Medium temperature carbon dioxide gas turbine reactor

Nuclear Engineering and Design 230 (2004) 195–207

Medium temperature carbon dioxide gas turbine reactor Yasuyoshi Kato∗ , Takeshi Nitawaki, Yasushi Muto Research Laboratory for Nuclear Reactors, Tokyo Institute of Technology, 2-12-1 O-okayama, Meguro-ku, Tokyo 152-8550, Japan Received 8 May 2003; received in revised form 2 October 2003; accepted 5 December 2003

Abstract A carbon dioxide (CO2 ) gas turbine reactor with a partial pre-cooling cycle attains comparable cycle efficiencies of 45.8% at medium temperature of 650 ◦ C and pressure of 7 MPa with a typical helium (He) gas turbine reactor of GT-MHR (47.7%) at high temperature of 850 ◦ C. This higher efficiency is ascribed to: reduced compression work around the critical point of CO2 ; and consideration of variation in CO2 specific heat at constant pressure, Cp , with pressure and temperature into cycle configuration. Lowering temperature to 650 ◦ C provides flexibility in choosing materials and eases maintenance through the lower diffusion leak rate of fission products from coated particle fuel by about two orders of magnitude. At medium temperature of 650 ◦ C, less expensive corrosion resistant materials such as type 316 stainless steel are applicable and their performance in CO2 have been proven during extensive operation in AGRs. In the previous study, the CO2 cycle gas turbomachinery weight was estimated to be about one-fifth compared with He cycles. The proposed medium temperature CO2 gas turbine reactor is expected to be an alternative solution to current high-temperature He gas turbine reactors. © 2004 Elsevier B.V. All rights reserved.

1. Introduction An overview of major gas-cooled power plant projects is shown in Fig. 1 that have been built or are in the planning stages. The first gas-cooled power reactor, the MAGNOX reactor, was developed by the UK and France using carbon dioxide (CO2 ) coolant in a closed cycle at 2 MPa pressure. The core consisted of graphite moderator blocks with holes and fuel elements. Fuel elements, consisting of a natural uranium metal bar and magnesium–aluminum alloy (called Magnox) cladding, were placed into the holes. Coolant flowed through moderator holes and the temperature at the core outlet was about 400 ◦ C. Steam cycle efficiency was about 31%. Subsequent plants, designated as ∗ Corresponding author. Tel.: +81-3-5734-3065; fax: +81-3-5734-2959. E-mail address: [email protected] (Y. Kato).

advanced gas-cooled reactors (AGRs), used 2.3% enriched uranium oxide fuel pellets and stainless steel cladding to achieve higher cycle efficiency of about 40% through elevating coolant outlet temperature to 650 ◦ C. Coolant pressure of AGRs was elevated to 4 MPa to increase average core power density 4–6 times that of MAGNOX reactors for reduction of plant capital costs. However, AGRs were not competitive with PWRs in electricity generation costs. The UK government announced in 1979 that the future nuclear power generation program would be based on construction of PWRs. The US and Germany began high-temperature gas-cooled reactor (HTGR) development programs in the 1960s. Using helium (He) as coolant and small particle fuels with enriched uranium oxide, a high core outlet temperature is attained ranging from 800 to 850 ◦ C without any chemical attack on moderator and fuel materials. Small particles are triso-coated with successive layers of porous carbon, pyro-carbon,

0029-5493/$ – see front matter © 2004 Elsevier B.V. All rights reserved. doi:10.1016/j.nucengdes.2003.12.002

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Y. Kato et al. / Nuclear Engineering and Design 230 (2004) 195–207

of a stack of graphite spheres of about 60 mm diameter with embedded coated fuel particles. The latter reactor core consists of hexagonal graphite blocks with holes and fuel rods. Fuel rods enclosing coated particles are inserted into the holes. A number of experimental and subsequent prototype HTGR plants were built for both types; however, no commercial plant was constructed because of the higher capital cost than contemporary LWRs. Helium heated to such high temperatures as 800–850 ◦ C can be used as a working fluid in a closed gas turbine direct cycle for a power generation system. The direct cycle eliminates the intermediate cooling circuit and steam generators required for PWR. Compared to a steam turbine, a gas turbine is far smaller because gas turbine outlet pressure is much higher than that of the former because of the low pressure ratio (about four) in the gas turbine. Moreover, the gas turbine is much simpler than that of the steam turbine because of absence of the moisture separation and steam extraction systems necessary for the steam turbine. The gas turbine system provides about 46–48% efficiency. These promising features indicate the possibility of the gas turbine direct cycle as a

