An Alternative Regeneration for Gas Turbines

8/6/2019 An Alternative Regeneration for Gas Turbines

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An Alternative Regeneration For Gas Turbines

Dept. of Mechanical Engg. N.S.S.C.E, Palakkad

ACKNOWLEDGEMENT

I would like to express my sincere gratitude and indebtness to Mr.D.PRAKASH,

Asst. Professor, Department of Mechanical Engineering, for his valuable guidance and

prompt advices in preparation and presentation of this seminar.

I am also indebted to Mr.V.P.SUKUMARAN NAIR, Professor and Head of Mechanical

Engineering Department for his valuable support and encouragement in presenting this

seminar.

Finally I also extend my sincere thanks to all the staff members of Mechanical

Engineering Department and all my friends for their co-operation and support.


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SYNOPSIS

An alternative configuration for a regenerative gas turbine engine cycle is

presented that yields higher cycle efficiencies than either simple or conventional

regenerative cycles operating under the same conditions. The essence of the scheme is to

preheat compressor discharge air with high temperature combustion gases before the

latter are fully expanded across the turbine. The efficiency is improved because air enters

the compressor at a higher temperature, and hence heat addition in the combustor occurs

at a higher average temperature. The heat exchanger operating conditions are more

demanding than for a conventional regeneration configuration, but well with in the

capability of modern heat exchangers. Models of cycle performance exhibit several

percentage improvement relative to either simple cycles or conventional regenerationschemes. The peak efficiencies of the alternative regeneration configuration occur at

optimum pressure ratios that are significantly lower than those required for simple cycle.

Model calculations for a wide range of parameters are presented, as are comparisons with

simple and conventional regeneration cycles.


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BACKGROUND

In recent years, ground-based gas turbine engine (GTE) applications have been

appreciably expanded due to significant improvements in cycle efficiency. Simple cycleefficiencies of over 40 percent are now possible from some designs, making GTEs

competitive alternatives to Diesel engines and Rankine steam cycles. Most ground-based

GTE applications can accommodate the space and mass requirements associated with

adding regeneration to a simple cycle, with the goal of even higher cycle efficiencies.

For many operating conditions, regenerators (heat exchangers) can improve

found- based GTE performance by recovering heat from high temperature exhaust gases.

Numerous applications for the recovered heat have been devised, including combined

cycle and cogeneration applications, but on stand alone GTE cycles the recovered heat is

usually used for preheating the air passing between compressor and combustor. In this

way, a well-known goal of thermodynamic design is satisfied by increasing the average

temperature at which heat is added to the air during combustion resulting in increased

efficiency. Regenerators have traditionally ([1-3]) used product gases leaving the final

turbine stage as the source of heat (referred to herein as conventional regeneration) so

that the maximum amount of work is extracted from the high-enhalpy gas stream before

any heat is recovered. How ever, such a regenerator location is inconsistent with a

fundamental lesson from Carnot-cycle thermodynamics,which is that cycle efficiency is

maximized by increasing the average temperature at which heat is added, and not

necessarily by maximizing the work output. Thus, the overall efficiency of conventional

regenerative GTE cycles can be improved through an alternative regenerators location,

and to the authors knowledge, this paper is the first discussion of such schemes.

Consider a GTE configuration with a high-pressure turbine (HPT) and a power

turbine (PT). If a heat exchanger is located between the two turbines as shown in Fig.1

then the cycle efficiency can be substantially improved beyond that available from the

conventional regeneration configuration. The thermodynamic effect is to increase the

amount of heat that is delivered to the compressed air beyond what conventional

regeneration is able to achieve, resulting in a higher average temperature for the heat

addition proves in the combustor. Although there is less work produced by the PT in the


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alternative regeneration scheme due to decreases in pressure and temperature of the gas

as it passes through the regenerator, the cycle efficiency is improved and the lower

specific work output can be compensated for by using larger engine components.

