-mately 510°C (950°F) to maintain a high cycle efficiency. The maximum salt temperature, however, is only 543°C (1010°F). High thermal perform-ance of the salt/steam heat exchanger therefore is required. Figure 10 illustrates the temperature profiles that occur in the salt/steam heat exchanger. Severe temperature pinching problems can occur in the heat exchanger if the cycle conditions are not selected properly. The pinch temperature can be relieved in two ways; first, the cold salt temperature can be increased, and second, the pressure of the steam cycle can be decreased. The first method results in a tradeoff of the design conditions of the helium/salt heat exchanger. As the cold salt temperature is raised (assuming a fixed hot salt temperature) the log mean temperature difference of the helium/salt heat exchanger decreases, resulting in a larger heat exchanger surface area requirement. The design temperature of the IHX can also be adjusted, but this imposes changes to the primary helium conditions and deviations from the standard HTGR technology.
The second method involves reducing the steam cycle operation pressure to relieve the temperature pinch. This method has a direct impact on the steam cycle efficiency.
A combination of the above two methods was used to select cycle con-
ditions that resulted in adequate cycle efficiency and an adequate pinch
point margin.
The resulting selection is based on a series flow configuration for
the salt/steam heat exchanger. The water flow enters the steam generator
section (consisting of the economizer, evaporator, and superheater) then
flows to the high-pressure turbine and back to the reheat section. The
salt flows in series through the reheater section and the steam generator
27 1
SECTION ECONOMIZER EVAPORATOR SUPERHEATER REHEATER
HEAT TRANSFER COEFFICIENTS W/m2-°C (BTU/HR-FT2 °F)
TUBE SIDE SHEI.L-SIDE 19300 (3400) 57760 (10000) 6470 (1140) 6470 (1140)
SECTIONS
2840 (500) 2840 (500) 2840 (500) 2840 (500)
1010°F
950°F
X
\VAPOR DOME \
\
N \
I I I I I i I I I L 0 0.2 0.4 0.6 0.8 1.0 1.2 1.4 1.6 1.8 2.0 2.2 2.4
ENTROPY (BTU/LB—1F) 1 I I I I 1
2000 4000 6000
J/Kg— °C
8000
GA-A14160
Fig. 10. Temperature profiles for the salt/steam generator and reheater
28 1
section. The salt cools from approximately 543°C (1010°F) to 501°C (933°F) in the reheater section and from 501°C (933°F) to 288 0C (550°F) in the steam generator section. Additional temperature and heat-transfer coefficient information for the salt/steam generator and reheater are supplied in Fig. 10.
The high-pressure turbine inlet conditions of 482°C (900°F) and 13.8 MPa (2000 psia) and the intermediate pressure turbine conditions of 510°C (950°F) and 3.8 MPa (550 psia) were selected to conform to conventional peaking turbine practice.
Steam cycle calculations were conducted to determine the system para-
meters for the entire loop. Figure 11 illustrates the peaking plant's
cycle diagram.
The peaking plant gross electrical output is 1026.7 MW(e), and 17.2 MW(e) auxiliary power is required for the station auxiliaries, which results in a net plant output of 1009.5 MW(e). Adjusting the plant output to account for the peaking plant's portion of circulator power (supplied by the base load plant) results in an adjusted plant efficiency of 35.1%. Table 2 gives a breakdown of the peaking plant electric power output, auxiliary loads, and adjusted circulator power requirements.
29 1
s
1491.2 H
OJ O
STM GENERATOR THERMAL POWER 2550 MW|!) GROSS PLANT ELECTRICAL OUTPUT 1026.7 MW|e) CIRC. WATER PUMP POWER OTHER STATION AUX.
143 MW(e) 2.9 MW(e)
6.4 M W W PIPING ANO TANK HEAT LOSS
NET PLANT ELECTRICAL OUTPUT 1009.5 MWIel
585 f 7D4P
HEUUM
LEGEND
H ENTHALPY (BTU/LB) W IB/HR P PSIA V VOLUME
0.4 MW(e!
GA-A14160
Fig. 11. Peaking plant cycle diagram (English engineering units) (Sheet 1 of 2)
346a SH STM GENERATOR THEBMA1 POWER 2550 MW|t) GROSS PLANT ELECTRICAL OUTPUT 1026.7 MW[e) CIRC. WATER PUMP POWER 14.3 MWIe) OTHER STATION AUX. 2.9 MWIe) NET PLANT ELECTRICAL OUTPUT 10095 MWIe)
LEGEND
H ENTHALPV KJ/Kg W Kg/SEC P MPa V VOLUME M 1
0.-. MWIe)
GA-A14160
Fig. 11. Peaking plant cycle diagram (SI units) (Sheet 2 of 2)
5. REFERENCE PLANT DESIGN
5.1 SYSTEM CONFIGURATION
The proposed system configuration is illustrated in Fig. 12. There are, however, a number of variations on the system diagram that can be considered. The proposed system configuration was selected because it best satisfied the system requirements and remained consistent with the established groundrules.
The basic system consists of the reactor system, the intermediate heat transfer helium system, the base load steam system, salt storage system, and the peaking plant steam system.
The primary helium coolant flows down through the reactor core,
transferring heat from the graphite reactor core to the primary helium
coolant. The primary helium transfers the absorbed heat to the inter-
mediate heat-transfer loop through four equally sized helium/helium heat
exchangers within the PCRV. The intermediate loop is comprised of the
helium/steam generators, the helium/salt heat exchangers, and intermediate
loop helium circulators. The thermal energy transferred by the inter-
mediate loop is delivered in parallel to the base load steam plant and the
HTS storage loop. The thermal energy delivered to the helium/steam gen-
erator is used directly to supply superheated steam for continuous oper-
ation of a conventional turbine generator to produce base load power.
The remaining thermal output continuously heats a conventional HTS. The
hot salt is stored in large reservoirs. During peak electrical demand
periods, the hot salt is circulated from the hot reservoirs through a salt/
steam heat exchanger to generate superheated steam for use by the turbine/
generator conversion equipment of the peaking plant. The cold salt is
32 1
BASE 10AD PLANT
STURAGE LOOP
PEAKING PLANT
GA-A14160
Fig. 12. HTGR/HTS thermal peaking plant
returned to the cold tanks for completion of its flow path. A more detailed system description of the various loops follows.
5.2 REACTOR SYSTEM
The reactor system is basically the GA 2000 MW(t) HTGR, as shown in Fig. 13, with the steam generators replaced by four helium-to-helium IHXs spaced symmetrically around the central cavity of the PCRV. The primary helium flows down through the graphite reactor core, picking up thermal energy and transferring it to four equally sized IHXs. The primary helium system is comprised of two independent loops each consisting of two IHXs and one helium circulator. The primary helium flow from the pair of IHXs in each loop supplies cool helium to the circulator. The helium circulators discharge the helium coolant into the core inlet plenum. The two CACSs, between the main cooling loops, are identical to what is used on the standard 2000 MW(t) SC-HTGR. The auxiliary cooling loops, each consisting of a circulator, a helium shutoff valve, and a helium/water heat exchanger, cool the core after a reactor shutdown when the main loops are out of service.
The thermal reactor employs U-235 as the fissile material, Th-232 as
the fertile material and graphite as the moderator, cladding, and reflector.
Each fuel element is made up of two basic parts: the graphite hexagonal
prism structure and the bonded rods of coated fuel particles. The entire
primary helium circuit is contained by the PCRV, which both contains the
coolant operating pressure and provides radiological shielding.
5.3 INTERMEDIATE HELIUM SYSTEM
The intermediate helium system transports the reactor heat from the
IHXs to the helium/steam generator and the helium/salt heat exchanger.
This intermediate loop eliminates the consequences of the salt reaction
with the graphite reactor core. The intermediate helium system consists
of intermediate helium circulators, piping, and isolation valves. Two
parallel loops are utilized as illustrated in Fig. 14.
34
P E N E T R A T I O N COVERS CONTROL ROD STORAGE W E L L S .
HOLD DOWN P L A T E S
CONTROL ROD AND D R I V E I N S T R U M E N T A T I O N .
P R E S T R E S S E D CONCRETE REACTOR V £ S S E L .
PCRV P R E S S U R E R E L I E F S V S T E H
PCRV SUPPORT STRUCTURE
L I N E A R ' P R E S T R E S S I N O S Y S T E M
GA-A14160
Fig. 13. Major components and systems of the HTGR NSS
35 1
OPERATING CONDITIONS
NO. FUNCTION tGAS
°C <°F)
PGAS MPa (PSIA)
© IHX OUTLET 704 (1300) 5.0 (725)
© CONCENTRIC IHX OUTLET 704 (1300) 5.0 (725)
© HTS/He HX INLET 704 (13D0) 5.0 (725)
© STM/He HX INLET 704 (1300) 5.0 (725)
© CIRCULATOR
"LINE (IHX INLET)
307 (585) 5.0 (725)
© IXH INLET CONCENTRIC 307 (585) 5.0 (725)
© HTS/He OUTLET 307 (585) 5.0 (725)
© STM/He OUTLET 307 (585) 5.0 (725)
HTS'HB
0 W 4 STM He © - A A A r - K D
CONTAINMENT ISOLATION VALVES
INTERMEDIATE HELIUM CIRCULATOR AND DRIVE
0 )
© r A A A r n ® STM/Hs
CONTAINMENT g H W W ®
HTS/He GA-A14160
Fig. 14. Flow diagram for the primary and intermediate loop helium systems
Most of the piping is in the reactor containment building (RCB) annulus.
A concentric configuration for this piping was selected for the following
reasons:
1. The IHX interface piping geometry is concentric.
2. Seventy percent of the heat loss is through the hot-leg pipe in the containment, and, rather than lose this heat to the contain-ment air conditioner system, it is recuperated to cold-leg coolant.
3. In-containment safety is enhanced, because the cold-leg pipe forms a guard pipe for the hot-leg pipe.
The concentric piping runs from the IHXs to the containment boundary. The concentric piping then separates into single hot- and cold—leg piping before penetrating the containment. Two pipe penetrations (hot and cold) per loop are required. A helium isolation valve is on each line just out-side the containment boundary. Missile protection is provided around each of the "Y" globe pattern helium isolation valves.
External piping runs are minimized by locating the helium/salt and helium/steam heat exchangers close to the RCB. The hot piping requires guard grating for personnel safety, as its surface temperature is 232°C (450°F). Thermal insulation, 2.5 cm (1 in.) thick and jacketed, will be applied to the outside of all cold-leg piping to keep the outer surface temperature at about 60°C (140°F) for personnel protection. Hot-leg piping will have 10 cm (4 in.) of insulation on its inner surface, covered by 1.25 cm (1/2 in.) thick HTGR-type cover plate, except that only 7.5 cm (3 in.) of insulation is needed inside the concentric hot leg pipes. The external hot-leg piping surface temperature is 232°C (450°F) and will require guard grating.
The total heat loss to the environment from the piping is 2.75 MW(t). A total of 9.1 MW(t) is transferred through pipe walls, but 6.35 MW(t) is recuperated.
37 1
The cold intermediate helium in each loop is routed to the helium circulator, which maintains the loop flow. The circulators are described in Section 6.2
Valves adjust flow distribution and isolate individual heat exchangers.
Maintenance operations, however, can be delayed until a reactor shutdown is deemed necessary for any reason other than a secondary helium system malfunction. Then, with the reactor depressurized, the IHXs intact, and the containment isolation valves locked closed, the piping can be cut and capped for component or piping repairs.
Table 3 gives the intermediate helium loop piping summary. The
piping material selected was SA-155, Grade KC-70, and wall thicknesses
were based on Section H I , Class 3 of the ASME Code design rules using a
316°C (600°F) wall temperature and the appropriate pressure differential.
5.4 BASE LOAD STEAM SYSTEM
The base load steam system is modeled directly from the 2000 MW(t)
SC-HTGR, using a conventional reheat cycle. The steam generators produce
superheated steam at 513°C (950°F) and 16.6 MPa (2400 psia) and reheat
steam at 537.8°C (1000°F) and 3.8 MPa (550 psia).
Steam from the high-pressure turbine exhaust passes through the primary helium circulator turbines before being reheated for admission to the intermediate-pressure turbine. Approximately 19% of the intermediate-pressure turbine exhaust is routed to the condensing drive turbines for the boiler feedpumps and intermediate helium circulators. The exhausts from the condensing turbines flow directly to the surface condenser. The remaining intermediate-pressure turbine exhaust is admitted to the low-pressure turbine. Turbine bleed flows are extracted from the low-pressure turbine for three stages of closed feed water heating and one.stage of deaeration.
38 1
TABLE 3 INTERMEDIATE LOOP Re PIPING SUMMARY
c
Number ( a ) Function
T gas
°C (°F)
F
MPa gas (psia)
Internal Flow Diam
cm (in.)
Pipe Inside Diam
cm (in.)
i
Pipe Outside
Diam cm (in.)
t External
Pipe Insulation Thickness cm (in.)
Internal Thermal Barrier
Thickness cm (in.)
1 IHX Outlet 704 (1300) 5 (725) 183 (72) 206 (81) 217 (85.5) — 11.4 (4.5)
lc Concentric IHX Outlet
704 (1300) 5 (725) 132 (52) 150 (59) 158 (62.3) — 8.9 (3.5)
2 HTS/He HX Inlet
704 (1300) 5 (725) 114 (45) 137 (54) 145 (57.0) — 11.4 (4.5)
3 Steam/He HX Inlet
704 (130C) 5 (725) 137 (54) 160 (63) 167 (66.5) — 11.4 (4.5)
4 Circulator Line (IHX Inlet
307 (585) 5 (725) 152 (60) 152 (60) 161 (63.3) 2.5 (1.0)
4c IXH Inlet Concentric
307 (585) 5 (725) 188 (74) 188 (74) 198 (78.0) 2.5 (1.0) —
5 HTS/He Outlet
307 (585) 5 (725) 97 (38) 97 (38) 102 (40.1) 2.5 (1.0) —
6 Steam/He Outlet
307 (585) 5 (725) 114 (45) 114 (45) 121 (47.5) 2.5 (1.0) —
GA-A14160 ^ T h e numbers correspond to sections of piping illustrated on Fig 14.
The low-pressure turbine exhaust is condensed by a surface condenser
operating at 57 ran (2.25 in.) of mercury. The heat load of 715 MW(t)
imposed on the condenser is rejected through a natural draft evaporative
cooling tower. The design parameters were selected to be consistent with
the 2000 MW(t) steam cycle design. The circulating water system design
is based on 24° C (75°F) dry bulb temperature, 50% relative humidity,
11°C (20°F) approach temperature, and 13.3°C (24°F) rise in the circulating
water temperature across the condenser.
The total design circulating water flow is 12,800 kg/s (102 x 10 lbm/hr). Three pumps provide operating flexibility and remain within capacity limitations of process pump designs. The large diameter piping between the condenser and cooling tower is concrete, and the smaller piping is carbon steel or fiberglass-reinforced plastic. Makeup water to replace evaporation losses is untreated lake water supplied by the raw water pumps. A blowdown line limits solids buildup.
The longitudinal-tube, three-pressure, divided-waterbox condenser
contains stainless steel tubes and a deaerating hot well. Conductivity
instruments monitor tube sheets and detect leaks and their location.
The condensed steam is transferred from the condenser hot well to the
full-flow deep-bed condensate demineralizer system by two 50% capacity
motor-driven condensate pumps. The condensate demineralization and chemical
treatment consist of hydrazine and ammonia injection.
The conventional, vertical, multistage electric-motor-driven condensate
pumps take suction through individual suction lines from the common con-
denser hot well. Condensate flow is controlled by deaerator level, with
anticipatory signals from the feedwater demand signals.
The hot well level control system consists of a level controller and three split-range pneumatic level control valves. Upon detection of a
40 1
high hot well level, the condensate reject valve is opened, permitting con-densate to be pumped back to the storage tank. Upon detection of a low hotwell level, the controller operates a normal makeup valve, which sprays makeup condensate from the storage tank into a deaerating spray header in the condenser neck. Upon detection of a very low hotwell level, an emer-gency makeup valve admits condensate directly to the not well at a rate dictated by certain unit transients.
