Cost and performance design approach for GTHTR300 power conversion system

Nuclear Engineering and Design 226 (2003) 351–373

Cost and performance design approach for GTHTR300 power conversion system Xing Yan a,∗ , Takakazu Takizuka a , Shoji Takada a , Kazuhiko Kunitomi a , Isao Minatsuki b , Yorikata Mizokami b a

Japan Atomic Energy Research Institute, Oarai-Machi, Ibaraki-ken 311-1394, Japan b Mitsubishi Heavy Industries, Ltd., Japan

Received 16 December 2002; received in revised form 12 June 2003; accepted 16 July 2003

Abstract Japan Atomic Energy Research Institute (JAERI) has been carrying out a design and developmental program for the gas turbine high temperature reactor of 300 MWe nominal-capacity (GTHTR300) power plant, aiming at prototype demonstration in Japan during 2010s. This paper introduces overall objectives of the program and describes the plant design and development approach taken to achieve these goals. A detailed description is focused on the power conversion system design and associated component research and development undertaken in the present program. The power conversion system incorporates unique design approach of non-intercooled cycle to attain economical performance at minimal system complexity, intrinsic cycle flow provision for reactor pressure vessel cooling, simplified and high-performance turbomachine in a horizontal design, and modularity of maintenance for all major power conversion equipment. This paper reports extensive technical evaluation related to these significant system design features, which are shown to offer the optimum solution of plant cost, efficiency potential, reliability and maintainability in addition to near-term commercial deployment. © 2003 Elsevier B.V. All rights reserved.

1. Introduction Design studies for gas turbine high temperature reactor conducted in JAERI over a period of several years had evaluated a range of reactor core and power conversion system options before the reference design of a nominal 300 MWe gas turbine high temperature reactor system (GTHTR300) was determined. Working with major domestic industries, JAERI has been carrying on the GTHTR300 detailed design evaluation and necessary component development and tests to validate the design. The JAERI program, which is en∗ Corresponding author. Tel.: +81-29-264-8897; fax: +81-29-264-8608. E-mail address: [email protected] (X. Yan).

trusted by the Ministry of Education, Culture, Sports, Science and Technology of Japan until 2007, aims at completing the basic design and development to support the system’s prototype demonstration in the next decade. For the near-term development of a successful system, the GTHTR300 takes on a design approach of SECO, i.e. simplicity, economical competitiveness and originality. It calls for incorporation of greatly simplified design solutions in order to minimize technological requirement or development risk and to yield the kind of system simplicity characteristic of least component count and effective modular system arrangement to facilitate construction, operation and maintenance. The approach to economical performance combines the system simplicity with

0029-5493/$ – see front matter © 2003 Elsevier B.V. All rights reserved. doi:10.1016/S0029-5493(03)00212-7

352

X. Yan et al. / Nuclear Engineering and Design 226 (2003) 351–373

maximized thermal power of reactor, commensurate with inherent and passive reactor safety, and high efficiency potential for power generation. The cost of electricity targets at
4/kWh (∼US3.5 /c/kWh) in order to be competitive economically in current and future Japanese power generation markets. The quest for the unique set of system and economical goals mandates an original design effort. The GTHTR300 primary system consists of three subsystem modules including a reactor module, a gas turbine generator (GTG) module, and a heat exchangers (HTX) module (Fig. 1). The prismatic reactor is of annular core whose thermal power is maximized within the inherent and passive reactor safety requirement and in the bulk limit of reactor pressure vessel construction. The fuel design improves from the pin-in-block element of the HTTR test reactor that JAERI has developed and is now operating successfully. The fuel cycle is of high (120 GWd/t) burnup and low (<1.4) power peaking factor in an extended (2 years) refueling interval.

The GTHTR300 baseline design and performance data are summarized (Table 1). Yan et al. (2003) provided a detailed description of the GTHTR300 system design and development. Katanishi and Kunitomi (2002) reported on the plant’s safety design and Nakata et al. (2002) on nuclear design. The present paper discusses the power conversion system with emphasis on the system’s unique cost and performance design approach.

2. Features of GTHTR300 power conversion system Non-intercooled power conversion cycle with intrinsic cooling provision for reactor pressure vessel (RPV), horizontal high-performance gas turbine generator, and separated arrangement of the turbomachine from heat exchangers represent the GTHTR300 power conversion system’s original design approach to economical performance and early deployment (Yan et al.,

REACTOR

HTX module

Control valves

Annular block core

GTG module Recuperator

Precooler Turbine

Compressor

Generator

GTHTR 300 rev. 080499

Fig. 1. GTHTR300 plant system arrangement.

X. Yan et al. / Nuclear Engineering and Design 226 (2003) 351–373 Table 1 GTHTR300 baseline design and performance data Overall plant Reactor power Reactor pressure vessel Reactor safety system Radioactive nuclide retention Plant cycle Power generation Net power output Net generating efficiency Plant capacity factor Reactor Coolant inlet/outlet temperature Coolant inlet pressure Core coolant pressure loss Average power density Fuel element Fuel cycle Enrichment Average burnup Shutdown refueling Refueling duration Turbomachine Shaft design type Shaft speed Turbine inlet pressure Turbine mass flow Turbine expansion ratio Number of turbine stages Turbine polytropic efficiency Compressor inlet temperature Compressor pressure ratio Number of compressor stages Compressor polytropic efficiency Generator drive Generator type Generator cooling Generator efficiency Heat exchangers Recuperator design type Recuperator thermal rating Recuperator effectiveness Recuperator total pressure loss Recuperator construction material Precooler design type Precooler thermal rating Cooling water inlet temperature Precooler design LMTD Precooler tubing material