Fig. 1 Schematic of Alternative Regeneration Cycle


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MODEL DEVELOPMENT

Computer models of simple, conventional regenerative, and alternative

regenerative cycles were developed to examine the influence of various parameters on the

performance of the cycles. The primary goal in developing the models was to

demonstrate the enhanced performance of the proposed alternative regeneration

scheme, and consequently, the models were not comprehensive in including all details of

gas turbine engine performance. An economic analysis was beyond the scope of this

paper, and would be difficult to implement in a general way for the wide variety of gas

turbine engine applications that exist today. However, the case of continuous duty power

generation is noteworthy since fuel costs over the lifetime of the plant are typically sohigh that is commonly cost-effective to invest capital to improve cycle efficiency by even

one percent. The following discussion will show that the alternative regeneration scheme

has the potential to improve cycle efficiency by several percentage points in some

scenarios.

Fig. 2 Schematic of a Simple Cycle


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Output by the cycle is proportional to the temperature drop across the PT. The

simple cycle efficiency can be written as

Net work input

Cycle = External heat input

mcpgT(T6-T7s) = T(T6-T7s) T(T6-T7s)

= = (1)

mcpg (T4-T2) (T4-T2)

The S subscript in Eq.(1) refers to the temperature that is computed when an isentropic

expansion is assumed:

T7s = (P7 )g/g-1

(2)

T6 (P6)

For the simple cycle, the pressure at state 7 is assumed to be atmosphere and the

temperature at state 6 is determined by writing an energy balance that equates the work

input at the compressor to the work output by the HPT.

mcpa (T2-T1) = mcpg (T4-T6) (3)

The simple cycle efficiency computed by Eq.(1) is used here as a comparison case

for the regenerative cycle configurations. It is important to note that slightly lower values

for simple cycle efficiency will be computed if the cycle is modeled as having only a

single turbine that provides both the compressor work and net shaft work. Single and

twin-turbine models give the same efficiency only when the turbines are modeled as

isentropic. This modeling artifact also arises when modeling non-isentropic Rankine

cycle turbines or multistage non isentropic compression. There are no references known

to the author that discuss this thermodynamic anomaly, both the calculation is so simple

that it presumably has not warranted attention previously. The alternative regeneration

cycle that is the focus of the present paper is modeled with a non isentropic, twin-turbine

configuration; hence it is more appropriate to compare those results to the larger values of

simple cycle efficiency that are computed for the twin-turbine configuration by using

Eq.(1).


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Consistent with the prior comments, a two-turbine model of a conventional

regeneration cycle was developed. In the conventional regeneration model, heat is added

to the cycle between states 2 and 3 as in fig. 2a, and that heat is extracted from the

exhaust gas stream leaving the PT at state 7.

Fig. 2a Schematic of Conventional Regeneration Cycle

The resulting expression for cycle efficiency is given by

mcpgT(T6-T7s)

Cycle (4)mcpg (T4-T3)

State 6 is evaluated by an energy balance just as it was for the simple cycle. Temperature

7S is again computed with Eq.(2), but the pressure at state 7 will be larger than

atmosphere pressure by an amount equal to the pressure drop through the regenerator.

The temperature at state 3 depends on the effectiveness of the regenerator in transferring

heat from the exhaust gases to the compressed air stream, and can be evaluated from the

regenerator effectiveness, defined as

Actual heat transferred T3-T2

n = = (5)Maximum possible heat transferred T7-T2

In.Eq.(5), the heat exchanger effectiveness is defined for the fluid with the smaller

product of mass flow rate and specified heat, which corresponds here to the compressor

discharge air since it has the smaller specific heat. The increase in working fluid mass


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due to fuel addition at the combustor was ignored in the present models, but in real engine

flows the mass flow rate of the air would be slightly less than that of the product gases,

again supporting the idea that the minimum fluid for purposes of defining the heat

exchanger effectiveness would be the compressed air stream between states 2 and 3. In

Eq(5), the temperature T7 is the highest possible temperature that could be obtained by the

air passing to the cobustor, and this temperature could only be realized at state 3 if the

heat exchanger had negligible heat transfer resistance or infinite surface area.