Nomenclature Cp d
hstage k Nu P Pcrit Pred Pr R Re Tred Tcrit
T V z

specific heat at constant pressure hydraulic equivalent diameter enthalpy drop per turbine stage heat conductivity Nusselt number gas pressure critical pressure reduced pressure Prandtl number gas constant Reynolds number reduced temperature critical temperature temperature drop in a turbine gas volume compressibility factor

and silicon carbide to prevent fission product release. Two types of HTGRs were developed: pebble-bed fuel type reactors in Germany and block fuel type reactors in the US. The former reactor core consists

Coolant

Cycle

1950

1960

1970

1980

1990

2000

Magnox reactor (60-655 MWe)

*1 UK & 1956

Steam Indirect France (Rankine)

AGR (660 MWe)

1976*1

UK

CO2

TIT

Gas Turbine Direct (Brayton)

Japan MIT US 1967*2 Peach Bottom (42 MWe)

Steam Indirect (Rankine)

US

1967*2 AVR (15 MWe)

Germany

He Gas Turbine Direct (Brayton) *

1982*1 Fort St. Vrain (324 MWe)

1986*2

THTR-300 (308 MWe) PBMR (100 MWe) S. Africa GT-MHR (286 MWe) Russia GTHTR-300 (275 MWe) JAERI

1: Start of operation, *2: Rated full power operation

Fig. 1. History of gas-cooled power plant development projects.

Y. Kato et al. / Nuclear Engineering and Design 230 (2004) 195–207

197

Table 1 Critical data of typical fluids Net Work Turbine Work

Turbine Work

Net Work Pump Work

SteamTurbine

Fluids

Critical temperature Tc (◦ C)

Critical pressure Pc (MPa)

CO2 H2 O N2 NH3

31.0 374.0 147.0 132.2

7.38 22.06 3.40 11.13

Compressor Work

Gas Turbine

Fig. 2. Comparison of compressor work in a gas turbine and pump work in a steam turbine.

potential alternative to LWRs. However, extensive development is required for high-temperature resistant materials and the He gas turbine. The principal difference is that the steam turbine works with a fluid that changes phase during the Rankine cycle, whereas the gas turbine works on fluids that remain in the gaseous phase during the Brayton cycle. As shown in Fig. 2, feed-pump work to pressurize water in the liquid phase in the steam turbine is few percents of steam turbine output (Wilson and Korakianitis, 1998), whereas the gas turbine compressor consumes about 50% of the power produced by the turbine. By far, the main efforts to enhance cycle efficiency have specifically addressed increasing turbine output by elevating turbine inlet temperature, while enhancement of recuperator efficiency is also crucial to the cycle in the gas turbine system. Fluids have peculiarities at their critical points or pseudo-critical points in thermo-mechanical properties: they exhibit strong peaks in specific heat, thermal conductivity and viscosity. Moreover, they have a peculiarly deep drop in their compressibility factors. Surface tension and heat of vaporization become zero at the critical point. This study investigated a new method in a CO2 cycle for improving cycle efficiency by reduction of compressor work through utilizing the peculiar drop in the compressibility factor around the critical point. The usual compressor inlet temperature of about 35 ◦ C, which is usually determined by cooling sea water temperature, and critical temperatures of fluids shown in Table 1 present a great advantage for this system: a peculiar drop in a compressibility factor around the critical point can be

utilized to reduce compressor work in a CO2 gas turbine cycle because its critical temperature (31.0 ◦ C) is approximately equal to the usual compressor inlet temperature (35 ◦ C). It can not be utilized practically in the conventional He gas turbine cycle because its critical temperature (−268.0 ◦ C) is much lower than the usual compressor inlet temperature. Coolant CO2 has preferable properties to He. Heat transfer coefficient h between the coolant and fuel cladding surface is calculated as