600 MWt/unit (4 units/plant) SA533/508 (Mn–Mo) steels No active emergency system Confinement Non-intercooled Brayton cycle 280 MWe 274 MWe 45.6% >90% 587/850 ◦ C 6.92 MPa 60 kPa 5.8 W/cc Pin-in-block prism LEU once through 14% 120 GWd/ton Once per 2 years 30 days Horizontal, single-shaft 3600 rpm 6.80 MPa 439.1 kg/s 1.88 6 92.8% 28 ◦ C 2.0 20 90.5% Cold-end, diaphragm coupling Synchronous 7 MPa helium cooled 98.7% Plate-fin, 6 modular units 1025 MWt 95% 1.7% Type 316SS Helical-coiled finned tube bundle 313 MWt 22 ◦ C 37 ◦ C Carbon steel (STB410)

353

2000). Intercooled cycle is ruled out because it results in substantial complexities in turbomachine and system but provides no compelling advantage in ultimate plant economy. An intrinsic cycle flow scheme provides RPV cooling and permits use of conventional carbon steels for the vessel construction. The effective RPV cooling provision also allows significant growth in cycle thermal efficiency without major change in plant structural design. These cycle design features are discussed more fully in Section 3. The gas turbine generator is oriented horizontally to lower load demands on bearings and to take advantage of the prevailing design and field experience of industry with turbomachines of similar size, speed and orientation. The helium gas turbine in the non-intercooled cycle, as opposed to its counterpart in an intercooled cycle, exhibits superior aerodynamic efficiency by a fewer number of stages and its rotor is lighter, shorter and more rigid, resulting in more robust rotordynamics. Section 4 discusses performance considerations in the aerodynamic and mechanical design of the GTG system module. Section 5 describes the thermal and structural design measures taken to minimize equipment cost of the heat exchangers included in the HTX module. Section 6 presents the power conversion system maintenance concepts and shows how the system modular arrangement contributes essentially to reduced maintenance operations.

3. Power conversion cycle design approach 3.1. Selection of non-intercooled cycle The GTHTR300 power conversion cycle selects no cycle intercooling (see Fig. 2). This selection was concluded from the system design studies that JAERI had performed on intercooled and non-intercooled cycle plants. The plant designs were studied in sufficient details and under similar design conditions to allow cost performance to be estimated and compared. Comparative system performance, cycle component design and costing, and plant construction and generation costs are summarized here (see Table 2). More details of economic evaluation were reported for the non-intercooled plant by Kunitomi et al. (2002).

354

X. Yan et al. / Nuclear Engineering and Design 226 (2003) 351–373

Fig. 2. GTHTR300 cycle diagram and baseline process parameters.

The intercooled cycle (single stage of compressor intercooling) was estimated to have an efficiency advantage of up to 2% when all important cycle performance parameters such as turbomachine efficiencies and circuit pressure losses are accurately accounted for. Even with the largest efficiency deficit estimated, however, the non-intercooled plant still offers comparable cost of electricity because its first or construction cost is substantially (about 7.8%) lower. The lower first cost of the non-intercooled plant results from its fundamentally simpler design, including a reduced component count (one less each of compressor and intercooler), a fewer number of turbine and compressor stages as a result of a reduced cycle pressure ratio for peak efficiency, a simpler integrated turbomachine system, less cooler surface due to more effective cycle waste heat rejection (in larger LMTD, i.e. logarithmic mean temperature difference, and using precooler only), and finally a smaller primary pressure vessel and piping system. The only major component that costs more in the non-intercooled cycle is a larger recuper-

ator needed to deliver 20% more thermal duty than that of a comparable intercooled-cycle recuperator. The non-intercooled cycle is selected because it allows significant system design simplification while achieving comparable economy, and is in keeping with the SECO design approach of the GTHTR300. 3.2. Cycle performance optimization and growth Cycle thermal efficiency is most sensitive to core outlet coolant temperature, which equals turbine inlet temperature. While keeping all other cycle parameters unchanged, cycle efficiency increases by 1.5% for each increment of 50 ◦ C in core outlet temperature (see Table 3). The net effect remains great even if turbine blade cooling is added due to turbine inlet temperature approaching to blade alloy temperature limit. This is because turbine cooling flow has relatively a far less effect on cycle efficiency. Therefore, the highest core outlet temperature allowed by fuel should be selected. Judged from the currently available fuel performance data, the SiC-coated and advanced ZrC-coated particle

X. Yan et al. / Nuclear Engineering and Design 226 (2003) 351–373

355

Table 2 Cost performance comparison of non-intercooled and intercooled cycle plants Design items

Non-intercooled cycle plant

Intercooled cycle plant

600 MWt 850 ◦ C 2.0

600 MWt 850 ◦ C 2.4

Cycle pressure loss Turbine polytropic efficiency Compressor polytropic efficiency

6.4% 92.8% 90.5%

7.1% 92.5% 90.0%

Turbine disc/casing cooling flow

4.48 kg/s

5.76 kg/s

Net generation efficiency Net electric power output Plant load factor Plant life

45.6% 274 MWe 90% 60 years

47.6% 286 MWe 90% 60 years

1

1.44

Recuperator

1

0.84

Precooler + intercooler

1

1.70

PCS vessels

1

1.29

Piping Building Overall PCS cost ratio

1 1 1

1.03 1.02 1.17

53,846

58,074

196

202

1.60 1.48 1.04 4.12

1.66 1.46 1.00 4.12

System performance Reactor power Reactor outlet gas temperature Cycle pressure ratio

PCS cost ratio Turbocompressor

Plant construction cost Plant construction cost (million
) (1 US$ ≈ 120 Japanese
) Unit construction cost (thousand
/kWe) Cost of electricity (
/kWh) Capital O&M Fuel Total

fuels are capable of core outlet coolant temperature in the range of 850–950 ◦ C. Given a core outlet temperature, selection of core inlet coolant temperature would essentially determine a nominal cycle design point together with cycle

Major differential design drivers: non-intercooled design → intercooled design

Selection of pressure ratio for peak efficiency in both cycles → added intercooler and associated ducting losses Average turbine volume flowrate: 188 m3 /s → 168 m3 /s Average compressor volume flowrate: 66 m3 /s → 58 m3 /s inlet/outlet losses: 1 intake + 1 diffuser → 2 intakes + 2 diffusers Cooling of 6 turbine discs/shrouds → 8 turbine discs/shrouds → up to 2% increase in efficiency