Condensate pumps are tripped by the usual electrical protective devices (undervoltage, fault, etc.), very low hotwell level, or very high deaerator storage tank level. When the unit is operating on one condensate pump, the idle condensate pump automatically starts on trip of the operating pump. If the unit is operating on two pumps and one pump trips, the unit begins an emergency runback to 50% power. The deaerator storage tank can hold adequate feedwater for necessary power ramps.
The condensate then flows through the feedwater heaters, which are sized based on the requirements of the turbine thermal cycle. All the heaters have stainless steel tubes to minimize iron pickup in the feedwater system. The heaters also have integral drain coolers, except the lowest pressure units, xjhich have an external drain cooler. External drain coolers were selected for these units to permit recovery of cycle energy during periods of prestart :ycle cleanup, thereby reducing the total auxiliary steam heating requ. rement.
The feedwater system pumps the deaerated condensate from the deaerator
to the steam generator. The feedwater system consists of a deaerator storage
tank, two turbine-driven feedwater booster pumps, and two turbine-driven
feedwater pumps, necessary valves, piping, instrumentation, and controls.
Feedwater pumps and feedwater booster pumps are driven by common steam
turbines. Booster pumps and feedwater pumps are equipped with flowmetering
and minimum-flow recirculation control to ensure minimum flow protection
under all conditions of unit operation, including pump startup and shutdown.
Idle pump warming is accomplished by recirculation of feedwater through
41 1
these same recirculation connections. The feedwater booster pumps are
rated at 213 m (700 ft) total dynamic head and 5.3 rad/s (2000 rpm).
The feedwater pumps are rated at 2440 m (8000 ft) total dynamic head and
14.6 rad/s (5500 rpm).
The turbines driving the booster-pump/feed-pump combination will trip on receipt of various turbine protective signals, such as low bearing oil pressure, low exhaust vacuum, thrust bearing failure, or high vibration. They are also tripped upon receipt of a very low deaerator storage tank level signal. Turbine speed is controlled by feedwater demand signals. The supply to each steam generator contains a flow measuring element and a remote manual trip valve, which is used to balance feedwater flow to the steam generators.
The flow diagram for the base load system is illustrated in Fig, 9.
A summary of the base load steam components is given in Table 4.
5.5 THERMAL ENERGY STORAGE SYSTEM
The heat storage system uses groups of four hot and four cold tanks
containing HTS at 543°C (1010°F) and 288°C (550°C). The top surface of
the fluid is covered with nitrogen at slightly above atmospheric pressure.
The fluid is pumped continuously from the cold tank, through twin heat exchangers (where it is heated from 288°C (550°F) to 543°C (1010°F),
c to the hot tank at the rate of 2134 kg/s (16.9 x 10 lb/hr). During
the 8-hr period of each day when peaking power is required, fluid is
pumped from the hot tank through twin steam generators (where it is
cooled to 288°C [550°F]) to the cold tank at the rate of 6400 kg/s (50.7
x 10 lb/hr). The capacities of the hot and cold tanks to accommodate
these flows are 73600 and 65100 m 3 (2.6 x 10 6 and 2.3 x 10 6 ft 3).
The heat storage system and its relationship to the peaking system
is shown diagrammatically in Fig. 12.
42 1
TABLE 4 BASE LOAD STEAM SYSTEM COMPONENTS
Turbine-Generator
Type
Throttle Plow
Throttle Pressure
Throttle Temperature
Reheat Steam Pressure (at intercept valve)
Reheat Steam Temperature (at intercept valve)
Turbine Back Pressure
Generator rating at 0.9 Power Factor
Condensate Pumps
Number
Flow (total)
Approximate TDH
Condensate Storage Tanks Number
Capacity
HPT/IPT Combined Unit LPT - Tandem Compound - 2 flow -
12.26 m 2 (132 ft 2) exhaust annulus area, 89 cm (35 in.) Last stage bucket.
403.7 kg/sec (3.2 x 10 6 lbm/hr)
16.64 MPa (2414 psia)
510°C (950°F)
3.8 MPa (550 psia)
537.8°C (1000°F)
57.15 mm Hg (2.25 in. Hg)
432 MVA
Minimum of 2
382 kg/sec (3.03 x 10 6 lbm/hr)
2.3 MPa (330 psi)
1
473 m 3 (125,000 gal.)
GA-A14160
43 1
TABLE 4 (Continued)
Feedwater Heaters Heater ill Heater HZ Heater II3 Deaerating Heater #4
Number 2 1 1 1
Feedwater Flow, kg/sec 191.1 382.2 382.2 403.7
(lbm/hr) each (1.52xl06) (3.03xl06) (3.03X106) (3.2xl06)
Steam Flow, kg/sec 11.8 21.2 22.5 21.4
(lbm/hr) each (93,500) (168,000) (178,000) (170,000) 1
Feedwater Inlet Temp., °C (°F)
41.1(106) 82.2(180) 116.1(241) 150.6(303)
Feedwater Outlet Temp., °C (°F)
82.2(180) 116.1(241) 150.6(303) 183.3(362)
Shell Pressure KPa (psia) 57.2(8.3) 191.7(27.8) 518.5(75.2) 1080(156.7)
Terminal Temp. Difference, °C (°F)
2.8(5) 2.8(5) 2.8(5) 2.8(0)
Drain Cooler Approach, °C (°F)
1 1 5.5(10) 5.5(10)
1
5.5(10) —
Boiler Feedpumps 1 !
Number 2
Drive i ! Steam Turbine |
Flow (eacii) kg/sec (lbm/hr )
} 202 (1.6xl06)
Approximate TDH MPa (psi) j ; 20.9 (3039)
Suction Temperature, °C (°F)
1 183.3(362)
i Main Condenser i i Turbine Exhaust Pressure,
mm (in.) Hg ; 57.2(2.25) i i
Condenser Heat Duty, MW(t)
j 715 1
Cooling Water Rise, °C (°F)
13.3 (24)
44
Tie HTS oxidizes arid promotes general corrosion of the ferrous alloys. Studies have shown, however, that corrosion rates are uniform and decrease with time (Refs. 19 and 21 through 24). These results substantiate the for-mation of a tenacious, passive oxide film, which is desirable. Up to tempera-tures of 454°C (850°F), carbon steels have been successfully used with virtu-ally no corrosion. For higher temperature systems operating near 593°C (1100°F), stainless steels or heat-resistant alloys are recommended by the supplier. They have shown imporved resistance to general corrosion over carbon steels and possess the strength required at the elevated temperatures.
With a large system involving heat transport, the potential for mass transport of chemical species within the HTS is of conccrn. If elements are removed from the structural containment materials and redeposited or absorbed at different locations, design integrity could be impaired. This phenomenon would not be anticipated in an HTS system. The tenacious, passive oxide present on the metal surfaces should act as a very effective barrier to diffusion of interstitial elements such as carbon.
Data indicate that temperatures up to 593°C (1100°C) can be handled
satisfactorily from a corrosion standpoint, the salt decomposition becoming
the limiting factor at this level. The low-temperature side of the loop
thus can utilize carbon steel, whereas the high-temperature side will
require stainless steel or other alloy.
The 142°C (288°F) freezing point of the pure salt presents some engineering difficulties in ensuring that either the salt does not freeze or if it does to provide for remelting. For the refinery study, a steam tracing system was selected for heating the salt during startup and shutdown. For the application that uses large storage tanks, it may be possible to design the system such that the loop can be drained to the low-temperature reservoir where heating of that unit only can be provided. The melting point of the salt can be lowered by adding water as the temperature is dropped. This method does have advantages for the initial shakedown and startup operations, and is discussed in the plant operation section.
45 1
5.6 PEAKING LOAD STEAM SYSTEM
The peaking steam cycle design is similar to the base load steam
cycle except for the following areas:
1. High-pressure turbine inlet conditions are 482°C/13.8 MPa
(900°F/2000 psia).
2. Intermediate-pressure turbine conditions are 510°C/3.8 MPa
(950°F/550 psia).
3. Helium circulator drive turbines are eliminated.
The lower temperature and pressure inlet conditions of the high-pressure turbine are advantageous since the turbine is operated in a cycling mode. This results in a lower turbine casing design conditions and reduces the physical mass of the machine. Conventional practice also indicates a lower turbine inlet temperature and pressure is desirable for turbines used in peaking applications.
The peaking cycle also includes reheat of the high-pressure turbine exhaust. The remainder of the peaking cycle is similar to the base load plant description.
The circulating water system is based on the same design conditions,
but has a heat duty of 1470 MW(t). A separate natural draft evaporative
cooling tower is utilized during operation of the peaking plant. The
total operating circulating water flow is 26360 kg/s (209 x 10^ lbm/hr)
and requires five pumps to provide operating flexibility and remain within
proven pump capacities.
Figure 11 illustrates the flow diagram for the peaking steam plant
and gives the plant operating conditions. Table 5 summarizes the major
component design parameters of the peaking plant.
46 1
TABLE 5 PEAKING LOAD STEAM SYSTEM COMPONENTS
Turbine-Generator Type HPT - 2 flow
IPT - 2 flow LPT - Tandem Compound - 6 flow -
36.78 m2 (396 ft 2) exhaust annulus area, 89 cm (35 in.) Last stage bucket.
Throttle Flow 863 kg/sec (6.84 x 106 lbm/hr)
Throttle Pressure 13.8 MPa (2000 psia)
Throttle Temperature 482°C (900°F)
Reheat Stefm Pressure (at intercept valve)
3.8 MPa (550 psia)
Reheat Steam Temperature (at intercept valve)
510°C (950°F)
Turbine Back Pressure 57.15 mm Hg (2.25 in. Hg)
Generator Rating at 0.9 Power Factor
1300 MVA
Condensate Pumps
Number Minimum of 3
Flow (total) 813 kg/sec (6.45 x 10 6 lbm/hr)
Approximate TDH 2.3 MPa (330 psi)
Condensate Storage Tanks
Number 2
Capacity 1 946 m 3 (250,000 gal.)
GA-A1416D
47
TABLE 5 (Continued)
Feedwater Heaters
Number
Feedwater Flow, kg/sec
(lbm/hr) each
Steam Flow, kg/sec
(lbm/hr) each
Feedwater Inlet Temp., °C (°F)
Feedwater Outlet Temp., °C (°F)
Shell Pressure, kPa (psia)
Terminal Temp. Difference, °C (°F)
Drain Cooler Approach, °C (°F)
Boiler Feedpumps
Nur'ier
Drive
Flow (each), kg/sec (lbm/hr)
Approximate TDH MPA (psi)
Suction Temperature, °C (°F)
Main Condenser
Turbine Exhaust Pressure, mm (in.) Hg
Condenser Heat Duty, MW(t)
Cooling Water Rise, °C (°F)
Heater #1 Heater #2 Heater #3
3
271.5 406.6
(2.15xl06) (3.22xl06)
20.2 26.0
(160,000) (206,000) 41.1(106) 82.2(180)
2
406.6
(3.22xl06)
24.6
(195,000)
116.1(241)
Deaerating Heater #4
2
430.6
(3.41xl06)
24.2
(192,000)
150.6(303)
82.2(180) 116.1(241) 150.6(303) 183.3(362)
57.2(8.3) 191.7(27.8) 518.5(75.2) 1080(156.7)
2.8(5) 2.8(5) 2.8(5) 2.8(0)
5.5(10) 5.5(10) 5.5(10)
Steam Turbine
430.6(3.4x10^)
15.1 (2194)
183.3(362)
57.2(2.25)
1470
13.3(24)
48
6. REFERENCE COMPONENT DESCRIPTIONS
Design studies were carried out on the major components comprising the HTGR/HTS peaking system. Components of the steam power-conversion loops that do not change significantly from other conventional power-conversion loops (i.e., steam turbines, generators, condensers, etc.) are not dis-cussed. Components that are unique to this concept or that represent significant design changes from conventional practice are discussed.
6.1 PCRV AND RCB
6.1.1 PCRV Structure
The selected PCRV configuration has two pairs of IHXs; each pair is connected to one primary helium circulator which circulates the primary helium coolant from the core outlet plenum through the two IHXs and into the core inlet plenum. Two auxiliary primary cooling loops are located between the main cooling loops in the PCRV. Secondary helium will be cir-culated through each of the four IHXs by secondary circulators located out-side the RCB. The primary functional requirement of the PCRV is to contain the reactor core, the primary coolant system, including the primary helium circulators, and the required heat exchangers. The PCRV serves as the primary coolant system pressure boundary together with the cavity liners and closures. The PCRV also provides the necessary biological shielding around the core. The general performance requirements for the PCRV are as follows:
1. The PCRV shall have an essentially elastic response to short-term
pressure changes up to the maximum cavity pressure.
49
2. The PCRV together with the liners and closures shall ensure, a leak-tight boundary for the primary coolant under normal, upset,
and emergency operating conditions throughout the life of the
PCRV.
The main safety requirements are that 1) the structural integrity shall be maintained throughout the PCRV design life, including exposure to seismic events, when the PCRV is at the maximum cavity pressure, and 2) a specified ultimate load capacity shall be available. The design shall comply with the ASME Code, Section III, Division 2.
The PCRV configuration is shown in Fig. 15. The PCRV is a multi-cavity, thick-walled cylindrical structure that contains the central cylindri-cal core cavity, which is surrounded by heat exchanger and circulator cavities. The top head of the PCRV also contains numerous other pits, wells, and penetrations, for refueling and for helium purification and other equipment housing. There are also primary pressure boundary penetrations for tem-perature and pressure measurements, neutron detection, and spare instruments.
The following design data were used for the PCRV design:
1. Normal working pressure (NWP) = 4.895 MPa (710 psig). 2. Maximum cavity pressure (MCP) = 5.350 MPa (776 psig).
3. Prestress safety factor (f) = 1.1.
Other design data are given in Table 6.
The diameter of the PCRV is generally determined by the largest side cavity. However, if a large number of side cavities is required, the diameter may be determined by geometric conditions owing to the number of cavities and the required ligament sizes between the cavities. For the minimum ligament between the cavities, a dimension of 1829 mm (6 ft) is used based on past experience.
50
en
GA-A14160
Fig. 15. Layout drawing of PCRV configuration
TABLE 6 PCRV DESIGN DATA
Item Dimensions m (ft)
Core cavity diameter, D^ 9. 957 (32.7)
Core cavity height, h 14. 417 (47.3)
Circulator cavity diameter 2. 438 (8.0)
Intermediate HX cavity diameter, D g 4. 250 (13.9)
Auxiliary cooler cavity diameter at core 2. 134 (7.0) cavity elevation
Precast panel thickness, t 0. 457 (1.5)
SUMMARY OF DIMENSIONS
PCRV
O.D. 28 .95 (95.0)
Height 25 .00 (82.0)
RCB
I.D. 38 .71 (127.0) (a) Height to spring line 50 .52 (165.75)
fa) Height to dome centerline 60 .96 (200.0)
(a) External height from top of foundation mat.
GA-A14160
52
A PCRV tlock diameter of 28.05 m (92.0 ft) was calculated based on only the average allowable compressive stress in the concrete. Block diameter is the PCRV diameter without the portion of the precast wire winding panels which contain the circumferential prestressing (CPS) wire and are 457 mm (18 in.) deep. Because there are only four major side cavities, the PCRV design is stress limited, rather than geometry limited, so that the overall PCRV diameter is 28.95 m (95 ft).
The height of the PCRV is determined by the height of the core cavity plus the thickness of the top and bottom heads.
H = h + h, + h c b t
= 14.5 + 4.9 + 5.6 m (47.5 + 16.0 + 18.5 ft)
= 25.0 m (82.0 ft)
where h t = the thickness of the top head of the 2000 MW HTGR, which has the same core cavity diameter and the same maximum cavity pressure.