From the states identified in Fig.1, and expression for the overall cycle efficiency

of the alternative regenerative cycle can be written as

cpgT(T6-T7s) = cpgT(T4-T5s) - cpa(T2S-T1)/ c

cycle = (6)

cpg (T4-T3)

The second and third terms in the numerator of Eq.(6) cancel one another if the

compressor work is supplied entirely by the HPT, but Eq.(6) is written in its full form

because later discussion will consider a second approach to providing the compressor

work.

The models for all the three cycles assume that air is the working fluid between

compressor and combustor inlets (cpa = 1.005 kJ/kgoC, a = 1.4), but that beyond the

combustor inlet the chemical reaction and increased temperature alter the gas properties

([1)] so that they are better represented by cpg = 1.147 kJ/kgoC and g = 1.33. All

calculation further assumed the isentropic compressor efficiency was 86 percent, the

isentropic turbine efficiencies were 89percent, the combustor pressure drop was 1.3.8 kPa

(2 psi), and that the compressor inlet conditions at state I were 21oC(70oF) and 101.4kPa

(14.7 psia). In addition, a reference case was established with moderate values estimated

for the remaining parameters used in the calculations. The reference case assumed that

the regenerator effectiveness was 70 percent, that the turbine inlet temperature was

1100oC (2011

oF), and that pressure drops associated with each pass through the heat

exchanger were 13.8 kPa (2 psi) each. In the following discussion, the parameters used in

the calculations were those associated with the reference case, unless otherwise specified.


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DISCUSSION

Conventional regeneration offers the benefit of improved cycle efficiency over

simple cycles for the ideal case where there is no pressure drop through the regenerator.

For example, for the reference case conditions, except with co pressure drops across the

regenerator, a conventional regeneration cycle achieves cycles efficiencies of 43.0 percent

at an optimum pressure ratio (PR) of 8, compared to the simple cycles efficiency of 42.7

percent at an optimum, and fairly high Proof 37 (this small benefit of conventional

regeneration improves as effectiveness increases eg., for an effectiveness of 90 percent,

the efficiency improves to 50.4 percent at a PR of 4). In addition to the higher cycle

efficiency of the conventional regenerative cycle, the lower optimum pressure ratio is

attractive because the compressor requirements are less demanding. The alternative

regeneration cycle represented in Fig.1, with no regenerator pressure drops, achieve a

peak efficiency of 45.9 percent at an optimum pressure ratio of 16, which is a substantial

improvement over conventional regeneration.

Even small pressure drops through the regenerator, like those specified in the

reference case, take a large toll on the performance of the conventional regenerative

cycle, as shown in fig.3.

Fig 3


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In fact, for modest regenerator pressure drops of 13.8 kPa (2psi), the performance of the

conventional regenerative cycle is inferior to that of the simple cycle, illustrating one

reason why conventional regeneration is frequently unsuitable for use on ground based

engines. Figure 3 shows that the peak efficiency of the alternative regeneration cycle

(44.6 percent), with the regenerator pressure losses, is superior to either of the other two

cycles, and this peak again occurs at the modest pressure ration of 16.

Some GTE applications might impose severe space limitations on regenerator

size, resulting in lower effectiveness or increased pressure drops. Figure 4 shows the

effect of these two parameters on the performance of the cycles, where each point on the

curve has been determined at the optimum pressure ratio for the particular operating

condition. For the conventional regeneration cycle having heat exchanger effectiveness

of 70 percent, there is actually a performance penalty when using a regenerative cycles

are compared at an effectiveness of 90 percent, it can be seen that the two curves are not

parallel, but that the conventional regenerative cycle is degraded more abruptly than the

alternative regenerative cycle as pressure drop through the heat exchanger increases. For

the alternative regeneration cycle,

Fig 4

there are various combinations of pressure drop and effectiveness that result in

performance superior to the simple cycle. Finally, it is interesting to note that the

optimum pressure ratio for the alternative regeneration scheme is only a very weak


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function of regenerator pressure drop. For example, for an effectiveness of 0.7, the

optimum pressure ratio for the alternative regeneration cycle varies from 16 (Preg = 0) to

18(Preg = 55kPa).