 k k h= Nu ∝ (Re0.8 Pr0.4 ), d d where k is heat conductivity, d = hydraulic equivalent diameter, Nu = Nusselt number, Re = Reynolds number, and Pr = Prandtl number. The above equation provides 1.5 times higher heat transfer coefficient h in CO2 at the same gas velocity, leading to about a 1.5-times-lower temperature difference between the coolant and fuel rod cladding surface. The higher heat transport capacity, measured as product of specific heat at constant pressure Cp and molecular weight, results in about 2.5 times more effective core decay heat removal under natural circulation conditions than He. The 3.6 times longer depressurization time of CO2 (Lewis, 1977) and higher heat transport capacity (mentioned above) over that of He mitigates the depressurization transient and simplifies design of a passive decay heat removal system described later. Carbon dioxide is about 250 times less expensive than He per unit weight and 24 times less expensive per unit volume. In addition, considering lower leakage rate characteristics of CO2 than He, coolant leakage problems of gas cooled reactors in operation are orders of magnitude less severe than with He. For the He direct cycle, a He gas turbine must be newly developed because He and air have substantially different gas properties, e.g. density (molecular weight), specific heat, and specific heat ratio.

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In contrast, most current gas turbine engine technologies developed for fossil fuel power generation systems could be borrowed because the difference between CO2 and air is much smaller than that between He and air. Considering turbine inlet temperature of modern gas turbines, the so-called “F” class gas turbine is 1350 ◦ C; therefore, material technology needed for a CO2 turbine up to 850 ◦ C is within current technology of fossil gas turbines. Carbon dioxide may react with structural materials and a graphite moderator under high temperature and radiation fields. During extensive experience in MAGNOX reactors and advanced gas cooled reactors (AGRs) over more than 20 years, structural material corrosion problems were eliminated by appropriate material selection according to service temperatures and by reduced vapor content (Gibbs and Popple, 1982). Regarding corrosion of the graphite moderator, addition of methane and carbon monoxide in a range of small concentrations (called the coolant “window”) inhibited reaction with CO2 ; thereafter, satisfactory operations were done without excessive corrosion (Hewitt and Collier, 1997). An indirect Rankine CO2 cycle with a water/steam system has been applied to AGRs and enhanced gas cooled reactors in the UK (Gratton, 1981; Lennox et al., 1998; Abram et al., 2000), for which the means proposed in this study to improve cycle efficiency can not be employed because recompressing process is absent in this indirect cycle. However, a medium temperature CO2 direct cycle proposed in this study could be inherited with the above advantages and experiences.

Now, worldwide interest in small and medium size reactors continues to increase for electricity generation and local heating for cities and islands. In December 1999, the Research Laboratory for Nuclear Reactors of the Tokyo Institute of Technology launched a new program to develop advanced small and medium size nuclear reactors. Under the program framework, we are designing advanced gas-cooled fast and thermal reactors (Kato et al., 2000, 2002, 2003; Nitawaki et al., 2001; Muto et al., 2003) and light water reactors (Yamashita et al., 2001; Ohtsuka et al., 2002) in collaboration with industry. Recently, various CO2 gas turbine cycles have been studied by us and at MIT (Kato et al., 2000, 2001, 2002, 2003; Dostal et al., 2002; Hejzlar et al., 2002). This paper addresses the possibility of a medium temperature CO2 gas turbine reactor with a partial pre-cooling cycle to provide alternatives to current high-temperature He gas turbine reactors, by evaluating thermal cycle performance.

2. Results and discussion Cycle efficiencies of He and CO2 cycles were calculated using PROPATH as a database for thermo-physical properties of fluids (Ito et al., 1990) and evaluation conditions are summarized in Table 2. Cycle efficiency of a Brayton cycle is improved when the number of compression stages with inter-cooling is increased as shown in Fig. 3(a)–(c). However, the contribution of each additional stage

Table 2 Cycle efficiency evaluation conditions Parameters Component efficiency

Turbine adiabatic efficiency (%) Compressor adiabatic efficiency (%) Recuperator effectiveness (%)

Pressure drop (%)a

Usual Brayton cycle (no pre-cooling)

Partial pre-cooling cycleb a b c

Values 90 90 95 Reactor Pre-cooler Intercooler Recuperator (Hb /Lc ) Reactor, pre-cooler, intercooler Recuperator (Hb /Lc )

Percent pressure drop in each component relative to the reactor outlet pressure. High-temperature side of a recuperator. Low-temperature side of a recuperator.