6.1.2 RCB Structure
The RCB is a safety structure which provides a leakage barrier against significant fission product release to the atmosphere if a primary coolant boundary failure and subsequent depressurization accident occur. The reactor containment building also allows the density of the coolant remaining in the PCRV to be maintained as required for adequate core cooling following a design basis depressurization accident. The concrete structure provides shielding from fission products which could be airborne within the building following such an accident.
The maximum pressure during a primary system depressurization is
expected to be on the order of 0.345 MPa (50 psia) in the free containment
volume, and therefore the RCB would be designed for a maximum internal
53
gage pressure of about 0.276 MPa (40 psig), including a safety factor of
1.1. The building will be designed against short-term tornado loadings,
including tornado-generated missiles.
The RCB is sized so that following the design basis accident, contain-ment pressure is maintained at a sufficiently high pressure for approximately 1 day to permit proper performance of the core auxiliary cooling loops.
2 The refueling floor around the PCRV is designed for 14.36 kPa (300 lbf/ft ).
Adequate pipe restraints will safeguard against whipping due to pipe
ruptures. The building will be laid out for a minimum 1.37 m (4 ft 6 in.)
annular space around the PCRV for future wire winding and maintenance
requirements.
The RCB shown in Fig. 16 is a cylindrical, reinforced, prestressed concrete structure bearing on solid soil. The inside of the shallow dome is designed as a 2:1 ellipse. The foundation mat is reinforced with con-ventional steel reinforcing. The mat has a total thickness of 5.180 m (17 ft). The cylindrical walls have a thickness of 915 mm (3.0 ft) for adequate seismic resistance and are prestressed with a posttensioning tendon system in the vertical and circumferential directions. The dome has a thickness of 0.762 m (2.5 ft) for adequate external missile resistance and is prestressed. The lower anchorages of the vertical tendons are in a tendon gallery below the base mate. The top vertical tendon anchor-age and the dome tendon anchorage are in the ring girder area.
It is possible that the RCB design shown in Fig. 16 is not optimum
for the subject plant. A nearly equal-cost alternate design would feature
a larger free volume because of the use of a hemispherical dome rather
than the elliptical one shown. This design could be chosen if the resulting
postaccident containment pressure were low enough to allow elimination of
poststressing tendons from the dome and wall of the building, along with
the ring girder at the springline and tendon galleries beneath the base
mat.
54
I i
I
GA-A14160
Fig. 16. Reactor containment building
55
The inside surface of the RCB is lined with carbon steel plate with stiffeners to ensure leak-tightness for the containment. There are leak chase channels for the bottom liner plate and for inaccessible areas. The dome liner plate, in conjunction with dome tendons, is designed to serve as formwork for support of the concrete dome during construction. The walls, dome, and mat are designed using 34.47 MPa (5000 psi) concrete and 413.7 MPa (60,000 psi) reinforcing steel in accordance with the latest applicable code requirements. A 100-ton polar crane will be supported from the vertical building walls. The containment structure will have personnel access hatches and one large equipment hatch. The equipment hatch is required for movement of refueling equipment into and out of the containment.
The inside diameter of the RCB is determined by the PCRV diameter plus a 4.57 m (15 ft) annular space provision for the PCRV circumferential wire winding machine, steam and helium piping, ventilation equipment, etc. The inside diameters for the integral and intermediate loop RCB configurations-are given in Table 6.
The inside height of the RCB is determined by the height of the PCRV
plus adequate space provisions for heat exchanger removal, polar crane
installation, and the core refueling machine. To satisfy these require-
ments, and assuming a 2:1 ellipse for the inner dome surface above the spring
line, the dimensions shown in Table 6 were established for the RCB.
6.1.3 Thermal Barrier
The functional requirements of the thermal barrier are twofold. The
first requirement is to control the temperature levels and gradients within
the PCRV to prevent concrete degradation and to maintain acceptable levels
of thermal stress within the PCRV. This requirement is performed in con-
junction with the liner cooling system. Secondly, the thermal barrier
minimizes heat loss from the primary coolant system.
56
The thermal barrier and liner cooling system together must maintain
the concrete and liner within the limits given in Table 7. The total heat
loss through the thermal barrier into the liner cooling system should be
limited to a maximum of 0.5% of the core thermal power.
Figure 17 is a schematic layout of the HTGR/HTS reactor. The layout indicates the nominal temperatures expected in various parts of the loop during normal reactor operation.
The thermal barrier design HTGR/HTS reactor system is based directly
on the designs developed for the HTGR steam plants.
6.2 HELIUM CIRCULATORS
The limited scope of the thermal storage system study did not allow designs of the circulators to be carried out. However, maximum use of design data from existing steam plant technology has enabled sufficient circulator information to be generated in the cost area for inclusion in the system economic evaluation. With the circulators, the plant layout embodies the steam-turbine-driven helium compressor of the type designed for the 4000 MW(t) steam plant (4 loop plant). The thermal rating of this circulator [1000 MW(t)] is identical to the value selected for the HTGR/HTS system. In the HTS thermal storage plant there are two turbomachines (excluding pumps and electrical generators), namely the primary loop cir-culator for transporting the helium in the reactor coolant loop, and the secondary or intermediate loop circulator. Details of these circulators are given in the following sections.
6.2.1 Circulator Design Requirements
6.2.1.1 Primary Loop Circulator. The steam-turbine-driven primary circula-
tor installed in the PCRV transports the reactor thermal heat, a total of
2000 MW(t), to the IHX, where the heat is transferred to the secondary
loop. The primary loop circulator is positioned in the reactor helium
57
TABLE 7 LINER AND CONCRETE TEMPERATURE LIMITS
Normal Operating Conditions
Upset Conditions
Emergency Conditions Faulted
°C °F °C °F °C °F °C °F
High Liner Temperature <82.2 <180 <132.2 <270 <82.2 <360
Effective Concrete Temperature <65.6 <150 <86.1 <187 <148.9 <300
Hotspot Concrete Temperature
Bulk Concrete Temperature
<121.1
<68.3
<250
<155
<190.6
<90
<375
<194
<260
<154.4
<500 , v (<204.4
* 3 1 0 j,315.6
<400 ( a )
<600 ( b )
Minimum Liner Temperature 37.8 100 37.8 100 37.8 100
(a) Pressurized.
^ ) Unp r es sur iz ed.
GA-A14160
U! VO
dlbJL.
TEMPERATURES LOCATION
A B C D E
343.6 352.7 760 307.2 704.4
650.5 666,8 1400 585 1300
GA-A14160
Fig. 17. Schematic layout of the HTGR/HTS reactor indicating nominal loop temperatures
circuit between the IHX and the core. There are two primary circulators in the loop, and the helium flow to the circulator is received from two parallel-connected IHXs in each primary loop circuit. After compression the helium is transported to the reactor core. The primary system pressure loss (circulator pressure rise) was established at 0.143 MPa (20.7 psia), a value very close to that for the steam plant. With two circulators in the primary loop, the helium mass flow per unit is 466,9 kg/sec (1039 lb/sec), and the inlet temperature and pressure are 343.6°C (650.5°F) and 4.86 MPa (704.3 psia), respectively.
The primary system helium compressor is driven by a single-stage turbine. The full steam flow from the steam generator of 201.9 kg/sec (445 lb/sec) flows through the circulator steam turbine. After leaving the steam generator, the steam flows through the high pressure section of the electrical power producing steam turbine and enters the circulator turbine at 398.9°C (750°F) and 7.58 MPa (1100 psia). For the two primary loop circulators the power necessary to drive the helium compressors is 44.4 MW(e). With a turbine enthalpy drop of 110 kJ/kg (47.3 Btu/lb), the steam return conditions to the reheater are 334.4°C (634°F) and 4.43 MPa (643 psia). Details of the primary system design conditions are given in Table 8.
6.2.1.2 Secondary Loop Circulator. In the secondary loop the role of the external steam turbine-driven circulator is to transport heat from the primary loop IHX to the closed secondary helium circuit. The secondary loop helium compressor circulates the heated gas (from the IHX) to the helium-to-HTS exchanger and helium-to-steam generator. The system pressure loss of 0.145 MPa (21.0 psia) is almost identical to the primary circulator. With two circulators in the secondary loop, the helium mass flow per unit is 501.8 kg/sec (1106 lb/sec), and the inlet temperature and pressure are 307.2°C (535aF) and 4.85 MPa (704 psia), respectively, as shown in Table 8.
The helium compressor is driven by a steam turbine that takes a small
fraction (15%) of the flow compared with the primary system turbine. After
60
TABLE 8 HELIUM CIRCULATOR DESIGN CONDITIONS
Circulator Primary Loop Secondary Loop
Steam HTGR Design for Comparison
Core Thermal Power, MW(t) 2000 2000 4000 Number of Circulators 2 2 4 Circulator rating, MW(t) 1000 1000 1000
Helium Compressor .
Unit Mass Flow, kg/sec (lb/sec)
466.9 (1039) 501.3 (1106) 430.9 (1060)
Inlet Temperature, °C (°F) 343.6 (650.5) 307.2 (585) 312.8 (595) Inlet Pressure, MPa (psia) 4.36 (704.3) 4.85 (704) 4.37 (706) Pressure rise, MPa (psi) 0.143 (20.7) 0.145 (21.0) 0.132 (19.1) Inlet volume flow,
in.3/sec (ft3/sec) 123.2 (4391) 122.8 (4337) 120.1 (4243)
Number of Axial Stages 1 1 1
Steam Turbine Unit Mass Flow, kg/sec
(lb/sec) 201.9 (445) 30.9 (68) 300.8 (663)
Inlet Temperature, °C (°F) 398.9 (750) 391.7 (737) 379.4 (715) Inlet Pressure, MPa (psia) 7.58 (1100) 1.29 (188). 6.27 (910) Inlet Enthalpy, kJ/kg
(Btu/lb) 3154. 0 (1356) 3240. 1 (1393) 3126.1 (1344)
Enthalpy Drop, kJ/kg (Btu/lb)
110.0 (47.3) 725.7 (312) 69.3 (30)
Inlet Volume Flow, in.^/sec (ft^/sec)
7.30 (258) 7.13 (252) 13.14 (464)
Unit Power, MW(e) (hp) 22.2 (29,760) 22.5 (30,160) 22,380 (30,000) Number of Axial Stages 1 6 1
GA-A141G0
61
leaving the reheater, the steam flows to the intermediate-pressure electri-
cal power producing turbine (connected in parallel on the steam side).
Steam exiting from one of the intermediate-pressure turbine units at 1.29 MPa
(188 psia), and 391.7°C (737°F) drives the secondary loop circulator steam
turbine. For the two secondary loop circulators the power necessary to
drive the helium compressors is 45 MW(e). Because of the small steam flow
the enthalpy drop across the steam turbine of 725.7 kJ/kg (312 Btu/lb) is
very large, implying a multistage turbine, and this will be discussed in
a later section. Details of the secondary system design conditions are
given in Table 8.
Also included in this table are the thermodynamic conditions for a helium circulator of 1000 MW(t) loop rating that was designed for an HTGR steam plant. As outlined in the following sections, the similarity between the steam plant and HTGR/HTS plant circulator design conditions enabled meaningful circulator size and cost data to be generated without actually doing design work in this area.
6.2.2 Circulator Configuration
With two circulators in each of the primary and secondary circuits the
unit rating for these machines is 1000 MW(t). This gives a circulator rating
equivalent to that for the 4000 MW(t) SC-HTGR (4-loop), and advantage was
taken of the turbomachinery conceptual design work done for this plant. A
simplified view of the circular concept selected is shown in Fig. 18, and
salient features are given in Table 9. While the HTGR/HTS plant circulators
are similar to the 4000 MW(t) SC-HTGR, the thermodynamic conditions, parti-
cularly for the steam turbines, are different and the effect of this is
discussed later.
Figure 18 shows the vertically mounted circulator assembly. The com-
pressor and turbine rotors are integrally mounted on a single shaft and
overhung from a center bearing housing and seal section. The rotors are
fixed to the shaft by curvic couplings. There are two journal bearings,
62
.(!•'•« I —'1
snffiaf-A-.
GA-A14160
Fig. 18. Circulator conceptual design for 4000 MW(t) SC-HTGR (4-loop) plant [1000 MW(t)] circulator rating
TABLE 9 HELIUM CIRCULATOR CONCEPTUAL DESIGN
DETAILS FOR IOOO MW(t) RATING (HTGR UNIT)
Unit Parameter
Circulator Thermal Rating, MW(t) 1000
Helium Compressor
Number of Axial Stages Tip Diara, mm (in.) Hub Diam, mm (in.) Blade Height, mm (in.) Hub/Tip Ratio Rotational Speed, rpm
Blade Tip Speed, m/sec (ft/sec) Gas Axial Velocity, m/sec (ft/sec)
1 1270 (50.0) 83C.2 (33.0) 215.9 (8.5) 0.66 6100
405.4 (1330) 176.8 (530)
Steam Turbine
Number of Axial Stages Tip Diam, mm (in.) Hub Di un, mm (in.) Blade Height, mm (in.) Hub/Tip Ratio Blade Tip Speed, m/sec (ft/sec) Inlet Axial Velocity, m/sec (ft/sec) Power/Circulator, MW(e) (hp)
1
723.9 (28.5) 584.2 (23.0) 69.9 (2.75) 0.807
231.6 (760) 91.4 (300) 22,380 (30,000)
^ C o n c e p t u a l circulator design for 4000 MW(t) SC-HTGR. For details of cycle conditions and comparison with HTGR/HTS plant circulators, refer to Table 8.
GA-A14160
64
a thrust bearing, and seals (plus a shutdown seal) within the bearing housing. The journal and thrust bearings are shrouded step bearings, lubri-cated by high-purity water under pressure. The hybrid bearings function hydrostatically at start-up and low speeds as well as hydrodynamically at high speeds. Water was selected as the bearing lubricant primarily because it greatly simplifies the seal design and the auxiliary systems as compared with the use of oil. Among the secondary advantages of water is the low and uniform housing temperature resulting from the substantial bearing water flows and the fact that the boiler feed system can be used as an emergency supply. Shaft sealing on the compressor end is accomplished by a limited-leakage labyrinth seal supplied with a buffer flow of purified helium.
Figure 18 shows that the helium enters the circulator cavity in a plane belcw the bladed section, flows upwards in an accelerating section, and is compressed by the rotor and stator stages. Leaving the bladed section, the helium is vertically discharged into a long conical diffuser. The function of the diffuser is to attain maximum static pressure recovery as the helium flow is discharged from the compressor stator blade outlet. The geometry of this diffuser for a single-stage compressor is of particular importance since the velocity head of the fluid entering the diffuser represents a sizeable fraction of the entire pressure rise. As a result, the overall static-to-static compressor efficiency is very sensitive to the diffuser recovery effectiveness, which should be high for the design geometry shown. From Fig. 18 it can be seen that the circulator is provided with a shutoff valve, the function of which is to limit backflow through the coolant loop when the associated circulator is shut down. The valve is operated open or closed by three actuators. Steam flow to the circulators is through coaxial concentric pipes. Steam flows downward in the inner pipe through the turbine nozzles and into the rotor blades, is turned through IT radians (180 deg) and flows upwards in the annulus of the con-centric pipe. All of the service lines and instrumentation,* etc., enter from the top of the circulator assembly.
65
6.2.3 Circulator Installation
6.2.3.1 Primary Loop Circulator. The primary loop circulator is installed in a cavity in the PCRV as shown in Fig. 19. The circulators are rigidly mounted from an extension of the inner liner of the plug that acts as the closure of the cavity. There are no mechanical connections at the bottom
of the circulator assembly, but a simple seal is necessary at the end of the
diffuser section to prevent flow bypass. For the circulator configuration
proposed, installation and removal of the complete unit is a relatively
simple operation.