# of turbines (# of stages): 1 (6) → 1 (8) # of compressors (# of stages): 1(20) → 2 (29) # of radial/axial bearings: 2/1 → 4/2 1025 MWt/95% effective/35,410 m2 surface → 820 MWt/95% effective/28,035,410 m2 surface Precooler: 313 MWt/37 ◦ C LMTD/4,206 m2 surface → 181 MWt/27 ◦ C LMTD/3,579 m2 surface Intercooler: none→ 123 MWt/19 ◦ C LMTD/3,502 m2 surface GTG vessel: Ø5.7/30 m length → Ø7.5/32 m length (including intercooler) HTX vessel: Ø5.8/29.7 m length → Ø5.8/27.0 m length → added intercooler piping → larger silo for intercooled PCS installation

pressure ratio and thermal efficiency (see Fig. 3). To maximize thermal efficiency, core inlet temperature should be selected higher than 580 ◦ C for the range of 850–950 ◦ C core outlet temperatures identified above. In the case where core inlet coolant flow is used to

356

X. Yan et al. / Nuclear Engineering and Design 226 (2003) 351–373

Table 3 Cycle efficiency sensitivity to key cycle parameters Cycle parameters

Change in cycle parameter

Change in cycle efficiency (%)

Turbine inlet temperature Recuperator effectiveness Reactor vessel cooling flowa Turbine cooling flowa HP circuit pressure drop LP circuit pressure drop Compressor inlet temperature Compressor efficiency Turbine efficiency Shaft seal leak flowa,b

+50 ◦ C +1% +1% +1% +1% +1% +5 ◦ C +1% +1% +1%

+1.5 +0.8 −0.8 −0.7 −0.7 −0.7 −0.6 +0.5 +0.4 −0.4

b

Percentage of compressor discharged flow. Seal considered is located at compressor intake.

cool reactor pressure vessel, core inlet coolant temperature must be limited to below 480 ◦ C to permit any cost-acceptable pressure vessel design. Fortunately, an innovative RPV cooling scheme incorporated into the GTHTR300 is made independent of core inlet coolant flow, as to be detailed in Section 3.3. This allows selection of 587 ◦ C core inlet temperature, for 55%

reactor ourlet temperature

4.0

o

950 C

3.5

Cycle Thermal Efficiency

50% o

850 C

3.0

45% 2.5 o

40%

950 C

Cycle parametric assumptions: 92.8% turbine polytropic efficiency 90.5% compressor poly. efficiency 95% recuperator effectiveness 6.4% cycle pressure loss o 28 C compressor inlet temperature

35% 350

400

450

500

Cycle Pressure Ratio

a

a given core outlet temperature of 850 ◦ C, simply to achieve peak cycle thermal efficiency. The selection of high core inlet temperature leads to other general design benefits such as low cycle pressure ratio and reduced peak fuel operating temperature. The cycle pressure ratio corresponding to peak cycle efficiency is around 2.0 in the non-intercooled cycle (see Fig. 3). This low cycle pressure ratio provides an optimum cost performance design condition for helium turbine and compressor. First, the cycle thermodynamics is such that the lower the cycle pressure ratio is, the higher the volume flow through turbomachine would be. Aerodynamic performance of helium turbomachine is strongly affected by design volume flow, and raising volume flow leads to more efficient turbomachine design, which in turn augments cycle thermal efficiency strongly. Second, low cycle pressure ratio helps simply turbomachine mechanical design by reducing the total number of stages in turbine and compressor. On the other hand, an adverse effect of low cycle pressure ratio is high recuperator thermal load. However, any increase in recuperator equipment size is partly mitigated by the extremely high heat exchanging capacity (30 MW/m3 ) of the plate-fin

2.0 o

850 C

550

600

650

1.5 700

o

Reactor Inlet Temperature ( C) Fig. 3. Correlations of cycle pressure ratio and thermal efficiency with respect to reactor core inlet and outlet coolant temperatures.

X. Yan et al. / Nuclear Engineering and Design 226 (2003) 351–373

357

Fig. 4. Projected plant efficiency range for the GTHTR300.

design whose optimization is the subject of discussion in Section 5. The net plant efficiency is at 45.6% estimated from base component performance parameters (see Fig. 4). Ultimately though detailed design validation and continual advancement in component technologies, up to 50% efficiency may be attained without any significant change in component or system structural design. This attractive growth option derives from the present cycle’s unique design. First, the cycle growth by increasing reactor core outlet temperature from 850 to 950 ◦ C requires no structural design change to the RPV system (e.g. material selection) because the intrinsic cycle design provision for the RPV cooling remains the same and effective. Furthermore, the increase in core outlet coolant temperature, i.e. turbine inlet temperature (TIT), causes the cycle pressure ratio for peak efficiency to increase whereas higher recuperator effectiveness causes the same to decrease, both weighing comparable effect. As a result, the optimum cycle pressure ratio remains practically unchanged from the baseline cycle having 850 ◦ C TIT and 0.95 recuperator effectiveness to the growth cycle having 950 ◦ C TIT and 0.96 recuperator effectiveness (see Fig. 4). The constant cycle pressure ratio would retain main cycle thermodynamic conditions and thus keep principal structural design features of components and sys-

tem unchanged as the GTHTR300 approaches to the advanced performance regime. 3.3. Intrinsic cycle flow provision for reactor pressure vessel cooling An innovative method of cycle flow circulation accomplishes cooling to reactor pressure vessel. The core hot coolant is circulated to and from the core through the inner piping of a pair of coaxial vessel ducts level-located near the reactor bottom, whereas the cycle top pressure helium discharged from the compressor at about 136 ◦ C and 7 MPa is circulated through the RPV bottom interior via the outer annular passages of the same coaxial vessel ducts as shown previously (see Fig. 1). The unique double vessel duct structure is intended to confine core inlet and outlet high temperature flows in the central graphite core while submerging all of core metallic externals in the cold top-pressure helium gas of the cycle. The core inlet coolant at 587 ◦ C flows inside the side reflector and effectively limits the core lateral graphite structures to similar temperatures. A small gas stream is bypassed off the main flow circulating in the RPV bottom interior and vented into the annulus between the vessel and core barrel (see Fig. 5). From there, the bypass flow is driven to move upward