1.5 1.02 0.58 1.99/0.66 Same as above cycle 2.65/0.88

Y. Kato et al. / Nuclear Engineering and Design 230 (2004) 195–207 Turbine

Compressor

199

Low Pressure High Pressure Turbine Compressor Compressor

Generator

Reactor

Generator

Reactor

Intercooler

Pre-Cooler

Pre-Cooler

Recuperator

Recuperator

(a)

(b)

Middle Pressure Compressor High Pressure Turbine Low Pressure Compressor Compressor

Generator Reactor Intercooler I Intercooler II Pre-Cooler

Recuperator

(c) Fig. 3. Variation of Brayton cycles with the number of intercoolers. (a) No intercooler, (b) one intercooler, (c) two intercoolers.

to cycle efficiency becomes less and less; use of more than three stages with two intercoolers can not be justified economically. Hence, comparison of cycle efficiency is done in this study for one and two compression stages between He and CO2 in closed gas-turbine direct cycles. It is difficult to apply the two-intercooler cycle to the He cycle because gas turbine rotor length is much longer in a He cycle than in a CO2 cycle; rotor dynamics design becomes much more difficult if another compressor is added. Cycle configurations of Fig. 3(a) and (b) were applied to actual designs of He cycle HTGRs (Muto, 2000; Kumar et al., 2001; Baydakov et al., 2001). Cycle efficiencies in these usual Brayton cycle configurations are almost identical between He and CO2 cycles as given in Table 3, although the configurations

are not preferable for the CO2 cycle, as explained hereafter. Specific heat at constant pressure Cp of CO2 is dependent upon pressure and temperature, whereas Cp of He is constant and Cp of CO2 in the recuperator is considerably lower in the low-pressure high-temperature side (connected to the turbine outlet) than in the high-pressure low-temperature side (connected to the compressor outlet). Consequently, CO2 can not be pre-heated to such temperature at the core inlet because it provides maximum cycle efficiency. If flow is by-passed to the compressor before pre-cooling, as shown in Fig. 4(a), this temperature mismatch problem is avoided as schematically explained in Fig. 5. Hereafter, a cycle with a bypass flow cycle is called a “partial pre-cooling” cycle. Partial pre-cooling is also achieved in the

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Y. Kato et al. / Nuclear Engineering and Design 230 (2004) 195–207

Table 3 Cycle thermal efficiencies of He and CO2 cycles at reactor outlet temperature of 800 ◦ C Cycle thermal efficiency (%) Items

No partial pre-cooling No partial pre-cooling Partial pre-cooling

One intercooler

Two intercoolers

7 MPa

12.5 MPa

7 MPa

12.5 MPa

7 MPa

12.5 MPa

45.3 45.8 49.2

45.1 45.5 49.7

47.5 47.8 51.4

47.4 47.4 51.9

48.4 48.7 –

48.3 48.1 –

High Pressure Compressor Bypass Low Pressure Compressor Compressor

Low Pressure Bypass Turbine Compressor Compressor

Reactor

Reactor Pre-Cooler

Pre-Cooler

Generator

Recuperator

Turbine

Generator

Intercooler

He CO2

No intercooler

Recuperator

(a)

(b)

Fig. 4. Partial pre-cooling in respective cycles. (a) No intercooler, (b) one intercooler.

Temp.

Reactor

Turbine

A E High pressure & low temp. side D Low pressure & high temp. side C

B Pre-cooler

Entropy T- s Diagram in a recuperator

Larger Cp value at low pressure & high temperature side results in temperature mismatch (∆ T C-D < ∆ T A-B ) not maximizing cycle efficiency .

Bypassing gas flow from B to E before a pre-cooler leads to ∆ T C-D = ∆T A-B improving

cycle

efficiency

by

about 4%-6% .

Fig. 5. Bypass flow to eliminate temperature mismatch problem in CO2 cycles. (A) Recuperator inlet of the low-pressure and high-temperature side, (B) recuperator outlet of the low-pressure and high-temperature side, (C) recuperator inlet of the high-pressure and low-temperature side, (D) and (E) recuperator outlet of the high-pressure and low-temperature side.

Y. Kato et al. / Nuclear Engineering and Design 230 (2004) 195–207

201

60 Reactor Outlet Pressure 20.0MPa 15.0MPa 10.0MPa 5.0MPa

Cycle Efficiency (%)

CO2 (827˚C)

50

He (827˚C)

40

CO2 (527˚C )

30 He (527˚C)

Pre-Cooler Outlet Temperature : 35˚C Compressor, Turbine Efficiency : 90% Effectiveness of Recuperator : 95%

20 0

0.1

0.2

0.3

0.4

Bypass Flow Ratio (-) Fig. 6. Cycle efficiency change with bypass flow fraction.