While thermodynamic conditions for the HTGR/HTS plant differ from the steam plant for which the circulators were designed, it can be seen from Table 8 that for the compressor, helium volume flow and pressure rise are quite similar. This implies that a single-stage compressor with very similar annulus geometries could be utilized. For the steam turbine the temperatures, pressures, and volumetric flows are quite different, implying that significant changes to the blading and annulus geometry are necessary. While the enthalpy drop for the HTGR/HTS plant steam turbine is greater than the steam plant design, it is postulated that to consider a single-stage turbine is still valid. The power levels of the two machines are virtually identical, and while internal flow path geometries will differ slightly (because of the different thermodynamic conditions) the overall envelope of the circulator will remain the same, and hence integration of this unit (Fig. 18) into the PCRV for the HTGR/HTS plant is valid. Cost data generated for this SC-HTGR circulator will be directly applicable for this current thermal storage study.
6.2.3.2 Secondary Loop Circulator. For the secondary loop the design
requirements for the circulator are less demanding since it is a nonnuclear
component and does not have an installation constraint imposed on it (as
in the case of the primary loop design where there is an obvious desire to
minimize cavity size within the PCRV). It is postulated that a commercially
avaiable steam-turbine-driven helium compressor could be used for this
66
STEAM IN
Fig. 19. View showing primary circulator installation in PCRV
application. Because of the limited scope of this study, however, a
survey of vendors for compressor information, particularly cost data,
was not possible. Because of this, the mechanical features and fluid
flow paths for the secondary loop circulator were taken to be identical
to the primary loop configuration, with installation of this vertical
turbomachine unit being in a relatively simple cylindrical steel pressure
vessel as shown in Fig. 20.
As mentioned above for the primary circulator, the thermodynamic con-
ditions for the compressor are very similar to the steam plant unit, and
with a single axial stage, only minor changes to the blading and annulus
geometries are necessary. For the steam turbine, however, the much reduced
inlet pressure, and the much higher enthalpy drop, imply a significantly
different machine. While no design work has been performed, it is postulated
that six axial turbine stages would be necessary. Clearly, this would result
in a much heavier rotor with substantially higher bearing loads, and indeed,
this would possibly extend the bearing design parameters beyond the range
of current test data. These aspects were not investigated in this study,
and for the purpose of sizing (and costing) the steel pressure vessel, the
envelope for the steam cycle circulator (Fig. 18) was retained. In the
circulator cost estimate account was taken for the increased number of steam
turbine stages.
The above circulator work was considered adequate to support the system
economic evaluation and plant layout study. In any follow-on effort, work
on the secondary loop circulator would be conducted in the areas of 1) investi-
gation of steam system rerouting to reduce turbine enthalpy drop, and
2) availability and cost of a commercial steam-turbine-driven helium com-
pressor.
6.3 HEAT EXCHANGERS
. Although detailed design was beyond the scope of this study, sufficent
engineering effort was carried out to generate conceptual heat exchanger
68
STEAM OUT
CIRCULATOR INSTALLED IN STEEL PRESSURE VESSEL
HELIUM OUTLET (TO IHX)
HELIUM INLET (FROM He/HTS EXCHANGER AND He/HnO STEAM GENERATOR)
a.?.:?.:-:-:-:- --v:^* > -•• •• GA-A14160
Fig. 20. View showing installation of secondary loop circulator in steel pressure vessel
69
designs in support of the system and cost analyses. Since the main thrust of this effort was directed primarily at identifying thermal sizes based on representative surface geometries and flow configurations, these designs, which are discussed in the following sections, are not necessarily optimized and should be considered conceptual in nature.
6.3.1 IHX
The IHX serves as the thermal interface between the primary helium reactor coolant and the secondary helium flowing in the intermediate heat transport loops external to the PCRV. Four such heat exchangers are required for the 4-loop plant. The technical approach adopted for this unit is based on that followed by GA in establishing the process heat HTGR IHX design, which bears a strong functional similarity to this application. The design effort therefore concentrated upon a plain tubular, pure counter-flow heat transfer matrix of modular construction, an approach offering the general advantages of compatible dimensional proportions for PCRV installa-tions, low-pressure losses, conventional construction, minimum transverse metal temperature gradients in the tube bundle, low potential for flow-induced tube vibrations, and relative ease of handling and maintenance.
General design features and thermal requirements for the IHX are summarized in Table 10. The fluid temperature requirements indicate that the steady-state metal temperature at the hot end of the heat exchanger will be approximately 732°C (1350°F), which is high enough to necessitate the use of 316 stainless steel as the tube and center return duct material. At the top end of the unit the metal temperatures are low enough to permit the use of ferritic alloys. Because the IHX is nearly pressure-balanced between the primary and secondary circuits during normal operation and experiences relatively light tube loading during short-term, abnormal operation (one circuit depressurized), practical handling considerations set the tube wall thickness at 0.6 mm (0.025 in.).
The IHX thermal design calls for a contiguous array of 168 hexagonal modules located on a pitch of 25 cm (9.74 in.), each module containing
70
TABLE 10 IHX DESIGN SUMMARY
Quantity per Plant Thermal Requirements
Heat duty per IHX, MW(t) LMTD, °C (°F)
UA, MW/°C, (Btu/hr-°F) pt:r HX UA, margin, %
Performance Fluid Flow per HX, kg/hr (lb/hr) Inlet temp, °C (°F) Exit temp, °C (°F) Inlet pressure, MPa (psia) AP, MPa (psid)
Mechanical Design Flow configuration Surface geometry HX o.d., m (ft) Effective tube length, m (ft) No. of Hex modules per HX Tubes per module Module pitch (A/F), mm (in.) Tube o.d. x wall, mm (in.) Tube p/d
Return duct i.d. x wall, mm (in.)
Materials, Return Duct Lower lead tubes
Module subheaders Module tubing Upper lead tubes
Primary tubesheet 2 2
Surface area, m (ft ) per HX ASME' Code Classification
4
507 54.8 (98.7) 13.01 (24.67 x 10 6
10
Primary He
852,000 (1.878 x 10 6) 757.2 (1395) 343.6 (650.5) 4.91 (712) 0.034 (5)
Counterflow Plain tubular 3.85 (12.6) 12.95 (42.5) 168 331
247.5 (9.74) 9.5 x 0.63 (0.375 x 0.025) 1.40 equilateral triangular 832 (32.8) 316 SS 316 SS 316 SS
316 SS (welded) 2-1/4 Cr - 1 Mo 1/2 Mo-Cr-Ni forging 21,550 (232,000)
Section VIII
GA-A14160
Secondary He 902,500 (1.99 x 10 6) 315.8 (600.5) 704.4 (1300) 4.91 (712) 0.055 (8)
71
(169) 9 mm (3/8-in.) o.d. x 0.6 mm (0.025 in.) wall x 13.2 m (42.5 ft)
effective length tubes pitched on a 1.40 equilateral triangular p/d. The
estimated pressure losses of 55 kPa (8 psi) and 34 kPa (5 psi) for the
secondary and primary sides, respectively, include parasitic losses (e.g.,
entry/exit and tube support losses) in addition to friction and flow
acceleration in the tube bundle.
The conceptual design of the IHX is shown in Fig. 21, which shows the unit mounted vertically within a reactor vessel side cavity. The primary structural interface between the reactor vessel and the IHX is the periphery of the upper tube sheet. This large-diameter tube sheet is torospherical because it forms a primary-to-secondary helium pressure boundary and the large differential pressure forces that can exist across it create unaccept-ably large thickness requirements for flat tubesheets. The tubesheet rests upon a thermal sleeve that is welded to the cavity liner and transmits all the dead weight and the majority of live IHX loads to the reactor vessel through this mount. A seal weld ring at this joint eliminates primary/ secondary helium leakage. The tubesheet permits the upper portion of the IHX cavity to contain only secondary inlet helium, which facilitates access for tube plugging and in-service inspection operations. The tubesheet manifolds 168 lead tubes of 127 mm (5 in.) diam to the upper ends of the 168 heat exchanger modules. These upper lead tubes gain longitudinal flexibility by being helically wound about a central secondary helium out-let duct. This duct is structurally welded to and penetrates the center of the sheet and acts as a central load-carrying member for the IHX.
The lower lead tubes, which operate at relatively high |>732°C (1350°F)]
metal temperatures, simply extend from the lower subheaders straight down-
ward to a gusseted lower lead tube support plate. The lead tubes pass
through the support plate and curve inward to terminate at a cylindrical
tubesheet that is integral with the central helium outlet duct. Direct
impingement of reactor outlet helium on the lower lead tubes is prevented
by shrouding the tubes below the support plate with an ellipsoidal deflector
shell.
72
xitiet ::ONO«I UM "ESS 4.55K .0' a . "04 PSIA)
INLET SECONDARY HEUUM TEMP 3l56"Ct60a5"FK f PRESS <91* 10'P. -y-M l! (712 PSIA l [ j
44ZO00<l«'-6">-LINER
PRIMARY HELIUM TO CIRCULATOR —
TEMP 343.6 'C 1650.5" F I
PRESS 4.87X10* P» (707PS1A1
COILED U«ER LEAD TUBES -IZ70Dl5lN),4 Iw»at-I6INI 168 QEQD
iota;j75i -WELDED STEEL CLOSURE
CAS 8TRASS SEAL ASSY-1 T1P» GUSSETED LOAES LEAD TUBE SUPFOGT PLATE INLET PRIMARY HELIUM-
TEM= m . r c 11395* Ft PRESS * 91 * 10" ft / "V • 7.2 PSIA) ' LOWER L£A3 TUBES -1J7 00 ISIN >,4.1 WALL U 6 I M 168 REQD
VEGTI<7AT SFCTIOW THRU INTFRMEDIATE H?AT EXCHANGES
SCALE R - I - Q '
9 5 2 5 0 0 ' 375IN|X64»MLL<025IN> SSI TUBES PER MODULE PITCHED ON 13 .335 -1525 INI EQUIU 'ESAL TRIANGLE ARRAY
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' 68 MODULES PITCHED ON 2 « 7 « 8 " ( 9 . 7 4 3 3 I N I EQUILATERAL TRIANGLE / ARRAY ^
THERMAL BARRIER
L'5.4J*SEF< 14.5466 W
2M.65 (9.36=. I ' .
A scAtf •-. ~-o" IHX MODULE ARRANGEMENT
\OTE I ALL DIMEMSTLS ABE IN
MILLIMETERS UNLESS CTHEBAISE S-EC:FIEC
GA-AH160
Fig. 21. Conceptual design of the intermediate helium heat exchanger
The entire IHX cavity is closed at the top by a welded steel closure
that features concentric secondary helium penetrations and redundant closure
retention bolts and seal rings. Although the top of the plug is shown
partially outside the PCRV in Fig. 21, it is actually below the reactor
operating floor level.
The primary helium flow path in the IHX cavity is entirely on the shell side of the IHX. Primary circuit helium at 757°C (1395°F) enters the cavity through a bottom cross duct that directs the hot helium upward toward the deflector. The hot helium, flowing around the deflector but inside the IHX and failing to bypass the gas seal rings, enters the IHX module shrouds and flows upward to exit at their upper ends. Having thus been cooled to 344°C (651°F), the primary helium leaves the IHX cavity through a cross duct leading to the circulator cavity.
Secondary helium at 316°C (600°F) enters the IHX cavity above the main
tubesheet through penetrations in the cavity closure. The helium enters
the upper lead tubes and flows to the individual modules, where it flows
downward through the 55,608 HX tubes. At the lower end of the modules,
the helium heated to 704°C (1300°F) enters the lower lead tubes and flows
to the central return duct. The heated secondary helium exits via the
central duct upward through the tubesheet and the cavity closure to the
external intermediate heat transport loop.
6.3.2 Helium/Steam Heat Exchangers
The helium-heated steam generator and reheater, which are outside
the PCRV, use thermal energy from the intermediate transport (secondary
helium) loop to generate steam at the throttle conditions required for the
high-pressure turbine and intermediate-pressure turbine, respectively.
The similarity of the operating conditions for these units to those of
their counterparts in an SC-HTGR provides the basis for applying existing
HTGR steam generator technology to the design of these components. The
chief difference between this application and the SC-HTGR components is
74
the additional premium placed on the heat exchanger envelope in the latter due to PCRV installation conisderations. The design approach taken in this
study accounted for these considerations as well as the limitations on time and effort by adopting the following groundrules.
1. The steam generator and reheater are connected in series on the
helium side.
2. The design effort considers two sets of components Cone set per
external loop) to enhance plant availability and component
transportability.
3. The steam generator and reheater are separate, free-standing com-ponents because there is little incentive at this point to accept the complexity and mechanical interaction associated with packaging one within the other to achieve a single unit.
4. The steam generator is a once-through design, comprising helically wound tube bundles with the water/steam flowing inside the tubes. To help ensure boiling flow stability, a bottom-fed, vertical bundle orientation was selected, and the same inlet orificing criteria as developed for HTGR steam generators were applied.
5. 'j'jje is a straiRht-tubular, pure counterflow design to
illllHitll e | f-'miii'/iJi'tJti lilol'-fji temperature gradients, shell-side pres--'!' ill!! j-lmHlUillhbrt VlltfatidUNi
'!l)e f)|]§lff||; (I) (4 tftij^lMtiim f!Mtj (iUUPWl- I M l fen features for both components
ttfe sUiiiiiidtized lit Trtb'ln j.j * H|iiatijMt! £i§|muiiB of l-lm d§r!-{J!in ni'f rtisQllnned
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6.3.2.1 Steam Generatdi- Design. The basic design features of the steam
generator are illustrated schematically in Fig. 22. The main bundle for
this unit was thermally sized through the use of NUSIZE, an internal GA
75
TABLE 11 HELIUM-TO-STEAM HEAT EXCHANGER DESIGN SUMMARY
Heat Exchanger Steam Generator
, .. — Reiieater
Quant i ty per P l a n t 2 2
Thermal Design:
Heat duty per p l a n t , MW(t) 1013 197
UA Margin, Z 18 15
F lu id Helium Uater /Steam Helium Steam
P lan t f low, k g / h r ( l b / h r ) 2 .12 x 10 6 (4.676 x 106) 1 .45 x 106 (3 .2 x 10 6 ) 2 .12 x L0 6 (4.676 x 106) 1 .45 x 10 6 ( 3 .2 x 10 6 ) I n l e t t empera tu re , °C (°F) 638.3 (1181) 187.2 (369) 702 .2 (1296) 334.4 (634)
Ex i t t empera tu re , °C (°F) 307.7 (585.9) 512.8 (955) 634 .4 (1174) 538.9 (1002)
I n l e t p r e s s u r e , MPa ( p s i a ) 4 .89 (710) 20.4 (2965) 91 (712) 4 .43 (643)
Lead- in ( o r i f i c e ) AP, MPa (ps id ) 1 .74 (252) —
Bundle AP, MPa (ps id ) 0.C41 (6) 1 .38 (200) 0. 014 (2 .0) 0 .19 (28)
Ove ra l l AT, MPa (ps id ) 0 .041 (6) 3 .11 (452) 0. 014 (2 .0 ) 0 .19 (28)
Mechanical Design:
Flow c o n f i g u r a t i o n Mul t ipas s c ro s s coun te r f low Counterf low
lube bundle type H e l i c a l once- through S t r a i g h t
Bundle, o . d . , m ( f t ) 3 .81 (12 .5) 1 .60 (5 .25)
Bundle, i . d . , si ( f t ) 1 . 31 (4 .3) —
Bundle h e i g h t , m ( f t ) 12.19 ( 4 0 ) ( a ) 12.37 (40.6)
Tube l e n g t h , m ( f t ) 141.7 (465) 12.37 (40.6) ( E f f e c t i v e )
Number of tubes 391 370
Tube o . d . x w a l l , mm ( i n . ) 25.4 x 3 .81 (1 .0 x 0 .15) 38 .1 x 3 .81 (1 .50 x 0 .15)
Type p i t c h i n g I n - l i n e r e c t a n g u l a r E q u i l a t e r a l t r i a n g l e
P/d 1 .66 t r a n s v e r s e , 1.50 l o n g i t u d i n a l
2 .0
Economizer /evapora tor s e c t i o n tube m a t e r i a l
2 - 1 / 4 Cr - 1 Mo —
Superheat s e c t i o n tube m a t e r i a l Incoloy 800 Incoloy 800 2 2
S u r f a c e a r e a per p l a n t , m ( f t ) 8547 (92,000) 1096 (11,800)
ASHE code classification Sec t ion VI I I S e c t i o n V I I I
GA-A14160 I n c l u d e s 0 .76 m ( 2 . 5 f t ) a l lowance f o r b i m e t a l l i c weld .