358

X. Yan et al. / Nuclear Engineering and Design 226 (2003) 351–373

RPV max. operating temperature (˚C)

Fig. 5. Reactor pressure vessel cooling bypass flow driven entirely by intrinsic positive pressure gradients toward the central graphite core.

under a succession of intrinsic positive pressure gradients toward the central core and enter the top core through the control rod guide-tubes. The pressure gradients driving the bypass flow are inherently kept small and nearly constant, only to match the relative pressure drops of the corresponding passages in the cycle high-pressure flow circuitry. It is this intrinsic design provision for cycle flow circulation that cools the reactor pressure vessel in addition to cooling the core barrel and bottom metallic support as well as the top control rod drive mechanism in suitable nominal operating temperatures. The actual operating temperature of the reactor pressure vessel varies with the rate of bypassed cooling flow (Fig. 6). A 0.2% bypassed cooling flow is seen adequate to keeping the vessel operating temperature in good margins from the vessel material design limit (i.e. 371 ◦ C for conventional carbon steel SA533/SA508) over the entire potential range of power operations. This is also in the temperature range where the vessel materials’ irradiation properties are sufficiently known. In case of accidents such as loss of coolant or loss of coolant circulation, where vessel cooling flow is lost, the reactor pressure vessel is cooled by the reactor cavity cooling system (RCCS) and maintained below the vessel temperature limits of safety and investment protection (Yan et al., 2003).

400 371˚C vessel material design temperature limit

350 20% power

300 100% power 250 0.0

0.5

1.0

1.5

2.0

Vessel cooling flow (% of main flow) Fig. 6. Reactor pressure vessel maximum operating temperature with respect to admitted cooling bypass flow rate as a percentage of main circuit flow rate.

X. Yan et al. / Nuclear Engineering and Design 226 (2003) 351–373

4. Gas turbine generator (GTG) module design approach 4.1. GTG system integration The GTG module is configured as a standalone, horizontally-oriented pressure vessel unit (see Fig. 7). The standalone configuration eases maintenance access to internal turbomachinery while the horizontal rather than vertical orientation simplifies turbomachinery bearings. The gas turbine is of axial-flow design including a six-stage turbine and a twenty-stage compressor and drives a synchronous generator on the same shaft at 3600 rpm. Since the GTG module interior can be easily accessed, the gas turbine casings are bolted, instead of supported on sliding rings, to vessel interior and can be disconnected during maintenance replacement as discussed in Section 6. A flexible shaft diaphragm couples the gas turbine rotor to the generator rotor, making alignment and vibration of each rotor group effectively independent of the other. Each of the rotors is supported by bearings at two rotor ends. The bearing loads are minimized by lightweight rotor designs, by avoidance of excessive load on any one bearing in the horizontal arrangement, and by an essentially balanced rotor thrust in the present gas turbine design. The horizontal orientation takes advantage of

359

traditional installation and maintenance practice for large-scale and high-speed turbomachinery. The successful experience with modern air-breathing gas turbines and steam turbines is instrumental to establishing most design details of the present helium gas turbine. This also means that less technical risk in development can be expected of the present helium gas turbine. With an exception of magnetic bearings used, which eliminates potential lubricant contamination to reactor circuit, the mechanical design features of the helium gas turbine closely resemble their respective counterparts in conventional gas turbines. These include similar count of turbine and compressor stages designed to comparable stresses in blades and disks, similar inlet and outlet geometries, similar bearing span and arrangement, and the same rotor orientation. Likewise, the aerodynamic designs of the helium turbine and compressor incorporate the same flowpath design principles including advanced blading proven in modern gas or steam turbines. The helium gas turbine has additional design advantages that augment aerodynamic performance, including low (<0.4) Mach number and high Reynolds number (e.g. ∼107 in compressor), which characterize the highly pressurized helium cycle. The reference design selects a submerged generator, while external generator remains a design option whose technical feasibility rests with the availability

Fig. 7. Design features of the standalone, horizontal gas turbine generator (GTG) module.

360

X. Yan et al. / Nuclear Engineering and Design 226 (2003) 351–373

and safe operations of a large shaft dry gas seal on primary pressure boundary. The external generator option would provide significant design benefits of conventional oil bearings, greater maintenance access as well as lower generator cost. In the case of the submerged design, the shaft penetration on primary pressure boundary is avoided for simplified operations and safety. However, the magnetic bearings are required on the submerged generator rotor for the same reason as is used in the gas turbine. Although the pressurization (about 7.0 MPa) tends to increase windage loss, this effect is fully compensated for by highly pressurized helium being an efficient coolant. As a result, the helium-submerged generator can be designed to deliver a level of efficiency comparable to traditionally-cooled units. 4.2. High performance aerodynamic design Turbomachine aerodynamic efficiencies are sensitive performance parameters of plant cycle (see Table 3). The non-intercooled cycle presents favorable aerodynamic design conditions such as high turbomachine volume flow rates. Further improvement in turbomachine performance can be obtained by exploiting advanced aerodynamic design techniques. Detailed aerodynamic design features of the helium turbine and compressor are described. 4.2.1. Turbine aerodynamic design features Turbine design is based on efficient reaction stages with similar tip speed and stage loading to those employed in conventional designs. This results in six turbine stages employing uncooled blades made of conventional directionally solidified alloy. In addition, the turbine includes advanced design techniques in order to break performance limitations of conventional design methods. In the case of conventional blading design the magnitude of secondary loss is in the same order as that of the profile loss. The advanced 3D blading can optimize secondary flows. The stator incorporates bowed blades of high aspect ratio that significantly reduces endwall losses whereas the rotor uses blades that incorporate increasing relative inlet flow angles toward the hub to mitigate the adverse Mach number effect in that region. The 3D blading design techniques are shown to reduce secondary loss significantly. As a re-