65

60

55

Cycle Efficiency (%)

one-intercooler cycle by bypassing flow to the third compressor before the pre-cooler, as shown in Fig. 4(b). The optimum bypass flow fraction depends on turbine inlet temperature and pressure, as shown in Fig. 6. As expected from the constant Cp of He, bypass flow is not necessary for the He cycle, but it degrades cycle efficiency. Bypass flow reduces heat rejected to cooling water through the pre-cooler and increases compressor work. Taking account of the opposing effect on cycle efficiency resulted from the bypass flow, cycle efficiency is improved by about 4% at 800 ◦ C, as shown in Table 3. Fig. 7 shows that cycle efficiency of the CO2 partial pre-cooling cycle is about 3–9% higher than that of the conventional He Brayton cycle, depending on turbine inlet temperature and pressure. At medium core outlet temperature of 650 ◦ C and turbine inlet pressure of 7 MPa, the CO2 cycle achieves cycle efficiency of 45.8%. This cycle efficiency value is comparable with that of a typical He cycle of GT-MHR (47.7%) at high temperature of 850 ◦ C and 7 MPa, as shown in Fig. 8. Its higher efficiency is ascribed to reduced compression work around the critical point of CO2 as explained below and consideration of variation in CO2 specific heat with pressure and temperature. Typical plant data are given in Fig. 9.

P re-C ooler T em p eratu re : 35˚C C om p ressor E fficien cy : 90% T u rb in e E fficien cy : 90% R ecu p erator E ffectiven ess: 95%

C O 2 P artial P re-C oolin g C ycle

50 He Cycle

45 Reactor Pressure

40

20.0 M P a 15.0 M P a 10.0 M P a 5.0 M P a

35

30 500

600

700

800

900

1000

1100

Reactor Outlet Temperature (˚C) Fig. 7. Cycle efficiencies of a CO2 partial pre-cooling cycle and a He cycle in a one-intercooler system.

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Cycle Thermal Efficiency (%)

60 CO2 Gas Turbine Cycle (Direct)

55 MTGR Partial Pre-Cooling Cycle (650˚C, 45.8%)

50

HTGR GT-MHR (850˚C, 47.6%)

45 Water/Steam Cycle (Indirect)

40

HTGR Fort St. Vrain (538˚C, 40.6%)

35

He Gas Turbine Cycle (Direct)

LWR (Avg. 278˚C, about 34%)

30 200

400

600

800

1000

Turbine Inlet Temperature (˚C) Fig. 8. Comparison of cycle efficiencies among MTGRs, direct cycle HTGRs and an indirect cycle HTGR.

Work W of one mole of real gas in isentropic expansion and compression processes is calculated as   dP W = − V dP = − zRT , P

T: 99.7 P: 7.17 H: 827.7

T: 35.0 P: 3.41 H: 783.7

T: 480.2 P: 7.11 H: 1265.3 T: 35.0 P: 2.02 H: 798.9 T: 79.5 P: 3.43 H: 832.4 T: 107.2 P: 2.04 H: 870.0

Intercooler

Reactor

PV = zRT,

T: Temperature (˚C) P: Pressure (MPa) H: Specific Enthalpy (kJ/kg)

High Bypass Low Pressure Pressure Compressor Turbine CompressorCompressor Generator

Pre-Cooler

T: 650.0 P: 7.00 H: 1470.2

where V = gas volume, P = gas pressure. For real gases, an equation of state is written as

Recuperator T: 243.9 P: 2.07 H: 1009.2

T: 502.2 P: 2.10 H: 1296.6

T: 229.5 P: 7.14 H: 977.9 Reactor Outlet Temperature: 650˚C Reactor Outlet Pressure: 7.0 MPa Pre-Cooler & Intercooler Outlet Temperature: 35˚C Turbine & Compressor Efficiency: 90% Recuperator Effectiveness: 95% Bypass Flow Fraction: 7.3% ↓ Cycle Efficiency: 45.8%

Fig. 9. Plant data of medium temperature CO2 partial pre-cooling cycle with one intercooler.