STEAM OUT
GRAVITY
HELIUM IN
HELIUM OUT
MAN-ACCESS HATCH " \ (OPTIONAL)
SUPERHEATER SECTION (INCOLOY 800) HELICAL BUNDLE
BIMETALLIC JOINT TRANSITION AREA
ECONOMIZER EVAPORTATOR SECTION (2- 1/4 CR - 1 MO) HELICAL BUNDLE
1,32 m (4.33 FT) APPROX. DIAM
FEEDWATER LEAD-INS
•3.81 m (IZ.5 FT) D I A M - H APPROX. DIAM
FEEDWATER INLET
GA-A14160
Fig. 22. Schematic of the helium/steam generator
77
computer code used extensively in HTGR steam generator optimization studies.
The selected design uses up both the helium and steam side allowable pres-
sure drops while providing heat exchanger dimensions that fall within con-
ventional fabrication and transportation limits.
Hot helium enters the steam generator through a vessel side opening near the top, flows downward over the coiled tube bundle, transferring its heat to boil the water flowing upward inside the tubes, and then discharges through a similar opening near the bottom. The feedwater enters the main bundle at the bottom via a set of small-diameter lead-in tubes, sized in accordance with inlet orificing criteria for hydrodynamic stability, receives heat during its upward passage inside the main bundle tubes, and leaves at the top in the form of superheated steam.
The main bundle comprises 391 helically coiled tubes, 25.4 mm (1 in.) o.d. x 3.8 mm (0.15 in.) wall x 141.7 m (465 ft) developed length, arranged on in-line pitching to produce overall effective bundle dimensions of 11.4 m (37.4 ft), 3.81 m (12.5 ft),, and 1.31 m (4.3 ft) on height, outer diameter, and inner diameter, respectively. The average transverse and longitudinal tube pitch-to-diameter ratios of 2.04 and 1.50 are consistent with current helical bundle steam generator design practice, and the innermost tube bend radius exceeds that necessary to avoid thinning the tube wall. The tube material selected for the economizer and evaporator sections is 2-1/4 Cr -1 Mo, a relatively inexpensive ferritic alloy. In the superheater section the metal temperatures exceed the capabilities of ferritic alloys, neces-sitating the selection of Incoloy 800. A space allowance of 762 mm (30 in.) axial bundle length should be added to the bundle dimensions cited above for accommodating the required bimetallic joint. The bundle is contained within an internally insulated mild steel pressure vessel that is 82.6 mm (3-1/4 in.) thick and is headered into conventional flat tubesheets. Differ-ential thermal expansion between the bundle and shell is taken up by a metal bellows in the main steam line, and sliding seals prevent helium by-passing of the bundle. Man-access hatches may be incorporated for main-tenance purposes in both the top and bottom vessel heads for inspection
78
and repair. Preliminary indications are that the bundle will require at least six radial tube support plates to help suppress tube vibrations and to maintain adequate tube spacing. In addition, a circumferential baffle midway through the bundle annulus may be required to prevent the possibility of acoustic excitation of the tube bundle.
6.3.2.2 Reheater Design. The reheater design follows a typical tube-in-shell approach, as illustrated schematically in Fig. 23, with the helium on the shell side and the steam inside the tubes. Special design features include 1) semitoroidal plenums in the shell at the helium entry and exit to enhance flow distribution, and 2) baffling as necessary to protect the tube bundle and tube/tubesheet joints at the hot end from direct impingement by the hot helium. The thermal sizing of the Incoioy 800 straight tube, pure counterflow heat transfer matrix was accomplished by generating a family of NUSIZE design solutions obtainable for representative tube pitch-ing (2.0 triangular p/d) and a typical tube o.d. of 38.1 mm (1-1/2 in.), using frontal area as the dependent variable. The selection of Incoioy 800 as the tube material is based on creep-buckling considerations arising from the higher helium pressure level relative to that of the steam. The design in Fig. 23 uses all of the available helium side pressure drop and has the smallest surface and frontal area requirements of the configurations considered, suggesting that follow-on detailed thermal size/cost tradeoff studies be aimed toward smaller frontal area as well as alternative tube diameters and pitching.
With two external reheaters per plant, the resulting bundle dimensions are 1.6 m (5.25 ft) diam by 12.4 m (40.6 ft) effective length, with each bun-dle comprising three hundred seventy 38.1 mm (1-1/2 in.) o.d. x 3.8 mm (0.15 in.) wall Incoioy 800 tubes pitched on a 2.0 p/d equilateral triangular array. The bundle fits within a 38.1 mm (1-1/2 in.) thick cylindrical shell that is insulated internally to permit, the use of mild steel, and the entire assembly is transportable by conventional methods. Support of the individual tubes within the bundle will be accomplished with cellular-type spacer grids, similar to those used to support light-water reactor (LWR) fuel
79
STEAM OUT-TEMP 538 9
- F L O A T I N G H J B E S H E E T »l MAY eE LOOTED »T COLD END IF ORIENTATION Of HX IS HOBIZOMTAL II TOUSLE SEALS WOVtOE OPTIONAL
Wrc-EO A«EA TO PfiOJIDE LEAKACE
NOTES:
I . ALL DIMENSIONS ARE IN MILLIMETERS UNLESS OTHERWISE SPECIF IED
TEMP 702.2'C CI296'FI PRESS 1.91 X 10* ft (712 PSIA L ^
FLOW 2 12X10, K I / W R C M.676XKTLB/,HP<
00 o
A f -ELIUM OUT
TEMP 6 3 7 . 8 ' C CII90" F) PRESS &J4XI0'Fl 1710 PSIA)
STEAM IN • TEMP 3 3 4 . 4 ' C
( 6 3 4 ' F I . PRESS 4 . 4 3 X 1 0 P.
< 6 4 3 PSIA I FLOW M 3 X I 0 KL/HR 1323?
THERMAL BARRIER
GA-A14160
Fig. 23. Conceptual helium/steam reheater design
elements, located at discrete axial intervals. For this preliminary study an additional pressure drop allowance of 25% of the friction loss on the shell side of the bundle was made to account for parasitic losses, and a corresponding 10% allowance was included on the steam side.
6.3.3 Helium/HTS Heat Exchanger
The He-to-HTS heat exchanger is a straight tube, single-pass, counter-flow assembly. A cylindrical vessel headering concept was chosen to cope with the high-pressure differential/long life/high-temperature requirements (Table 12) that, due to the resulting low allowable stress levels, obviate the use of more conventional tube sheets. Twenty-five thousand 1.3 cm (1/2-in.) o.d. tubes are used in each at the two heat exchangers for the plant. The required tube wall thickness is 0.3 mm (1/32-in.) and is dictated by the hot end material temperature and the pressure differential of about 4.8 MPa (700 psi). The tube pitch was set at 1.4 times the o.d., which was taken as a lower practical limit. The high density of the HTS which flows on the shell side allows low tube pitch values without undue friction pressure losses.
A tradeoff was made regarding number of tubes, tube length, and
resulting pressure drops and heat-transfer coefficients. Twenty-five thousand tubes per exchanger was chosen which, with the chosen tube size and pitch, Sets the fluid velocities and establishes the surface area
2 2 required. The required surface is 23,700 m (255,000 ft ), which is achieved in the 11.9-m (39-ft) long tubes. This results in a tube side (helium) friction loss of 31 kPa (4.5 psi) and a shell side (HTS) loss of about 3 kPa (1/2 psi).
The heat exchanger is divided into two halves (see Fig. 24), each packaged within a common insulated sheetmetal shell. Each half has a helium inlet and discharge header into which the ends of the heat exchanger tubes are welded. Both inlet and discharge headers have the same 87 cm (34.4 in.) o.d. The inlet headers, however, have a wall thickness of
81
TABLE 12 PEAKING PLANT HELIUM-TO-HTS HEAT EXCHANGER
Each Heat Exchanger (Two required per plant)
Total power 425 MW (1.45 x 10 9 Btu/hr) Number of tubes 24960 Tube active lengths 11.89 m (39 ft) Tube o.d. 12.7 mm (1/2 in.) Tube wall thickness 0.8 mm (1/32 in.) Tube pitch 17.8 mm (0.7 in.) Tube pitch pattern Triangular Tube bundle width 1.55 m (61 in.) Tube bundle depth 4.82 m (15.8 ft) Active bundle surface area 11859 m 2 (127,650 ft 2) Flow area (He-tube side) 2.43 m 2 (26.1 ft 2) Flow area (HTS shell side) 3.68 m 2 (39.6 ft 2) He temperature in 704°C (1300°F) He temperature out 307°C (585°F) HTS temperature in 288°C (550°F) HTS temperature out 543°C (1010°F)
He pressure in (header inlet) 4.91 MPa (712 psia) He pressure out (header discharge) 4.83 MPa (701 psia) HTS pressure in (manifold inlet) 0.4 MPa (^60 psia)
He flow rate 744.8 kg/hr (1.642 x 10 6 lb/hr)
HTS flow rate 3832.9 kg/hr (8.45 x 10 6 lb/hr)
Materials:
Tubes, He inlet header, internal baffles, pressure shell
316 Stainless Steel
Discharge header, insulation cover plates
Carbon Steel
Insulation Kaolin wool and fiberglass
GA-A14160
82
- i t INLET HEADER or- 874M*'J« 4*) T«. i09UU(4 3T
t"™" ro4'c (i30C*ri J
ZZT v 3 1 Z L
J ..j..... -j
iz: i
saw nzi
ECIION A * A
NOW SAFFLE L VERTICAL Support MEMBER
f l o w b a f f l e s i s h e l l RESTRAINT MEMBERS
39RANCH MANIFOLD id*&o«hu " 9 t 2 » U U L W )
GA-A14160
Fig. 24. Conceptual design of helium/salt heat exchanger
11.0 cm (4.3 In.) while the discharge header walls are only 3.2 cm (l-l/4-in.). The tubes are welded to the inner surface of the drilled headers in rows which have a circumferential pitch of 5 cm (2 in.). The rows are spaced longitudinally about 0.6-in. apart (the same as the tube rows in the tube bundle) and the holes in adjacent rows are staggered. There are 40 tubes per circumferential row and 312 rows per header. The helium is introduced into one end of each of the two inlet headers, the other end being closed. Similarly, the cooled helium is discharged from one end of the two discharge headers. Flow splitter manifolds join the appropriate ends of each header pair.
The cool HTS is introduced at the bottom of the tube bundle through
a manifold feeding three 30.5 cm (12-in.) i.d. inlet ports spaced equally
along the 16-ft base of the tube bundle. Three similar ports, but with
35.6 cm (14-in.) diameters, are provided at the top of the heat exchanger
to effect the HTS discharge to the discharge manifold. The HTS flows into
the shell cavity between the two helium discharge headers and into the main
heat transfer area. The HTS is allowed to circulate around the helium
headers and tubes in the inlet and discharge sections. Two 1.9-cm (3/4-in.)
thick sheet metal baffles are included in the main bundle and header areas
to resist any cross flow or recirculation tendencies in the HTS. These
baffles are welded to the cylindrical intersections along the sides, giving
internal support to withstand the head load of the HTS. The baffles are
horizontally segmented to allow for a difference in thermal expansion
between the outer cylindrical walls and the baffles.
The headers are supported at the penetration ends by the shell end
plates to which they are welded. The closed ends, which have opposing
juxtaposition for inlet and discharge headers for pressure loss balance
purposes, are also supported from the shell end plates, but by means of a
sliding bracket. The distance between any two of the four headers is there-
by controlled by the shell end plates. Differences in longitudinal expansion
between these end plates and the heat exchanger tubes are absorbed by the
flexibility provided by the bends in the tubes. Lateral expansion difference
84
is quite small and will likewise be absorbed by tube bend flexibility. The tube positions relative to each other are maintained by spacer grids located at approximately 1.2 m (4 ft) intervals in the main bundle sections. These grids are fixed to the baffle plates.
The outer shell m the main tube bundle region is formed from 1.2 cm (l/2-in.) plates rolled into cylindrical segments of about 1.7 rad (100 deg), which when welded together form scalloped-shaped sides with cylindrical ends. The two internal baffles are welded to the apex of the convolutions to resist spreading of the two sides of the heat exchanger. Cylindrical sections are also used for the shell around the inlet and exit headers and tubes as shown in Fig. 24. These are joined to the scalloped sides and to spherical end caps to form the entire external pressure shell. The maxi-mum pressure on this shell is approximately 400 kPa (60 psia) at the inlet manifold. Flat 1.9-cm (3/4-in.) sheet baffles are included to contain the HTS flow within the cube bundles. These are welded to semicircular baffles that close off the cylindrical sides to HTS flow.
Ten to 15 cm (4 to 6 in.) of thermal insulation must be added to the outside surface of the shell to minimize heat losses and to provide a safe temperature level to the exposed surface. One or more of several industrial moderate temperature fibrous insulating materials such as Kaolin wool, glass fiber, etc., can be used as the insulation, sandwiched between the shell and thin steel cover plates hung from insulated studs attached to the shell.
Detailed dimensions of the heat exchanger are shown in Table 12. Mate-
rials are also included, which indicates extensive use of stainless steel.
Carbon steels and other alloy steels are avoided at temperatures above 371°C
(700°F) when in contact with the KTS because of corrosion problems.
6.3.4 HTS/Steam Heat Exchangers
The steam generator and reheater components in the peaking power plant
are functionally identical to their helium-heated counterparts in the
85
base-loaded plant, except that the thermal energy required to generate the steam is extracted from the HTS being pumped from the hot storage tank. Even though' the steam outlet conditions for these units are similar in both the peaking and the base-loaded plants, the HTS-to-steam heat exchangers must operate with a lower temperature heat source to produce over twice the thermal output, thus creating considerably larger overall conductance (UA) requirements. Despite the higher heat transfer coefficients obtainable with HTS as compared to helium, the surface area requirements are still very large. The size of the steam generator, in particular, suggests that further thermodynamic cycle tradeoff studies should be included in any follow-on effort.
The design approach and general groundrules followed for these com-ponents are the same as those discussed in Section 6.3.2 for the He/steam heat exchangers, except that more than two sets of steam generators appear to be advisable to help keep the components within fabrication and shipping limits.
The operating conditions and general design features for the steam
generator and reheater are summarized in Table 13. Specific aspects of
the designs are discussed below.