sult, the secondary loss in the present helium turbine is shown to contribute to a small fraction of the total stage loss in all six stages (see Fig. 8). The stage efficiency is improved by a 0.5–1.0% point over performance of an otherwise conventionally-designed stage. The polytropic efficiency of the final helium turbine design is estimated at 92.8%, which accounts for turbine exhaust loss. 4.2.2. Compressor aerodynamic design features Despite the many common design parameters (number of stages, tip speed, flow coefficient, diffusion factor, etc.) shared with air-breathing gas turbine compressors, the helium compressor presents some unique design challenges. For example, the helium compressor flowpath is essentially parallel with high hub-to-tip ratios (around 0.9) throughout the twenty stages as opposed to a similar flowpath being observed only in the rear stages of a high pressure ratio air-breathing compressor. This unique design feature makes secondary flow design considerations more crucial to meet efficiency goal and stall margin of helium compressor. A high reaction stage design, proven effective in production gas turbine experience to reduce flowpath friction loss and tendency for stall, is employed (Fig. 9a). The advanced blading techniques incorporated in the present helium compressor design include the blade-end bends to eliminate flow separation and the blade over-camber to compensate for flow distortion near the endwall (see Fig. 9b and c). These advanced design features are combined with optimum blade row solidity and aspect ratio for efficiency and flow stability. Finally, CFD optimizations of inlet and outlet geometries are performed to control pressure losses in these locations (Fig. 9d). The gas turbine standalone vessel installation provides ample vessel interior space in sizing the inlet and outlet geometries to facilitate fluid dynamic designs. The final compressor performance is projected of 90.5% polytropic efficiency, including inlet and outlet losses, with a 30% conservative surge margin desired in the initial design development. The plan to validate the helium compressor design includes two consecutive steps of testing for a model compressor (Takada et al., 2003). The model compressor consists of four stages having one-third dimensional scale and about one-ninth aerodynamic scale of the full machine flowpath. The testing of the

X. Yan et al. / Nuclear Engineering and Design 226 (2003) 351–373

361

Fig. 8. Comparison of stage-wise aerodynamic losses of helium turbines employing conventional and advanced blades, indicating the potential of incorporating advanced blading techniques to reduce secondary losses.

model compressor incorporating the above-identified design features is being conducted in a dedicated helium loop to map the compressor performance in the high speed range. The key performance evaluations include compressor stall/surge margin, endwall losses in the multiple rows of rotating blades in addition to inlet and outlet helium flow performance. The second step of testing includes transient performance verification for startup, turndown, load control of the model compressor. Construction of a subscale helium gas turbine plant is planned for the transient compressor test as part of overall operational and control tests of the integrated plant system. The performance data generated in the tests of the helium compressor model will be used to benchmark the design codes that are used for development of the full-size helium compressor and which have until now been established from air-breathing compressor experience. 4.3. Turbomachinery mechanical design 4.3.1. Turbomachinery orientation Turbomachine may be installed in either vertical or horizontal orientation, and the former was preferred by

traditional direct cycle modular reactor designs. Both installation options have been evaluated in the present work with respect to pertinent design conditions of large scale and high speed turbomachine supported on magnetic bearings. The effects of turbomachinery orientation on rotordynamics, bearing requirements, maintainability and plant building arrangement were investigated (see Table 4). Optimum building arrangements were studied for the horizontal and vertical installation options, and major difference between the arrangements was limited to the turbine building area. Taking account of the spaces required for both installation and maintenance, the horizontal design calls for a larger plan area in turbine building whereas the vertical design for a taller turbine building elevation. Overall plant building volume is about 6% smaller for the vertical design than for the horizontal design, potentially saving 1% in plant construction cost for the vertical design. Notwithstanding its potential economical benefit, the vertical turbomachinery installation poses significant technical challenges, especially to existing bearing technologies. Thrust loads in magnetic bearing and auxiliary catch bearing were near or beyond the

362

X. Yan et al. / Nuclear Engineering and Design 226 (2003) 351–373

Fig. 9. Elements of high performance helium compressor design approach in the GTHTR300.

current design limits in the vertical design. In the case of horizontal turbogenerator arrangement where static and dynamic loads are not jointly concentrated on any one bearing, bearing loads are reduced to within the technical feasibility of the present technologies. In addition, the horizontal turbomachinery design benefits importantly from the prevailing experience of industry in installation, operations and maintenance

of machines in the weight and speed of interest. The horizontal turbomachinery orientation was considered advantageous in the present system because it simplifies technical solutions in critical design areas. 4.3.2. Rotor thrust balance Thrust bearing is used to counter the net of aerodynamic and pressure thrust forces of a rotor. Ideally,

X. Yan et al. / Nuclear Engineering and Design 226 (2003) 351–373

363

Table 4 Evaluation of horizontal and vertical rotor orientations for 300 MWe turbomachinery Evaluation item

Horizontal orientation

Vertical orientation

Evaluation

Rotor design

Single-shaft and diaphragm-coupled turbogenerator rotor with submerged generator

The same

Rotordynamics

Critical vibration amplitude < 125 ␮mp–p (ISO G2.5)

As good as horizontal rotor’s if the same bearing support stiffness as horizontal one’s is achieved

No difference

Generator rotor: 44 ton (=35 ton weight + 9 ton dynamic load)

Generator rotor: 9 ton (=0 weight + 9 ton dynamic load)

Thrust load is near stress–strain limit (110 ton) of bearing disc for vertical orientation

Turbine rotor: 63 ton (=0 weight + 63 ton steady plus transient load)

Turbine rotor: 108 ton (=45 ton weight + 63 ton steady plus transient load)

Operational vibration amplitude < 75 ␮mp–p (ISO G2.5) Magnetic bearing loading Radial load (max.)