It is well known that the maximum cycle temperature has a large effect on overall

efficiency, and this is demonstrated in fig.5. In addition to the expected trends, Fig.5

shows two important results. For the reference case pressure drops, the alternative

regeneration cycle is superior to the other two for any turbine inlet temperature, and the

conventional regenerative cycle performance falls further behind the that of the other two

cycles as turbine inlet temperature increases. However, the simple cycle results shown in

Fig.5 could be misleading because they imply that a simple cycle could be useful at the

higher turbine inlet temperatures, but the optimum pressure ratios required to achieve the

efficiencies at higher turbine inlet temperature become excessive for a practical design.

For example, the simple cycle requires optimum pressure ratios of about 37.58, and 90 for

turbine inlet temperature of 1100oC, 1300

oC, and 1500

oC, respectively, by contrast, the

optimum pressure ratio of 30 for the alternative regeneration cycle operated at 1500oC is

feasible with current compressor designs, and results in a cycle efficiency of 54.2 percent.

Fig. 5

For engines that run continuously, cycle efficiency is likely to be the most

important criterion used in designing the cycle because fuel costs over the lifetime of the

engine will far exceed the initial capital costs, However, for some applications, the lowest

specific cost or smallest engine size may be the more important criteria [(3)]. The


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physical size and cost of an engine are directly related to the specific power output, and

Fig.6 shows that there is a penalty associated with the alternative regeneration scheme in

this regard. To be consistent with this papers theme of improved cycle efficiency, the

operating points in Fig.6 were determined by finding the pressure ratios that gave the

highest efficiencies (alternatively, operating points could have been determined by

finding the pressure ratios that resulted in maximum specific work output if that were the

dominant concern driving engine design). For effectiveness between 52 percent 75

percent, the alternative regeneration scheme has efficiency and specific work equal to, or

better than the simple cycle. For effectiveness greater than 75 percent, the cycle

efficiency climbs sharply, but the specific work decreases below that of the simple cycle.

The conventional regeneration cycle has specific work output superior to that of the

simple cycle for any effectiveness, but the efficiency is inferior for effectiveness less than

82 percent. Also, for a particular effectiveness the conventional regeneration cycle has

better specific work output than the alternative regeneration cycle, but its cycle efficiency

is 3.5 to 5.5 percentage points lower, depending on the particular conditions. Since a heat

exchanger increases the overall size and cost of an engine, the Fig.6 data would have to

be weighted carefully for a particular application to determine which cycle would be

preferable, especially for a space-limited or low cost application where these

characteristics are more important than cycle efficiency. To summarize, for the reference

case effectiveness of 0.7, the cycle incorporating conventional regeneration yields about

13 percent more specific work output than the alternative regeneration cycle, suggesting

that the engine components could be about 13 percent smaller than an engine utilizing the

alternative regeneration scheme for a given power requirement, but the cycle with

conventional regeneration would have much lower efficiency (39.7 percent versus 44.6

percent).

In order to design a suitable regenerator, the magnitude of the heat load in the

device must be known. In the alternative regeneration scheme, the air preheating

operation proceeds to a higher temperature than in conventional regeneration, but the hot

side of the heat exchanger utilizes higher temperature gases to do the heating so that the

overall surface area in the heat exchanger can be comparable to that in a conventional

regeneration cycle.


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Fig. 6

As one measure of heat exchanger size, a regenerator heat ratio was defined by

Regenerator heat ratio

heat transferred in regenerator

=

network output from cycle

= cpa(T3-T2 ) (7)

cpg(T6-T7 )+ cpg(T4-T5 ) - cpa(T2-T1)

For the reference case conditions, conventional regeneration results in a regenerator heat

ratio of about 0.77, while the alternative regeneration cycle operated in the gas-generator

configuration has a ratio of about 0.91, indicating that the heat exchanger would have to

be somewhat larger to transfer about 18 percent more heat in the latter case.


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OPTIMIZING PERFORMANCE-SINGLE SHAFT CONFIGURATION

Figure I depicts the purpose of the HPT as providing power input to the

compressor, consistent with many GTE configurations that have a gas generator (ie.