6.3.4.1 Steam Generator. From a conceptual standpoint, the HTS-heated steam generator is schematically identical to its helium-heated counter-part shown in F:!g. 22. The once-through helical tube bundle approach appears to be particularly suited to this component because it offers the favorable combination of a near-counterflow log mean temperature difference (LMTD) with a relatively high overall heat transfer coefficient (U) to help keep heat transfer surface area requirements to a minimum. Both the tube and pressure vessel wall thickness requirements benefit from routing the high-pressure water/steam fluid inside the tubes with HTS at near-ambient pressure levels on the shell side. For these reasons, the selected flow arrangement calls for multipass crosscounterflow heat transfer to take place between the HTS, which makes one shell-side pass over the helically coiled tubes, and the two-phase water/steam mixture flowing counter-currently inside the tubes. Q(-
TABLE 13 SALT/STEAM HEAT EXCHANGER DESIGN SUMMARY
Heat Exchanger Steam Generator Reheater Quantity per Plant 8 2 Thermal Design:
Heat duty per plant, MW(t) 2125 424 UA margin, % 15 15 Fluid Plant flow, lb/hr
HTS 50.71 x 106
H2O 6.835 x 106
HTS 50.71
H2O x 106 6.835 x 106
InLet temperature, °F 933.5 370 1010 585 Exit temperature, °F 550 900 933. 5 950 Inlet pressure, psia (55 psia) 2290 (60 psia) 595 Pressure drop, psi 5 270 5 35
Mechanical Design: Flow arrangement Helical once-through
Multipass cross-counterflow Straight tube counterflow
ASME code classification Section VIII Section VIII Heat exchanger section Economizer/
Evaporator Superheat Reheat Bundle o.d., ft 11 11 5.5 Bundle i.d., ft 6.8 6.8 (NA) Bundle height, ft 80 29 57 Tube length, ft 510 184 57 No. of tubes/heat exchanger 238 238 1900 Tube o.d. x wall, in. 1.0 x 0.15 1.0 x 0.15 1.0 x 0.10 Tube pitching 1.40
triangular 1.40
triangular 1.40
triangular Tube material 2-1/4 Cr-1 Mo Incoloy 800 2-1/4 Cr-1 Mo Surface area, ft /unit 30,000 10,840 28,360
GA-A14160
87
Since the overall height, diameter, and weight of the tube bundle thermally sized to meet the preliminary cycle conditions adopted for this study far exceed conventional handling and transportation limitations, a modular approach has been employed to subdivide the steam generator into eight individual assemblies (four per loop), each of which is separated further into a superheater section and an economizer-evaporator section. The additional headering complexity involved in this style of modulari-zation is offset partially by the elimination of the need for bimetallic tube joints at the material transition between the Incoloy 800 super-heater and the 2-1/4 Cr - 1 Mo economizer-evaporator sections. With this arrangement the superheater section comprises 238 helically wound 2.5 cm (1-in.) o.d. x 56 m (184 ft) long tubes pitched on a 1.40 equilateral triangle p/d and packaged in a 3.4 m (11 ft) o.d. x 2.1 m (6.8 ft) i.d. x 8.8 m (28.9 ft) high bundle. The tube bundle for the economizer-evaporator section is geometrically identical to the superheater, except that a tube length of 155 m (510 ft) is to be accommodated within a bundle height of 24.4 m (80 ft). While these parameters fall within handling and transportation limits, additional study of the fabrication parameters involved in helically winding the economizer-evaporator tubes into their tube support plates should be included in any follow-on effort.
6.3.4.2 Reheater. The approach taken.in the design of the HTS/steam
reheater was identical to that of the He/steam reheater discussed in
Section 6.3.2.2, except that it was necessary to pitch the tubes more
closely together to avoid size penalties attributable to low HTS velocities.
For the reasons discussed in Section 6.3.4, the HTS/steam reheater is con~
siderably larger than its helium-heated counterpart, requiring almost
five times the heat transfer surface area. This size increase was absorbed
primarily in the increased active length of the bundle [17.3 (57 ft) compared
to 12.4 m (40.6 ft)] and the tighter pitching [1900 2.5 cm (1 in.)] o.d.
tubes compared to 370 3.8 cm (l-l/2-in.) o.d. tubes. While the low HTS
pvessure level results in minimal shell thickness requirements, sub-
stantial tubesheet thicknesses—on the order of 27.5 cm (11 i n . ) — a r e needed
to accommodate the large operating pressure differential between the tube-
side ;fluid (steam) and the shell side fluid (HTS). Further study is required
88
to determine whether or not measures to reduce this tubesheet thickness (e.g., going to ellipsoidal or hemispherical configurations to reduce the bending component) justify the additional fabrication complexity that may be Involved. Unlike the He/steam reheater, the tube material selection for this unit is not governed by creep-buckling considerations, and it is therefore possible to specify 2-1/4 Cr - 1 Mo tubing, which is considerably less expensive than Incoloy 800. With these provisos, the HTS/steam reheater is conceptually identical to the He/steam reheater shown schemati-cally in Fig. 22. Vital statistics and basic design details for this unit are summarized in Table! 13.
6.4 SALT LOOP COMPONENTS
The conceptual design information for the salt loop components was
identified in support of the system and cost analysis. Detailed design
studies were beyond the scope of this study; however, sufficient engineering
effort was carried out to determine the technical, availability of the salt
loop components.
6.4.1 Heat Storage Tank Design
The storage tanks are free-standing cylindrical vessels made from welded steel plates and are conventionally designed except that the "hot" fluid tanks are made of stainless steel. The wall of each tank is tapered in thickness from base to top by using successively thinner plates to suit the reducing hoop stress. All tanks are insulated on their external surfaces to reduce the surface temperature to 66°C (150°F). This minimizes heat loss and the danger to personnel.
Design features of the storage tanks and principal dimensions are listed in Table 14.
The roof of each tank is supported by radial and auxiliary Warren girder trusses. The girders are mounted on vertical columns welded to the inner wall of the tank and on a central A section Warren girder column assembly.
TABLE 14 HEAT STORAGE TANKS
Hot Tank Cold Tank
Diameter (inside) 36.6 m (120 ft) 36.6 m (120 ft)
Height (to roof truss) 17.i m (57 ft) 16.1 m (53 ft)
Wall Thickness at Base • • 6.7 cm (2.65 in.). 6-7 cm (2.65 in.)
Wall Thickness at Top 2.5 cm (1.0 in.) 2.5 cm (1.0 in.)
Wall Material 316 Stainless Steel (type) Carbon Steel
Outside insulation 5.1 cm (2 in.) ceramic fibre 2.5 mm (+0.1 in.) cover plate
5.1 cm 2.5 mm
(•2 in.) mineral wool (+0.1 in.) cover plate
Roof 1.2 cm (1/2-in.) stainless steel plate
1.2 cm (1/2-in.) carbon steel plate
Roof Truss [12 radial, 10.1 cm (4 in.) deep Warren girders, welded]
[24 auxiliary, 10.1 cm (4 in.) deep Warren girders, - -Ided]
Roof Truss Material 316 Stainless Steel (type) Carbon Steel
Center Support Pillar (A section Warren girders, welded)
Center Support Pillar Material 316 Stainless Steel (cype) Carbon Steel
Roof Truss Support Columns [24 I Bearns, 61 cm (24 in.) x 30.5 cm (12 in.)]
Roof Truss Support Columns Material Stainless Steel Carbon Steel
GA-A14160
Fluid is introduced into each tank via a standpipe which distributes
the flow below a minimum fluid level to prevent erosion damage to structure.
Fluid is removed via a bellmouth locate^ 4 ft above the lowest point of
the conical tank floor.
6.4.2 HTS Piping System
This system comprises conventionally designed pipes and fittings simply mounted above grade. Pipes transporting the hot heat transfer salt are made of stainless steel to minimize corrosion to a level acceptable for the life of the plant. Pipes transporting cold fluid are made of plain carbon steel. Table 15 summarizes the piping specifications.
Conventional expansion loops will accommodate changes in pipe temper-
ature from ambient to operating temperature. The criteria used to select
supply and return pipe materials were 1) low cost, 2) state-of-the-art mate-
rials, with proven fabricability, 3) field assembly experience, 4) known
environmental compatibility, and 5) mechanical properties covered in appli-
cable design codes.
All pipes will be insulated externally to limit surface temperatures to 66°C (150°F). This minimizes heat loss and removes the personnel burn hazard.
The external pipe insulation was selected for its high strength, thermal stability, low cost, ease of application and repair, low conductivity, and proven environmental compatibility. Thermobestos is a standard external pipe insulation in high-temperature refinery applications. References that investigated compatibility between HTS and containment materials were reviewed. No data were found that indicated any concern over the maximum operating conditions anticipated for the heat transport system under study.
The environment outside the piping system will vary with site locations,
but typical refinery and/or seacoast atmospheres probably represent the most
91
TABLE 15 HEAT-TRANSPORT LOOP MATERIALS
Description Material Specification Comments
Supply Piping Austenitic stainless steel (Type 316)
ANSI: A-312 Welded pipe, 594°C (1100°F) operation
Return Piping Carbon steel (Type 1030)
ANSI: A-155 Grade 70
Welded pipe, 288°C (550°F) operation
External Insulation Thermobestos (hydrous calcium silicate/ asbestos fibers)
Federal HH-1-523 Sectional pieces, 15.2 cm (6 in.) thick for supply piping, 5.1 cm (2 in.) thick for return piping
GA-A14160
92
severe conditions. The selected steels and insulation materials have per-
formed very well at high temperature in the presence of these environments
for a few decades. Table 16 briefly summarizes known data and history used
for justifying the expectation of material/environmental compatibility for
the loop and storage system.
The system incorporates dual heaters and steam generators. These components may be isolated from the system by valves, blanking plates and removable sections of ducting (as explained in the "valves" section of this report). This permits component replacement or servicing while the system is operating with the parallel component.
The estimated piping quantities and fittings required for the fluid system outside of the component buildings for the 8-tank system are shown in Table 17.
6.4.3 Valves
Valves will be standard on/off gate valves (or equivalent) in the cold part of the system. In the hot part of the system valve design will be conventional, but materials will be selected for compatibility with hot heat transfer salt. Although the operating temperature of the hot valves is high, the pressure levels are very modest, and valve vendors contacted expect no difficulty in designing valves for the operating conditions of the system.
Adjacent to each valve used to isolate a component a spectacle blind blanking plate and flange, and a removable section of ducting will be incorporated. The valve and blanking plate provide a double block, and the removable section of duct interrupts the flow system. With these pro-visions, one component can be worked on while the parallel component is operating in the system.
Shutoff valves required in the ducting system are listed in Table 18.
93
TABLE 16 COMPATIBILITY OF MATERIALS WITH ANTICIPATED CHEMICAL ENVIRONMENTS
M a t e r i a l Component
J u s t i f i c a t i o n o f C o m p a t i b i l i t y
I n t e r n a l (HTS) E x t e r n a l
( R e f i n e r y o r S e a c o a s t Atms)
C a r b o n - s t e e l
316 S t a i n l e s s S t e e l
T h e r m o b e s t o s
Welded p i p e a t 343°C ( 6 5 0 ° F )
Welded p i p e a t 593°C ( 1 1 0 0 ° F )
E x t e r n a l i n s u l a t i o n a t 5 9 3 c C . ( 1 1 0 0 ° F ) max .
1 . V i r t u a l l y no c o r r o s i o n t o 454°C (850°F)
2 . S t a n d a r d c o n t a i n m e n t m a t e r i a l f o r i n d u s t r i a l h e a t t r e a t i n g t a n k s , h e a t e x c h a n g e r s
3 . No h i s t o r y o f n o n u n i f o r m c o r r o s i o n
1 . I n i t i a l u n i f o r m c o r r o s i o n a t t - 0 . 0 0 1 i n . / m o , d e c r e a s i n g a s p a s s i v e o x i d e d e v e l o p s ; 0 . 1 i n . t h i c k n e s s a l l o w a n c e f o r c o r r o s i o n
2 . J t a n d a r d c o n t a i n m e n t m a t e r i a l f o r i n d u s t r i a l a p p l i c a t i o n s b e t w e e n 454° and 593°C ( 8 5 0 ° a n d 1 1 0 0 ° F )
1 . T h e r m a l l y s t a b l e a t 649°C ( 1 2 0 0 ° F )
2 . N o n r e a c t i v e w i t h a i r and HTS
3 . S t a n d a r d g a s k e t and p a c k i n g m a t e r i a l i n HTS pumping s y s t e m s
Low g e n e r a l c o r r o s i o n a t 343°C ( 6 5 0 ° F ) < 0 . 0 0 2 i n . / l O y r
C o n d i t i o n s n o t c o n d u c i v e t o m a t e r i a l e m b r i t t l e m e n t , o r n o n u n i f o r m , a c c e l e r a t e d c o r r o s i o n
400° t o 1 1 0 0 ° F ; c o n d i t i o n s n o t c o n d u c i v e t o n o n u n i f o r m , a c c e l e r a t e d c o r r o s i o n ( i . e . , p i t t i n g , c r e v i c e , o r s t r e s s - c o r r o s i o n )
A s t a n d a r d h i g h - t e m p e r a t u r e p i p e m a t e r i a l f o r r e f i n e r y a p p l i c a t i o n s a t 593°C ( 1 1 0 0 ° F )
N o n r e a c t i v e
S t a n d a r d e x t e r n a l p i p e i n s u l a t i o n i n h i g h -t e m p e r a t u r e r e f i n e r y a p p l i c a t i o n s
GA-A14160
TABLE 17 PIPING AND FITTINGS
[8 Tank System, (4 Hot, 4 Cold)]
I cm
.D. (ft)
Length m (ft)
Minimum Wall Thickness mm (in.)
Expansion Loops R = 8 i.d. T Joints
90-deg Bends
Stainless Steel 110 (3.6) 41 (136) 5.6 (0.22) 3 2 2
55 (1.8) 122 (400) 2.8 (0.11) 1 2 2
64 (2.1) 119 (391) 3.3 (0.13) 2 2 5
33 (1.1) 122
404
(400)
(1327)
1.5 (0.06) 1 0 4
Carbon Steel 110 (3.6) 122 (401) 5.6 (0.22) 2 2 2
64 (2.1) 228 (750) 2.8 (0.13) 4 2 5
55 (1.8) 122 (400) 3.3 (0.11) 1 2 2
33 (1.1) 122
594
(400)
(1951)
1.5 (0.06) 2 0 4
GA-A14160
TABLE 18 SHUTOFF VALVES
Operating Inside Number Temperature Diam cm (ft) Required
593°C (1010°F) 110 (3.6) 1
76 (2.5) 2
64 (2.1) 1 43 (1.4) 2
288°C (550°F) 76 (2.5) 2
43 (1.4) 2
GA-A14160 6.4.4 Pumps
The HTS is pumped continuously from the cold s t o r e s tanks, through che heat exchangers and into the hot fluid storage tanks. A standard verti-cal drive, high-temperature pump will be used. A standby pump will provide redundancy and permit the replacement or servicing of one pump while the other is operating.
During the 8-hr heat delivery period a pair of pumps operating in
parallel will be used to pump HTS from the hot tanks through the dual steam
generators to the cold storage tank. These will also be vertical drive,
high-temperature pumps but made of materials compatible with HTS at 543°C
(1010°F). A standby pair of hot fluid pumps will be installed for redundancy
and to permit replacement or servicing of one pair while the other pair
is on line.
The principal pump performance requirements are listed in Table 19.
6.5 HTS PEAKING PLANT PLOT PLAN
A scale plot plan of the HTS peaking plant (Fig. 25) shows relative
locations of the major plant components buildings and piping runs. The
96
vo
L E G E N D
' CORE t iHTERMECriME -EAT E*CNA*G£R 2 P8iV»aT "CLIUU CfflCUL*W « AUMUFLT MEAT EXCHANGER S « t H i t - t * ? E*CWM.6£B
FC ^ E / I I E T W B£H£R>TOA
t S t C C A W R f -£L<UU CIRCULATOR
GA-A14160
Fig. 25. HTGR/HTS thermal storage peaking plant plot plan
TABLE 19 PUMP PERFORMANCE
Cold Pumps Each of 2 Hot Pumps
Flowrate, kg/sec (gpm) 2133 (20,000) 3197 (30,000) Head, m (ft) 39 (128) 28.6 (94)
Pumping Horsepower 1,093 1,208 Efficiency 0.75 0.75
Shaft Horsepower 1.457 1,611
Shaft MW 1.09 1.20 GA-A14160
nuclear heat source within its containment building is surrounded by control and service buildings, secondary helium heat exchanger sets, and the base load turbine generator building. Salt and helium storage tanks are nearby and adjacent to the peaking turbine generator building. Auxili-ary buildings and rail and roadways surround the plant.
Short runs oi secondary helium piping interconnect the in-vessel IHXs
and the adjacent helium/steam and helium/salt heat exchangers and circu-
lators. More lengthy, smaller diameter piping runs connect the salt heat
sources with the storage tank and the peaking turbine building. Inter-
mediate length steam lines run to and from the base load turbine building.