Thrust load (max.)

Auxiliary (catch) bearing loading Radial load (max.) Generator rotor: 55 ton (=MB load capacity + 11 ton seismic load) Thrust load (max.) Turbine rotor: 77 ton (=MB load capacity + 14 ton seismic load)

Generator rotor: 20 ton (=MB load capacity + 11 ton seismic load) Turbine rotor: 117 ton (=MB load capacity + 9 ton seismic load)

Thrust load is beyond current design base for vertical orientation

Maintainability

Mostly conventional operations

Mostly first-of-a-kind operations

More difficult for vertical orientation

Static helium seals

Not necessary (bolted connection is used)

Not used (seals must be avoided to prevent large helium leaks; use of bolted connection is assumed)

No difference

Plant building volume (for a 4-reactor plant)

461 km3 (= reactor building 302 km3 + auxiliary building 159 km3 )

433 km3 (= reactor building 274 km3 + auxiliary building 159 km3 )

6% smaller building volume for vertical orientation

thrust bearing load becomes zero if the aerodynamic and pressure thrust forces cancel each other completely. This is nearly the case in the present gas turbine aerodynamic design and layout (see Fig. 10). The sum of the rotor thrusts at full-power operation is about 1 ton only. Nevertheless, a conservatively larger load capacity of 50 ton is adopted in the gas turbine thrust bearing design in consideration of transient thrust and seismic force and a load capacity of about

10 ton is set for the generator rotor thrust bearing. These thrust loading requirements will be verified in system transient simulation and subsequent tests of the subscale helium turbine plant. 4.3.3. Lightweight rotor design The achievable stiffness of magnetic bearing is considerably lower than that allowed by oil bearing. In light of this, rotor weight reduction becomes

364

X. Yan et al. / Nuclear Engineering and Design 226 (2003) 351–373

Fig. 10. Turbomachinery design to achieve a nearly balanced axial rotor thrust, which minimizes load demands on magnetic and auxiliary thrust bearings.

critical to shaping satisfactory rotordynamics. Available lightweight rotor construction methods were contemplated whenever feasible to accomplish this. A hollow disc design is selected for both turbine and compressor rotors. While the turbine rotor is assembled of typical built-up discs, a welded rotor is selected for the compressor to minimize weight. Preliminary design of turbine disc considers a conventional hollow disc structure, in which the stress concentration near the fore catenary ring was found to have exceeded the material design limit (see Fig. 11). The disc structure was then modified to allow optimum stress distribution by means of finite element method (FEM) modeling. The modifications made include moving the catenary rings radially outwards, relocating stator seal diaphragm, and slightly increasing disc thickness around inner and outer rims. As a result, all design stress limits are satisfied in the final disc design albeit the disc weighs slightly heavier. 4.3.4. Rotordynamic design optimization The sheer rotor size, high transmitted shaft power and inevitable use of magnetic bearings complicate

rotordynamic design. In additional to reducing rotor weights, rotordynamics can be significantly improved by a flexible diaphragm coupling of the gas turbine and generator rotors. Each rotor is significantly shortened and stiffened. As a result, the critical vibration modes are reduced in the operating speed range, and alignment and vibration control are made essentially independent of the two rotors. The results of finite element vibration analysis (see Table 5) indicate that the nominal operating speed (3600 rpm) has a margin of 20% or larger from the nearest critical speeds for both rotors, based on the magnetic bearing design characteristics of stiffness and damping. Although one critical bending mode remains below the rated speed for the turbine rotor and two critical bending modes for the generator rotor, all resonant Q-factors are smaller than 10, which indicates adequate damping to these and other critical speed modes. In addition, the magnetic bearing system enables additional control of critical speed vibration through active modulation of bearing stiffness and damping. The calculated unbalance vibration amplitudes at critical and selected operating speeds meet

X. Yan et al. / Nuclear Engineering and Design 226 (2003) 351–373

365

Fig. 11. Turbine disc structural design modifications to optimize strength and disc cooling.

the appropriate standards of ISO G2.5 for the turbine and generator rotors. The validation for the rotordynamic design involves two ongoing development and test programs. First is development and test for a diaphragm coupling of

the size and rating required of the present design. Its ability to transmit steady and dynamic torques with the coupling flexibility specified by the rotordynamic design will be verified. Second, a two-span rotor model in one-third of the full scale rotor train will

366

X. Yan et al. / Nuclear Engineering and Design 226 (2003) 351–373

Table 5 Critical vibration modes and damping factors of the gas turbine and generator rotors Vibration mode

1st 2nd 3rd 4th 5th

Gas turbine rotor

Generator rotor

Critical speed (rpm)

Margin from rated speed (3600 rpm) (%)

Resonance damping Q-factor

Critical speed (rpm)

Margin from rated speed (3600 rpm) (%)

Resonance damping Q-factor

641 1700 2588 4447

−82 −53 −28 +24

1.8 1.5 2.8 7.7

824 1007 1812 2882 4661

−77 −72 −50 −20 +29

4.7 5.6 1.8 5.1 5.2

be constructed and tested for evaluation of appropriate integrated vibration control methods of multiple magnetic bearings (Takizuka et al., 2003). Analytical techniques of finite element method and potential control algorithms for the magnetic bearing system will be evaluated for rotordynamic controllability up to the second rotor bending mode. 4.3.5. Minimization of turbine cooling flow Because overall cycle efficiency drops rapidly with increased flow consumption in turbine cooling (see Table 3), the present turbine design employs a turbine cooling scheme (see Fig. 12) that successfully limits the total of turbine rotor and casing cooling flows to less than 1% of main flow. First of all, absolute turbine cooling flow requirement is low because only six stages are present in the low-expansion-ratio turbine. The relative effect on cycle efficiency of this already reduced cooling flow

becomes even less pronounced because the main turbine flow rate is large in the present non-intercooled cycle. Further design factor contributing to the low turbine cooling flow requirement stems from the standalone gas turbine installation in the pressure vessel, in which the entire turbine casing externals, including inlet and exhaust pieces, are exposed to strong convective cooling of the low temperature (about 136 ◦ C) main flow that is discharged from the compressor. Only casing internals need dedicated cooling. The main section of casing internals are cooled with flow streams bled off of the main flow and the exhaust scroll casing internals are cooled by recovered leak flow of the turbine aft shaft. The temperatures and related stress and displacement of the integral turbine casing are confirmed satisfactory by FEM analysis (Fig. 13). Finally, the turbine cooling flow is conserved by an efficient cooling flowpath design (see Fig. 12),

Fig. 12. Details of turbine cooling flowpath designed to minimize cooling flow requirements for discs, casing and bearing.