Y. Kato et al. / Nuclear Engineering and Design 230 (2004) 195–207

where R = gas constant, and z = compressibility factor. The compressibility factor represents the departure from the ideal gas and is defined by z=

At the critical temperature and pressure, the z value dips sharply below the ideal line of unity and takes an extremely low value as low as about 0.2 as shown in Fig. 10. Factor z gives fractional deviation of real gas from the ideal gas (z = 1) in isentropic expansion and compression processes. A low z value indicates that the gas is more compressible than the ideal gas. Compressor inlet temperature is usually around 35 ◦ C in gas cooled reactors which is determined by available cooling water temperature. Since compressor inlet temperature is very close to the CO2 critical temperature (31.0 ◦ C), z values of CO2 in the compressing condition are estimated to be smaller than those of He, as seen from critical parameters. Both CO2 and He z values approach unity at high-temperature and high-pressure conditions in the turbine, so the difference is small. Judging from z values in compression and expansion, CO2 cycle efficiency is expected to be higher than that of the He cycle. Cycle efficiency is dependent on the lowest temperature that appears usually at the inlet of compressors in a Brayton cycle. Cycle efficiencies are plotted against the compressor inlet temperature in Fig. 11, increasing linearly with temperature decrement at the rate of 1.3%/10 ◦ C. Corrosion resistant structural materials and reliable components in the CO2 environment can be used at

RT . PV

If P, V and T are expressed in terms of the respective reduced pressure (Pred = P/Pcrit , Pcrit is critical pressure), reduced volume Vred (Vred = V /Vcrit , Vcrit is critical volume) and reduced temperature Tred (Tred = T /Tcrit , Tcrit is critical temperature), the above equation is rewritten as 

Pred Vred Pcrit Vcrit . z= R Vcrit Tred The term of (Pcrit Vcrit /RVcrit ) is known to be approximately constant for many gases; for that reason z appears to be a universal function of Pred and Tred because of the law of corresponding states. If z is plotted against Pred and Tred , a single curve will be obtained for all gases. Gases with an equal z value display the same behavior according to the law of corresponding states. The compressibility factor z at various Tred is plotted against Pred shown in Fig. 10 using the data given by Hougen et al. (1960). At these pressure and temperature conditions, the fit is good to within about 1% for gases (Moore, 1972).

1.6 Reduced Temperature T red =1.0

1.1 1.2 1.6 2 1.4 3 4

1.4

Compressibility Factor z

203

6

1.2

8

10

15

1.0 2

He Compression Condition

1.6 1.4

0.8 0.6

Higher cycle efficiency could be attained in CO2 cycles compared with He cycles by utilizing reduced compression work around the critical point.

1.2

0.4

1.1

0.2

Drawn from the data in O.A.Hougen, et al., "Chemical Process Principles,Part II, Thermodynamics", John Wiley & Sons

1.0

0.0 0

2

4

6

8

10

12

14

16

18

20

22

24

Reduced Pressure P red Fig. 10. Compressibility factor with reduced temperature and reduced pressure.

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Y. Kato et al. / Nuclear Engineering and Design 230 (2004) 195–207

Cycle Efficiency (%)

56 54 52 Core Outlet Temperature = 800˚C

50 48 46

Core Outlet Temperature = 650˚C

44 42

25

30

35

40

System Lowest Temperature (˚C) Fig. 11. System lowest temperature dependency on cycle efficiency in CO2 partial pre-cooling cycle with one intercooler.

the medium temperature of 650 ◦ C, as has been proven in AGRs during extensive operation with core outlet temperature of 650 ◦ C. The AGR graphite moderator was prevented from excessive oxidation reaction with CO2 by adding methane and carbon monoxide added to CO2 in a range of small concentrations in CO2 (Hewitt and Collier, 1997). The average plant availability of AGRs over the last 10 years is as high as that of current LWRs (Knox, 1999). Regarding anti-oxidation of the graphite, surface coating with inert materials would also be promising: it has been developed for oxidation prevention of cutting tools up to 1000 ◦ C in air atmosphere. Graphite