The only remote system component, the cold HTS storage tanks, requires
low-temperature, small diameter piping.
The various auxiliary buildings that are associated with the HTGR/HTS
peaking plant are described briefly below.
6.5.1 Reactor Service Building
The Safety Class 2 and Seismic Category I reactor service building houses new and used fuel storage facilities and reactor auxiliary systems not located in the RCB. The primary activity in the reactor service build-ing will be refueling.
98
The fuel storage concrete monolith is sized to contain 28% of the spent fuel elements of one reactor. This massive structure comprises the inner core of the reactor service building and is also used to store radio-active tools, control rods and other equipment. The remaining portion of the reactor service building below the refueling floor houses fuel inspection and sealing facility, equipment service facility, fuel shipping facility, shipping cask transporter parking and service area, and other auxiliary systems. A crane atop the reactor service building makes all lifts neces-sary in storing fuel in the monolith and transferring fuel to the reactor.
6.5.2 Control Building
This is a Safety Class 2 and Seismic Category I structure housing
the control room, auxiliary electrical equipment, ventilation equipment,
controlled access area, and the reactor plant cooling water system.
The following functional areas comprise the control building:
1. Control and Computer Rooms.
2. Essential Switchgear Room.
3. Relay and Instrument Room.
4. Battery Charging Rooms.
5. Ventilation Equipment.
6. Reactor Plant Cooling Water System Equipment.
6.5.3 Penetration Building
The Safety Class 2 and Seismic Category I penetration building pro-
tects both safety and nonsafety related penetrations.
The building consists of a tunnel housing electrical control and power cables. It is laid out so that it completely surrounds the RCB, thereby protecting containment penetrations at different locations around the perimeter of the containment structure.
99
6.5.4 CACS Power Supply Building
The Safety Class 2 and Seismic Category I auxiliary circulator power supply building houses the two auxiliary circulator power supplies,
6.5.5 Diesel Generator Building
The Safety Class and Seismic Category I diesel generator building
houses the standby power supplies.
6.5.6 Turbine Buildings
The turbine buildings are nonsafety class nonseismi^ category struc-
tures containing the turbine generators and other equipment related to the
conventional portion of the base load and peaking plants.
No specific design requirements are imposed on the turbine conventional design; therefore, it is based on the same criteria as would be used for a conventional fossil-fired steam plant turbine building.
6.5.7 Helium Storage Building
This nonsafety class nonseismic category structure houses the primary coolant helium storage bottles.
100
7. PLANT OPERATION
The HTGR/HTS peaking plant has two modes of operation, continuous base load and cyclic peak load operation. The reactor system portion of the plant follows the operation pattern of the standard 2000 MW(t) steam cycle plant. The reactor core operates continuously at full power supplying thermal energy to both the base load steam plant and the HTS thermal storage loop.
The primary helium system, intermediate helium loop, base load, and thermal storage peaking load systems are divided into two completely isolated loops. This allows one-half of the total plant to operate while the other portion is down for either scheduled or unscheduled maintenance. In addition, the individual heat exchangers (He/HTS and He/Steam) within the same loop can be taken out of service while allowing the other to function normally. This allows full or partial base load operation with full or partial peaking plant operation. The operation of only the base load system can be accommo-dated; however, the peaking system cannot be operated unless the base load plant is functioning. As a result, the primary and intermediate helium circulator drive turbines are supplied by steam from the base load system.
The steam plant portions of the HTGR/HTS plant are similar in oper-ation to conventional base load steam and peaking load steam plants. Only those aspects of operation that differ significantly for conventional plants will be discussed below.
The peaking plant has the capability of accommodating a sudden load
rejection and turbine trip without rejecting large quantities of heat.. This
is accomplished by simply throttling back or completely stopping the flow
101
of hot salt to the peaking plant steam generator. This differs signifi-
cantly from fossil-fired power-conversion systems, which are limited by
the shutdown time of the burners.
Additional differences in the operational sequence of the HTGR/HTS system lies in the initial startup of the molten salt loop. The eutectic mixture of KN0 3/NaN0 2/NaN0 5 salt melts at 142°C (288°F), but a proprietary salt dilution technique developed by the American Hydrotherm Corporation (Ref.,25) allows controlled reduction of the salts melting point down to ambient.
The plant startup sequence would be to bring the base load system on-line first. Then, after all the reactor plant systems and base load plant systems are in operation, the salt storage loop startup can be initiated. For the initial startup the salt tank is charged with water and HTS of the solution strength desired. As the first step the circulating pump is turned on and the solution is circulated through the helium/salt heat exchangers to the hot reservoir and through a piping bypass back to the cold reservoir. The system temperature can be raised uniformly by steam coils located in the hot reservoir and supplied by steam from the base load plant.
Heat-up is normally accomplished in two steps. All of the water in
the solution can be boiled off simply by uhe heat available in the tank
steam coils. The limiting temperature reached with the steam pressure
available is selected above the eutectic melting point of the salt.
Through its lower range of between ambient and 175°C (350°F), the solution is a boiling mixture until ail the water is evaporated and the salt reaches the anhydrous molten state. It must be remembered, however, that all boiling takes place in the salt tank, which is vented to the atmosphere through the steam/salt separator. Consequently, the only pres-sure in the system is the circulating pump discharge pressure required to overcome static head and pressure drops through the lines and process equipment.
102
The second step of salt heat-up commences after all the water has been removed from the salt. Heat addition to the salt through the helium/salt heat exchanger is initiated and the steam flow to the steam coils is dis-continued. Hot helium from the intermediate loop heats the salt to its normal operating storage temperature. This continues until all the salt inventory has been heated to its hot storage temperature.
During this thermal storage mode of operation of the HTS loop, the components of the peaking plant can be thermally conditioned before its startup sequence is initiated. The normal daily startup and shutdown of the peaking steam system would be similar to conventional fossil-fired peaking plants. During the nonpeaking hours, the peaking plant would be thermally conditioned and in a standby mode. Normal operation of the salt loop should not result in any salt freeze-up problems. The lowest oper-ating temperature encountered is approximately 288°C (550°F), which is well above its eutectic temperature.
When shutting down the entire peaking plant for scheduled maintenance, water can be introduced into the circulating molten salt at the temperature well above the freezing point of anhydrous HTS. Part of the water vaporizes but some is absorbed. The rate of water addition is predetermined and as dilution continues, maintains the salt in solution down to ambient temper-ature. At no time is there any danger of solidification, and the cooling rate is designed to avoid uneven thermal stresses. The salt/water solution can then be stored indefinitely in the hot storage reservoirs until the next scheduled startup.
103
8. PARAMETRIC STUDIES
8.1 ALTERNATIVE REACTOR SIZES
The effect on the system design parameters was investigated for two additional reactor plant sizes utilizing the identical salt storage loop and peaking turbine facility as defined in the reference 2000 MW(t) HTGR/HTS design. System design studies for an 850 MW(t) reactor system size solely dedicated to producing peak power and a 3000 MW(t) reactor system size with additional base power generation were carried out for the same 8-hr duty cycle. The same basic groundrules that were adopted for the 2000 MW(t) size were used in selecting the cycle parameters for the alternative sizes.
Figure 26 illustrates the primary and secondary helium cycle parameters
for the 3000 MW(t) reactor system. The thermal power split is 850 MW(t)
for thermal storage and 2260 MW(t) for steam generation in the base load
plant. Figure 27 gives the steam cycle parameters for the base load portion
of the plant. The net base plant electrical output is 760 MW(e). The
peaking portion electrical output is 1010 MW(e) during an 8-hr duty cycle.
The design parameters for an 850 MW(t) reactor system are shown in
Fig. 28. The plant would differ slightly from the larger size plants in the
following areas: the Fort St. Vrain size reactor has no separate CACS and
special provision would be required to supply steam continuously to the
helium circulator turbine drive (an alternative is electric motor drives).
8.2 VARIATION OF DUTY CYCLE
Early in the study, selection was made of an 8-hr daily duty cycle
on which the reference plant study would be based. However, each utility
104
PRIMARY HELIUM SYSTEM I N T E R M E D I A T E LOOP HELIUM SYSTEM
355.9°C (672.6°F)
5 MPa (725 PSIA)
i 76.5 MW(t) 3!7°C (602.9°F|
3000 MW(I)
760°C (1400°F)
O cn
4.8 MPa (702 3 PSIA)
5.1 MW(t) IHX HEAT LOSS
HEAT LOSS TO CAHE 10.2 MW(t)
366.2° C (691°F) \
315.9°C (600.5° FI 5 MPa 025 PSIA)
67.9 MWItl
• 3057.3 MW(t) >
4.9 MPa (713PSIAI
4.9 MPa (712 PSIA)
704.4°C (1300°F)
9 6 MW(t) RECUPERATED FROM PIPING HEAT LOSS
PIPING HEAT LOSS 13 6 MW(t)
1429 KG/SEC (11.33 * 10G LB/HR)
758.3°C (1397°F)
4.9 MPa (712 PSIA)
1520 KG/SEC (12 05 * 106 LB/HRI
HOT SALT PIPING LOSS 6.4 imit)
307.2°C (58E°F)
3116.4 MW(t) —
/ t 850 MW(t}
856.4 MW(t) THERMAL STORAGE
2260 MW(tl BASE LOAD PLANT
4.89 MPa (710 PSIA)
702.8DC I1297°F)
PRIMARY LOOP HEAT BALANCE CORE POWER 3000 MW(t) HEAT ADDED BY CIRCULATORS 76.5 MW(l) HEAT LOSS FROM CORE TO PCRV 3.9 MW(l) HEAT LOSS TO CAHE 10.2 MW(t) HEAT LOSS FROM IHX TO PCRV 5.1 MW(t) NET HEAT TRANSFER FROM IHX 3057.3 KWlt)
INTERMEDIATE LOOP HEAT BALANCE IHX HEAT LOAD 3057.3 MW(i) HEAT ADDED BY CIRCULATORS 67.9 MW(t) HEAT LOSS FROM PIPING 4 0 MlV(t) HEAT LOSS FROM HX's 4.8 MW(t) NET HEAT TRANSFER 3116.4 MtY(i)
GA-A14160
Fig. 26. Primary and secondary helium cycle parameters for a 3000 MW(t) reactor system
1518.1H
O cr>
HELIUM
1297 T 71a?
SH EVAP ECON
I2513P J955T
1210 MW(t)
586 P 1002 F
REHEAT
29S5P 369=1:
4.67 X 106 W
585T 7MP
HELIUM
1308.4H 643P 634*F
3.20 X 10s W
LEGEND H - ENTHALPY (BTU/LB) W = LB/HR P - PSIA
346.4H
1427.3H
HEAT BALANCE STEAM GEN. THERMAL POWER 2260 MWItl HE CIR. POWER 144.4 MWIt) NSS THERMAL POWER 2115.6 MW(t) GROSS PLANT OUTPUT 773.6 M W I e l CIR WATER PUMP POWER 11.2 M W I e l OTHER STATION AUX. 2.4 M W I e l NET PLANT OUTPUT 760 0 MWIe)
PRIMARY HE CIRC. TURBINE DRIVE 44.4MW
INT. LOOP HE CIRC DRIVE 45 M W
GA-A14160
Fig. 27. Base load plant with 3000 MW(t) reactor system (English engineering units) (Sheet 1 of 2)
3531.1H
O -J
HEUUM
1702.8X U.9P
SH EVAP ECON
HEAT BALANCE STEAM GEN. THERMAL POWER 2260 M'n{t) HE C1R. POWER 144.4 MW(t) NSS THERMAL POWER 2115.5 MWIt l GROSS PLANT OUTPUT 773.6 M W I e ) CIR WATER PUMP POWER 11.2 M W I e ) OTHER STATION AUX. 2.4 MWIe) NET PLANT OUTPUT 760.0 MWIel
GA-A14160
Fig. 27. Base load plant with 3000 MW(t) reactor system (SI units) (Sheet 2 of 2)
352.7°C (666.8°F)
18.9 MW(t) SHAFT WORK
315.8°C | '600.5°F)
5 MPa (725 PSIA)
760°C (1400°F)
1.5 MW(t) CORE HEAT LOSS
343.6°C (650.5°F)
1HX HEAT LOSS 1.4 MW(t)
4.8 HPa (704.3 PSIA)
317°C (602.7°F)
866 MW(t) :
V 5 MPa (725 PSIA)
19.1 MW(t) SHAFT WORK
HOT SALT PIPING & TANK HEAT LOSS S.4 MW(t)
4.9'MPa (713 PSIA)
760°C 11400° F)
4.9 MPa (712 PSIA)
704.4°C (1300°F)
401 KG/SEC (3.18 x 1 0 6 LB/HR)
2.7 MW(t) RECUPERATED FROM PIPING HEAT LOSS
HX HEAT LOSS 1.4 MW(l)
4.8 HPa (704 PSIA)
307.4°C (58o°F)
4.9 MPa (712 PSIA) 1
3.9 MWU) PIPING HEAT LOSS
4.9 MPa (711 PSIA)
882.5 MW(t) —
• 856.4 MW(t) 85C MW(t) T H E R M A L STORAGE
702.8°C (1297°F
26.1 MW(t) , A U X I L I A R Y
STEAM BOILER
426.4 KG/SEC (3.38 x 10 6 LB/HR)
PRIMARY HEL IUM HEAT BALANCE
CORE POWER 850 MW(t) HEAT ADDED BY CIRCULATOR 18.9 MW(t) HEAT LOSS FROM CORE 1.5 MW(t) HEAT LOSS FROM IHX 1.4 MW(t) NET HEAT TRANSFER BY IHX 866. MWIt)
INTERMEDIATE LOOP HEAT BALANCE
IHX NET HEAT TRANSFER 866. MW(t) HEAT ADDED BY CIRCULATOR 19.1 MWW HEAT LOSS FROM PIPING 1.2 MW(t) HEAT LOSS FROM HX 1.4 MW(t) NET HEAT TRANSFER 882.5 MW(t)
GA-A14160
Fig. 28. Primary and secondary helium cycle parameters for an 850 Mtf(t) reactor system
has its own particular system load demand characteristics. The Thermal Energy Storage Peaking Plant can be tailored to fit the needs of the different utilities. The effect of varying the duration of the daily duty cycle or changing it to a weekly duty cycle are important consider-ations in the evaluation of the system.
Figure 29 illustrates the peaking power available for variations in da .y duty cycle. The heat input to the thermal storage loop is held fixed at 350 MW(t). The peaking power is approximately 4040 MW(e) for a 2-hr duty cycle and 653 MW(e) for a 12-hr duty cycle,. This curve remains identi-cal for all three different reactor sizes, because all the plants utilize the same thermal storage and peaking plant system. The base load power output portions for the 3000, 2000, and 850 MW(t) are 760, 395, and zero MW(e), respectively.
Fixing the peaking power at the reference plant value [1009 MW(e)] and varying the duty cycle results in a change in the base load power output. Figure 30 illustrates this effect for the 2000 MW(t) and 3000 MW(t) plants.
The effect of changing the duty cycle on the salt inventory has a direct impact on the storage tank cost, salt inventory cost, and salt replacement cost. Figure 31 shows how the salt inventory is affected by duty cycle variations for both daily and weekly cycles. In all instances the thermal power devoted to thermal storage is 850 MW(t). The 5-day cycle results in a significant increase in salt inventory. Approximately 2.45 x 8 6
10 kb (540 x 10 lbm) of additional salt inventory is required by changing from a 7-day cycle to a 5-day cycle.