X. Yan et al. / Nuclear Engineering and Design 226 (2003) 351–373

367

Fig. 13. Results of thermal and stress FEM analysis of turbine casing and exhaust scroll.

in which turbine discs share the common cooling flow streams used to cool turbine fore bearing and bearing support. The turbine disc cooling flow is then introduced near disc outer rim, rather than from disc bore, thereby increasing convection effectiveness. The disc cooling flowpath was depicted previously (Fig. 11) and the discs are maintained at temperatures below the 450 ◦ C limit of Ni–Cr disc material.

5. Heat exchangers (HTX) module design approach 5.1. HTX module system integration The HTX module contains the recuperator and precooler in a vertical steel pressure vessel (see Fig. 1). The recuperator is made up of six compact platefin heat transfer units that operate in parallel and

368

X. Yan et al. / Nuclear Engineering and Design 226 (2003) 351–373

locate in the upper pressure vessel. The precooler is a helically-coiled tube bundle with water circulating in the tubes and helium in the shell. The tube bundle is placed in the lower pressure vessel section. Since essential technologies employed in the recuperator and precooler designs have been developed in Japan, the primary design objective concerns with cost minimization through optimal thermal design and fabrication process. 5.2. Recuperator cost reduction The first approach to lower recuperator cost is minimization of weight and volume given the thermal duty specified of the cycle design. The fact that both sides of the recuperator operate in non-corrosive and non-erosive helium flows permits use of extremely compact plate-fin surface. For fixed heat transfer duty and effectiveness, the weight and volume of recuperator depends heavily on selection of fin compactness (see Fig. 14). This means that as small a fin geometry (fin height and fin pitch) as practical in fabrication should be used. The recuperator design selects a fin size of 1.2 mm height by 1.2 mm pitch (designated as 1.2 × 1.2 surface in Fig. 14). Further attempt at cost reduction examined the high cost areas of the recuperator fabrication pro-

cess. One significant question was what level of fin-to-plate surface connectivity is desired and should be acquired in fabrication. Fins are brazed to plates in furnace to form a strong integral structure to withstand internal high-pressure and thermal loads. Fabrication cost closely correlates to the rate of perfect bounding to be achieved between fins and plates. The FEM analysis shows the recuperator structure suffers little loss of strength, in terms of stress and displacement under static thermal and pressure loads, even if the fin-to-plate connectivity is as low as 80% (see Fig. 15). If similar results are to be confirmed under transient and cyclic conditions, modification of recuperator design specification at an appropriately reduced fin-plate surface connectivity promises considerable saving in fabrication cost. 5.3. Precooler cost optimization Similar installation and cost design objectives were observed in precooler design. A nominal design benefit provided by the non-intercooled cycle is that the cycle waste heat rejection occurs in entirely in precooler over a broad and high temperature range. This increases effective logarithmic mean temperature difference (LMTD) between the gas and water

Fig. 14. Effective volume and weight of a recuperator unit with respect to plate-fin surface compactness.

X. Yan et al. / Nuclear Engineering and Design 226 (2003) 351–373

369

Fig. 15. FEM analysis of the proposed low-production-cost recuperator with a reduced fin-to-plate brazing connectivity.

flows and thus reduces heat exchanging surface area requirement for precooler thermal design. To obtain a compact installation envelop, smooth and finned tube designs are compared in equivalent thermal performance conditions (Table 6). The finned tube design is selected and shown to have 20–30% saving in tube bundle volume and weight compared to the smooth tube design variant.

6. Power conversion system maintenance approach 6.1. Maintenance requirements The maintenance requirements and methods for the power conversion system have been studied in detail (Kosogiyama et al., 2003). Basic maintenance consid-

370

X. Yan et al. / Nuclear Engineering and Design 226 (2003) 351–373

Table 6 Comparison of smooth and finned tube precooler designs under equivalent thermal duty Items

Unit

Smooth tube design

Finned tube design

Design type Thermal rating LMTD

– MWt ◦C

Helical tube bundle 323 37.9

Helical tube bundle 323 37.9

Fluid design condition Shell side Fluid Flowrate Inlet/outlet temperature Inlet pressure Pressure loss

– kg/s ◦C MPa kPa

Helium gas 440.8 169.2/28.0 3.55 20

Helium gas 440.8 169.2/28.0 3.55 24

Tube side Fluid Flowrate Inlet/outlet temperature Inlet pressure Pressure loss

– kg/s ◦C MPa kPa

Water 2763 22/50 0.50 47

Water 2763 22/50 0.50 31

– mm mm mm m m3 ton

3249 28.8 2.85 – 23.2 93 138

3249 28.8 2.85 1.5 14.4 70 95

Sizing dimensions Number of tubes Base tube outer diameters Base tube wall thickness Fin height Tube length Tube bundle volume Tube bundle weight

erations given in the present modular system design and related to the horizontal turbomachine layout are discussed here. As evident as the separated arrangement and housing of the three pressure vessels containing the reactor, turbomachine and heat exchanging equipment, respectively (see Fig. 16), the plant system is laid out with not only modular construction but modular maintainability in mind. The latter proves effective in simplifying inspection and maintenance (I&M) operations for the power conversion system. The PCS components are of diverse I&M requirements because of different working conditions and due to pertinent regulatory requirements and industrial practice. For instance, the gas turbine equipment is expected of the most frequently scheduled, and likely forced, maintenance activities of all power conversion system. Therefore, easy access and replacement is essential in turbine and compressor maintenance. Further maintainability goal for the PCS includes completion of most I&M activities within a reactor refueling period, without adverse effect on plant availability.