balls coated on surface with TiSiN, TiCN + TiN and TiCN + Al2 O3 + TiN are shown in Fig. 12. Fracture cross section SEM micrographs of TiCN + TiN coating on the graphite ball are shown in Fig. 13. Their irradiation testing in the JMTR (Japan Material Testing Reactor) core has been planned. Any type of fuel can be applicable to CO2 gas turbine cycles with partial pre-cooling such as a conventional metal cladding fuel of the type used for LWRs and LMFRs, or a pebble bed or block fuel of the type used for HTGRs. In the case of the conventional fuel type encapsulated with type 316 stainless steel cladding material, encapsulation completely (except in the case of cladding failure) prevents release of fission products (FPs) from the fuel to coolant and consequent contamination of turbomachinery; this reduces radiation dosage incurred through maintenance. Leakage of a small amount of FPs is unavoidable in the case of particle fuel for the pebble and block type fuel coated with porous carbon, pyro-carbon, and SiC. The leak fraction of FPs to total quantity produced in the fuel depends on their diffusion coefficients in pyro-carbon and SiC. Recent irradiation experiments (IAEA, 1997) for the coated particle fuel show that diffusion coefficients of typical FP elements such as cesium, strontium, silver, and iodine are evaluated from Arrhenius plots to be lower by about two orders of magnitude in the medium temperature of 650 ◦ C than at the high temperature of 850 ◦ C. These results may lead to considerable reduction of radiation dosage

Fig. 12. Graphite balls with coated surfaces.

Y. Kato et al. / Nuclear Engineering and Design 230 (2004) 195–207

205

Fig. 13. Fracture cross section SEM micrographs of TiCN + TiN coating on graphite ball surfaces.

during maintenance of turbomachinery and heat exchangers. Various heat-resistant materials have been used for HTGRs, as shown in Table 4: Hastelloy XR, Alloy 800H, and 2.25 Cr–1 Mo steel. Instead of these materials, less-expensive materials, such as type 316 stainless steel, can be applied to a medium temperature

reactor which has been proven to be compatible with CO2 up to medium temperature of 650 ◦ C in AGRs, and used for LWRs and LMFRs for long periods. Thus, the medium temperature offers more flexibility for structural material choice. Lowering temperature to 650 ◦ C not only eases maintenance through the lower diffusion leak rate of fission products from coated

Fig. 14. Bird’s-eye view of a CO2 gas turbine reactor.

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Y. Kato et al. / Nuclear Engineering and Design 230 (2004) 195–207

Table 4 Typical material selection in HTGRs and MTGRs Positions

Core barrel, control rod drive mechanism Core support plate Pressure vessel liner

Typical materials HTGRs (850 ◦ C)

MTGRs (650 ◦ C)

Alloy 800H

316 SS (proven in AGR)

2.25 Cr–1 Mo steel Hastelloy XR

particle fuel, but also provides flexibility in choosing materials. All components (core, turbine, recuperator, pump, and pre-cooler) are installed into pressure vessels as typically illustrated in Fig. 14. Number (N) of turbine stages is given as N = Cp


T ,
hstage

where Cp = specific heat at constant pressure,
T = temperature drop in a turbine,
hstage = enthalpy drop per turbine stage. It is seen from the above equation that the number of the turbine stages (or turbine length) is proportional to the Cp value at the same
T and
hstage conditions. Values of
T and
hstage are determined by optimization of the cycle design and the turbine diameter, respectively. Because the specific heat value of CO2 at constant pressure is smaller by a factor of about five

He

than that of He, number of stages (or length of turbine) is smaller at the same turbine diameter and then volume (or approximately weight) of a CO2 cycle is less by a factor of about five than that of a He cycle. We did a comparative design study of turbomachinery (turbine and compressor) between He and CO2 cycles optimizing the turbine length and diameter. The optimization results shown in Fig. 15 indicate that turbine length is 44% [= (0.81/1.84) × 100] and turbine diameter is 65% [= (1.25 + 1.33)/(1.94 + 2.01) × 100] in a CO2 cycle compared with those of a He cycle. Consequently, the CO2 cycle system offers one-fifth (19%) volume (or approximate weight) of CO2 cycle reference to that of He cycle (Muto et al., 2003). Judging from the key component size and cycle efficiency, power generation cost per unit electricity for the CO2 cycle design might be less than that for a He cycle design. A gas turbine is a much simpler system than a steam turbine because it has no moisture separators, steam extraction systems, or coolant purity control systems required for steam turbine cycles. Furthermore, it has much smaller size because of its lower turbine pressure ratio of about four. Our proposed direct cycle has only a single coolant circuit because of its direct cycle, whereas PWRs have primary and secondary cooling circuits with huge steam generators because of their indirect cycle. Consequently, our CO2 gas turbine direct cycle reactor is much simpler and smaller than PWRs.

CO2

1.84m

0.81m

1.94m

1.80m

2.01m

1.25m

1.10m 1.00m

Fig. 15. Comparison of turbine size between He and CO2 cycles.

1.33m

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