109
D A I L Y D U T Y C Y C L E (HR) GA-A14160
Fig. 29. The peaking plant power available for variations in the daily duty cycle
110
1000
800
600 cc UJ 5 o . a . a < o
CO. < oo
400
200
3000 MW(t) PLANT
2000 MW(t) PLANT
CONSTANT PEAKING PLANT ELECTRICAL OUTPUT OF 1009 HW(e)
_L 6 8
DAILY DUTY CYCLE (HR) 10 12 14
GA-A14160
Fig. 30. Base load power variation with daily duty cycle with a constant peaking plant output of 1009 MW(e)
1000
800
600
400 h
200
-A 400
•5 DAY CYCLE
•7 DAY CYCLE
X _L
300
200
100
6 8 DUTY CYCLE (HR)
10 12 GA-A14160
Fig. 31. The relationship between the salt inventory and variations of the duty cycle
9. ECONOMICS
9.1 REFERENCE PLANT COST
Estimates of the cost of the HTGR/HTS systems were developed for both the nuclear steam supply (NSS) and balance of plant (BOP) system for the reference 2000 MW(t) design. All cost data is escalated to July 1, 1976 dollars.
The cost breakdown for the reference plant [8-hr daily duty cycle, 395 MW(e) base load, and 1010 MW(e) peaking load] Is summarized in Table 20. The cost breakdown of the NSS and BOP among the various portions of the plant are also shown in the table. The cost for the intermediate loop portions includes piping, intermediate helium circulators, support systems, and the associated structures, instrumentation and controls. The base plant portion costs includes the helium/steam heat exchangers, which are part of the NSS, and all the BOP systems associated with the base load steam plant equipment. The salt storage and peaking plant costs includes the helium/salt and salt/steam heat exchangers in the NSS portion and the remaining equipment in the BOP portion.
The information to determine the cost of electricity (COE) is given in Table 21. The assumptions and methods used in estimating the power cost are summarized in Table 22. The COE can be determined by two different methods; one called the base load credit method and the other based on actual plant cost method, both are summarized in Table 21. The base load credit method is highly dependent on what value credited the base power. In the present study a base credit value of 22.3 mills/kW(e)-hr was selected because it represented the COE of a purely base loaded 2000 MW(t) steam HTGR calculated on a consistent basis. The COE for the base load credit method results in a peaking power cost of 51 mills/kW(e)-hr.
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TABLE 20 COST BREAKDOWN FOR THE 2000 Mtf(t) PEAKING PLANT STUDY
(All Costs in Millions of July 1976 Dollars)
Nuclear Intermediate Salt Storage and Task Plant Loop Base Plant Peaking Plant Total Plant
NSS(a)
22 Nuclear Steam Supply 118.7 29.6 11.8 63.1 223.2
B0P(b)
21 Structures 77.7 0.2 15.4 48.2 141.5 22 Reactor Plant Equipment 41.0 5.1 1.2 4.5 51.8 23 Turbine Plant Equipment 0 0 94.1 127.6 221.7 24 Electric Plant 36.4 0 13.9 35.9 86.2 25 Misc. Plant Equipment 18.0 0 0 4.2 22.2 26 Salt Storage Tanks 0 0 0 83.8 83.8
Total BOP 173.1 5.3 124.6 304.2 607.2 Salt Inventory 0 0 0 36.0 36.0
Total Plant Cost 291.8 34.9 136.4 403.3 866.4
(a) Costs include Direct Cost, Indirect Costs, Contractual Risk, GA Allocation, and Cost Margins.
^Costs include Direct Cost, Indirect Costs, and allowance for indeterminates.
GA-A14160
TABLE 21 REFERENCE 2000 Mtf(t) PEAKING PLANT
COST OF ELECTRICITY
Method of Calculation
Item Base Load Credit Method
Actual Plant Peaking
Cost Method Base Load
Plant Capital Cost Interest During Construction
(At 30.262 NSS and Base Plant) (At 18.91% Salt Storage and Peaking Plant)
866.4
140.1 76.2
542.2
42.0 76.2
324.2
98.1
Total Cost at Completion 1082.7 660.4 422.3
Annual Capital Charge (At 15%)
Operating and Maintenance Annual Fuel Cost, 57.64/106 Btu Annual Salt Replacement Base Load Credit (At 22.3 mills/kW-hr) Annual Operating Cost
162.4 3.3 29.3 7.6
(65.8) 136.8
99.0 1.4 12.5 7.6 0
120.5
63.4 1.9 16.8 0 0 82.1
Peakine Power Cost •• Annual Operating Cost mills/kW(e)-hr (Availability)(Plant Power)(Operating Time) 51.5 45.4 27.8
GA-A1416Q
TABLE 22 ASSUMPTIONS FOR ESTIMATING POWER COSTS
Fixed Capital Rate
Operation and Maintenance
Fuel Cycle Costs
Salt Replacement Costs
Base Load Credit
Peak Power Cost
Assumed at 15%/yr on total capital costs.
Assumed to be constant at $3.3 million/yr, same as the standard 2000 MW(t) HTGR.
Based on recent uranium costs giving fuel cycle cost of about 57.6<?/l06 Btu.
Salt loss based on measured N£ evolution due to thermal decomposition of NaN02-
Determined on the basis of 22.3 mills/kW-hr at a 0.85 availability of base load plant.
Total operating costs minus credit for base load divided by peak power produced. It is assumed that for the specifind operating interval, the availability factor is 0.9 to account for maintenance opportunity during off-peak time.
GA-A14160
1 1 6
Figure 32 illustrates how the results of the peak power cost is affected by changing the base power credit.
The actual plant cost method involves separating that portion of the total plant cost that is associated with the peak power generation and cal-culating the COE. This method results in a COE for the peaking power of 45 mills/kW(e)-hr and 28 mills/kW(e)~hr for the base load power cost.
9.2 ALTERNATE PLANT COSTS FOR 850 MW(t) AND 3000 MW(t) PLANTS
Plant costs for two additional reactor sizes 850 MW(t) and 3000 MW(t) coupled to the same salt storage and peaking steam plant system were developed. The cost associated with the salt storage and peaking plant remained constant, however, the NSS and BOP costs associated with the nuclear plant, intermediate loop, and base load portion were adjusted by scaling factors developed from various steam plant HTGR reactor sizes. Table 23 summarizes the results of the various plant costs and cost of electricity [base load credit = 22.3 mills/kW(e)-hr]. The resulting peak power costs are 54.7 mills/kW(e)-hr and 47.9 mills/kW(e)-hr for the 850 MW(t) and 3000 MW(t) reactor sizes, respectively.
The 850 MW(t) plant generates only peaking power, therefore a cost penalty was associated with the circulator drive power requirement at the rate of the base load power credit [22.3 mills/kW(e)-hr].
The results of the comparison made in Table 23 indicates a relative plant comparison at a base load credit of 22.3 mills/kW(e)-hr. Figure 32 illustrates the peaking plants COE as a function of base power credit.
9.3 COST COMPARISONS
Comparison of the capital costs of the energy storage systems in re-cent studies (Ref. 26) has been described as a sum of two terms. The costs C (in $/kW) for the power-related equipment are associated with
1 1 7
Fig, 32. Peaking power cost for variation of the base load power credit
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TABLE 23 COST COMPARISON OF 850 MU(t), 2000 MW(t), AMD 3000 MW(t) PKAKINC PLANTS
(Costs in Millions of July 197b Dollars)
2000 MVJ(t) 3000 MU'(t) S50 MK(t) 395 MU'(e) liase 76C MK(e) Base
Item 1010 MW(e) Peaking 1010 >IU(e) Peaking 1010 MW(e) Peaking
NSS Nuclear Plant 71.0 113.7 151.4 Intermediate Loop 17.7 29.6 37.8 Base Plant 0 11.8 19.5 Salt Storage and Peaking Plant 63.1 63.1 63.1
BOP Nuclear Plant 103.6 173.1 220.3 Intermediate Loop 3.2 5.3 6.8 Base Plant 0 124.6 205.5 Salt Storage and Peaking Plant 304.2 304.2 304.2
Salt Inventory • 36.0 36.0 36.0 Plant Capital Cost 598.8 866.4 1045.1
Interest During Construction At 30.26% 140.1 194.2 At 18.91% 181.2 76.2 76.2
Total Cost at Completion 780.0 1082.7 1315.5 Annual Capital Charge (At 15%) 117.0 162.4 197.3 Operating and Maintenance 1.4 3.3 5.0 Fuel Cost (57.6C/106 Btu) 12.5 29.3 44.0 Annual Salt Replacement 7.6 7.6 <7.6> Base Load Credit at 22.3 mills/kW-hr 6.7 (65.8) 126.8 Annual Operating Cost 145.2 136.8 127.1 Peaking Power Cost, mills/kW-hr 54.7 51.5 47.9
GA-A14160
the power output (kW) of the device. The storage costs C are determined s by the specific costs (in $/kW-hr) of energy storage capacity and the time T (in hours) for which the energy storage system c.an deliver energy at rated output power. Analyzed in this fashion, the per-unit capital cost C (in $/kW) is:
The estimates of the terms C , C , and T are $544/kW, 15, and 8, respec-P s
tively, for the systems considered in this study. These costs include both equipment and installation costs and normal contingencies and over-heads, but do not include an allowance for interest during construction. Table 24 gives the breakdown of the calculations of C and C . p s
Comparisons with gas turbine plant peaking power costs were made as a function of the duty cycle. Scaling equations were developed to determine the influence of the duty cycle on the cost of electricity for the 2000 MW(t) HTGR/HTS peaking plant. The results are shown in Fig. 33. The basis for the HTGR/HTS curves is for a fixed thermal rating of 2000 MW(t) of which 850 MW(t) is utilized by the thermal storage loop, allowing the peaking capacity to increase as the operating period decreases. Two cost curves are shown for the peaking power; one is based on a base load credit of 22.3 mills/kW(e)-hr and the other on 30 mills/kW(e)-hr. A range of fossil fuel costs are illustrated in Fig. 33, with the lower value repre-senting July 1976 prices.
The power costs for gas turbines are a weak function of operating period, but are very sensitive to fuel prices. The HTS system appears to be an attractive alternative to gas turbines, even at the lower range of fuel prices.
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TABLE 24 CAPITAL COST OP ENERGY STORAGE SYSTEM
Item Capital Cost in $ Million
Storage System Coat Storage Tanks Salt Inventory Misc. Plant Equipment (pumps, piping)
Total Storage System Cost
$124 x 106 Js 1010 MW(e) x 8 hr
$ 83.8 36.0 4.2
$124.0
$15/kW-hr
Power-Related Equipment Cost Capital Cost of HTS/HTGR $866.4 Time-Averaged Power Output 732 MW(e) Capital Cost per kW(e) $1184/kW(e)
Capital Cost per kW(e) for a 2000 MW(t) HTGR $640/kW(e)
Differential Cost of Power From a Base 2000 MW(t) HTGR, C $544/kW(e)
GA-A14160
1 2 1
140
120
3 JC c/5
CA o o
3 o a. C9
<
100
80
60
40
20 1
GAS TURBINE 30 S/BBL DISTILLATE 20 $/BBL DISTILLATE
HTGR-HTS SYSTEM
BASE LOAD CREDIT 22.3 MILLS/kW(e)-HR 30.0 MILLS/kW(e)-HR
_L J_ 4 6 8 10 12 14
HR/DAY OF PEAKING POWER OPERATION GA-A14160
Fig. 33. Power cost comparison, HTGR-HTS and gas turbine
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10. CONCLUSIONS AND RECOMMENDATIONS
The HTS thermal storage concept makes highly efficient use of energy generated during off-peak periods to provide electrical generation for daily Intermediate and peaking power demands, The coupling of this con-cept with an HTGR nuclear plant shows peaking power costs less than for fossil-fueled systems for this type of application. In addition to potential cost savings, the HTGR/HTS concept provides an important means of conserving fossil fuels by substituting nuclear energy.
The system utilizes existing HTGR technology as incorporated into the Fort St. Vrain plant and the design of the large HTGR for commercial appli-cation. The salt has been used on a smaller scale in commercial application under similar conditions for many years.
The economic attractiveness results from the high efficiency, low fuel costs and the economy of large size. Except for very large utilities with geographically concentrated loads, the attractiveness of the concept may depend on the utility requiring additional base load as well as peaking power. The plant capacity can be adapted to a utility's peaking power requirements by selecting an appropriate base load/peaking load power split.
Typically, U.S. utilities have peaking and intermediate power demand which are 40 to 50% of the peak load level. Based< on FEA projections (Ref. 27) for a "moderate/low" growth in electrical demand, this corresponds to 700,000 to 875,000 MW(e) for peaking/intermediate power by the year 2000. Considering that the HTGR/HTS concept appears an attractive substitute for fossil peaking systems, and that its attractiveness will grow as utilities approach full nuclear base load capacity, a sizeable market for the concept could develop.
123
The purpose of this study was to establish a reference plant design in which to serve as the basis for evaluating the technical and economic aspects of the concept. Optimization of the system's performance and cost were not within the scope of this study. Still, the results of selected plant designs do appear competitive with fossil-fueled alternatives.
There are uncertainties in the plant capital costs, particularly those associated with the salt loop and storage system, because of a lack of design definition. However, it is not anticipated that any new tech-nology is required and the costs are based on existing materials and sim-ilar types of components. Conceptual design studies carried out on the plant components show that no technical barriers exist.
The study revealed a number of areas of potential improvements for the HTGR/HTS peaking plant concept. Significant cost savings are avail-able by optimizing the design conditions of the salt/steam generator. This can be accomplished by increasing the salt storage temperatures, in-creasing the intermediate helium temperatures, and increasing the primary helium temperatures.
Improvements in the design of the storage tanks can also be made, such as utilizing a thermocline reservoir or using an internal stainless steel thermal liner with an external carbon steel structure. These im-provements could significantly reduce the cost of the storage tanks.
A major uncertainty in the annual operating costs, and, therefore, directly impacting the peak power cost is the cost of replacing the salt. The available data on the decomposition of the NalK^ component is incon-clusive. It is expected that the replacement cost estimates used are conservative and could be significantly reduced. Other salt mixtures could be considered in further studies.
An investigation of how this type of power peaking plant fits in with a typical utility grid may show other inherent advantages. Concen-tration of a large portion of a utility's peaking power In the HTGR/HTS
124
power peaking system has the net effect of increasing the utility's over-all capacity. Normal utility operation requires turndown of even the large base loaded nuclear and fossil-fired plants during certain periods of operation. This causes a reduction in the overall capacity. If future electrical growth potential is met with HTGR/HTS peaking plants, the mix of power generating plants would change, allowing the base load plants to be operated continuously at full power and the large power swings to be accommodated by the HTGR/HTS peaking system. The economic advantages of such a system in a utility's grid needs to be fully investigated.
The efficiency of the thermal energy utilization is approximately 90% of that for direct use of the energy;(without storage) in a modern steam electric power plant. The thermal storage system developed for this study may also be coupled with other types of high-temperature heat sources, in-cluding LMFBR, GCFR, and solar applications.
125
REFERENCES
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11. Fernandes, R. A., 0. D. Gildersleeve, and T. R. S<*.l>.'"..'/.rfr, "Assessment of Advance Concepts In Energy Storage and Theix Application on Electric Utility Systems," Paper No. 6, 1-17, Proceedings of the World Energy Conference, Detroit, MI, 1974.
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18. Silverman, M. D., and J. R. Engel, "Survey of Technology for Storage of Thermal in Heat Transfer Salt," Oak Ridge National Laboratory Report ORNL-TM-5682 (January 1977).
19. DuPont Hitec Heat Transfer Salt, E. I. duPont de Nemours and Co., Explosives Dept., Chem. Products Saves Div., Wilmington, Del., 19898.
20. Bohlmann, E. C., "Heat Transfer Salt for High Temperature Steam Generation, Oak Ridge National Laboratory Report 0RNL-TM-3777 (Decem-ber 1972).
127
21. Brownlie, D., "Inorganic Heat Transfer Liquids," The Steam Engineer 115, (1941).
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24. Hoffman, H. W., and S. I. Cohen, "Fused Salt Heat Transfer - Part III: Forced-Convection Heat Transfer in Circular Tubes Containing the Salt Mixture NaNO2 • NaNC>3 • KN03," Oak Ridge National Laboratory Report 0RNL-2433.
25. Hydrotherm Molten Salt and Transfer Systems Corp., 470 Park Avenue, South, New York.
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