6.2. Gas turbine and generator maintenance The gas turbine removal and replacement is performed of overhaul maintenance. Two gas turbine removal schemes are proposed. In the first scheme (refer to Fig. 16), the gas turbine and generator pressure vessel sections are separated at the vessel flanges. The generator vessel section is pulled away on rails in axial direction (marked as step 1). The gas turbine is then removed from the pressure vessel interior and lifted by crane (marked as step 2) to a turbine service area common to multiple reactors on a plant site. A refurbished gas turbine is replaced via the reversed steps. A 250-ton crane capacity is required for the gas turbine movement. In an alternative, modular replacement scheme (see Fig. 17), a remotely-operated tool is inserted into the pressure vessel interior through the opened hatches located on the vessel exterior wall (step 1) and disconnect the internal gas duct from the turbine intake and separate the diaphragm coupling between the

X. Yan et al. / Nuclear Engineering and Design 226 (2003) 351–373

371

Fig. 16. Reactor refueling and power conversion system maintenance equipment and operations in the GTHTR300.

gas turbine and generator shafts. Once the bolts on the vessel flanges and on the lower vessel duct are removed (step 2), the turbine building crane is used to lift the gas turbine vessel section as “cartridge” up to the ground floor (step 3). The crane capacity approximately doubles that of the previous scheme. Sealed at both ends, the pressure vessel acting as a cask provides shielding and dust retention from interior radioactivity during movement. The removed gas turbine vessel is then taken to the turbine service area where maintenance is performed over a period of 6 months including initial cooling or necessary decontamination. A previously refurbished and fully-aligned gas turbine vessel is installed in the reversed steps. The proposed maintenance method requires one

spare gas turbine vessel “cartridge” in a four-reactor plant. Generator maintenance follows removal of the gas turbine. The generator rotor, which is radioactively free, is pulled out of the pressure vessel shell toward the vacated space by the removed gas turbine and can be serviced in the turbine building floor. 6.3. Heat exchangers maintenance No scheduled maintenance removal is expected for the recuperator and precooler. Should a damage in life warrants a replacement, they may be removed through the opened top closure of the HTX pressure vessel by using the reactor service crane (see Fig. 16). The

372

X. Yan et al. / Nuclear Engineering and Design 226 (2003) 351–373

removal and replacement scheme of the recuperator units involves a few straightforward steps (Fig. 18). The precooler tubes can be inspected or repaired, by plugging of any damaged tube, from outside of the pressure vessel. The water in the tubes is completely drainable.

7. Summary and conclusion

Fig. 17. An alternative gas turbine replacement scheme—integral turbine-vessel “cartridge” replacement.

The GTHTR300 follows the SECO design approach. The reference plant system is established and the detailed design including safety and economical evaluation is being performed. The design validation includes development and testing of scaled-down key components and integrated system. This paper discussed the unique cost and performance design approach to the GTHTR300 power conversion system, which features the non-intercooled cycle optimized to secure economical performance and strong growth potential, the intrinsic cycle provision for reactor pressure vessel cooling that makes the conventional steel vessel system practical, the horizontal and high-performance turbomachinery,

Fig. 18. Removal and replacement steps of recuperator core modular units. These steps are not required in scheduled maintenance rather than in need of repair for an unexpected structural damage.

X. Yan et al. / Nuclear Engineering and Design 226 (2003) 351–373

and modular system construction and maintainability. These design features were shown to offer the optimum solution of efficiency potential, economics, reliability and maintainability in addition to early market deployment. Acknowledgements The authors wish to acknowledge the significant contributions to the GTHTR300 design, particularly to the power conversion system design and development, by a larger number of their colleagues in JAERI, MHI, other industries and utilities. References Katanishi, S., Kunitomi, K., 2002. Safety design philosophy of the gas turbine high temperature reactor (GTHTR300). Trans. At. Energy Soc. Jpn. 2 (1), 55–67 (in Japanese). Kosogiyama, S., Takei, M., et al., 2003. Maintenance methods and procedures for the power conversion system of the gas turbine

373

high temperature reactor 300 (GTHTR300), in preparation (in Japanese). Kunitomi, K., Katanishi, S., et al., 2002. Design study on gas turbine high temperature reactor (GTHTR300). Trans. At. Energy Soc. Jpn. 1 (4), 352–360 (in Japanese). Nakata, T., Katanishi, S., et al., 2002. Nuclear design of the gas turbine high temperature reactor (GTHTR300). Japan Atomic Energy Research Institute, JAERI-Tech 2002-066 (in Japanese). Takada, S., Itaka, H., et al., 2003. The 1/3-scale aerodynamics performance test of helium compressor for GTHTR300 turbo machine of JAERI (step 1), ICONE11-36368. In: Proceedings of the 11th International Conference on Nuclear Engineering, Tokyo, Japan, April 20–23. Takizuka, T., Matsumoto, I., et al., 2003. A verification plan of rotor dynamics of turbo-machine system supported by AMB for the GTHTR300 (step 1), ICONE11-36369. In: Proceedings of the 11th International Conference on Nuclear Engineering, Tokyo, Japan, April 20–23. Yan, X., Shiozawa, S., et al., 2000. Design innovations in GTHTR300 gas turbine high temperature reactor, ICONE-8016. In: Proceedings of the 8th International Conference on Nuclear Engineering, Baltimore, MD, USA, April 2–6. Yan, X., Kunitomi, K., et al., 2003. GTHTR300 design and development. Nucl. Eng. Des. 222, 247–264.