Compressor, combustor, and HPT) and a power turbine. There are certain advantages to

this arrangement, including the ability to operate the two turbines at different speeds.

However, there is no thermodynamic reason why the optimum performance of the

alternative regeneration cycle should correspond to this particular hardware configuration,

and in fact, results discussed below will show that the best overall cycle efficiency

usually occurs when the temperature drop across the HPT, and hence the HPT work

output, is less than that required to drive the compressor. Thus, some work from the PT

would also have to be directed to the compressor in the optimum efficiency, or single

shaft scenario. Figure 7 shows how the efficiency varies as a function of HPT outlet

pressure.

Fig. 7

A vertical dashed line on Fig.7 indicates the HPT outlet pressure when the cycle

configuration is that of a gas generator with separate power turbine. HPT outlet pressures

to the left of the dashed line correspond to cases where the HPT is supplying all of the

compressor work and some net shaft work. Points to the right of the dashed line indicate


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that the HPT and PT are working together to supply the compressor work. The cycle

efficiency peaks at an HPT outlet pressure slightly above that which would be required if

the HPT alone drove the compressor, and this is the usual case for the range of parameters

considered here. In fig.7 the cycle efficiency increases from the reference case value of

44.6 percent to a maximum value of 45.3 percent when the PT is utilized to provide 26

percent of the compressor work requirement. The optimum cycle pressure ratio increases

from 16 in the gas-generator configuration to 20 in the single-shaft case.

One important concern with the alternative regeneration scheme is the temperature

experienced by the materials in the heat exchanger itself. Since the air preheating in the

alternative regeneration scheme occurs at higher pressures and temperatures than in

conventional regeneration, the heat exchanger requirements are more severe. However,

for most of the operating conditions presented herein, the maximum heat exchanger

temperatures (i.e., the peak temperatures at state 5) were in the range 700-900oC.

Exceptions occurred for cases with higher turbine inlet temperatures and higher

effectivenesses, where HPT outlet temperatures as high as 1100oC were computed for the

single-shaft configurations. Modern gas/gas heat exchangers are capable of temperatures

as high as 1100oC and pressures as high as 30 atmospheres ([4]), so the alternative

regeneration scheme does not appear to pose any insurmountable problems in this regard.

Figure 7 shows that for the optimum HPT outlet pressure and reference case conditions,

the peak regenerator temperatures would be about 800 oC. For contrast, if the HPT is

used to provide all the compressor work in a gas-generator configuration, then the peak

temperature in the regenerator is only about 690 oC for the reference case conditions.

For consistency with other calculations presented herein, Fig.7 was generated for

the reference case conditions. However, fig.7 is somewhat misleading because the

improvement in cycle efficiency when using the single-shaft configuration is much more

significant when considering lower-technology regenerators having low values of

effectiveness and/or large pressure drops. Figure 8 demonstrates how the single-shaft

configuration is superior to the gas-generator configuration, as a function of regenerator

performance parameters.


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Fig. 8

As with previous figures, the overall cycle pressure ratio has been optimized to

achieve the highest cycle efficiency for each point on the curve. The fig.8 data

correspond to optimum PRs of between 8 and 33, with the larger values required for the

lower effectivenesses. Figure 8 shows that the single-shaft configuration effectively

decreases the slope of the efficiency curves, resulting in improved performance,

especially for larger regenerator pressure drops. Consequently, there are more

combinations of regenerator pressure drop and effectiveness that result in performance

superior to the simple cycle. The maximum regenerator temperatures required for the

single-shaft operating points shown in Fig.8 lie between 760oC and 850

oC, with the

higher temperatures required for the higher effectiveness.

For the single-shaft configuration, the peak in cycle efficiency is fairly flat and

incentive to overall cycle pressure ratio, as shown in Fig.9. For each value of overall PR

in fig.9, the cycle efficiency has been determined at the optimum HPT outlet pressure.

Because the efficiency curve is fairly flat near its peak. It is conceivable that certain

compromises might be attractive when designing a cycles operating point. For example,

fig.9 shows that if a designer were willing to use a larger compressor to increase the

pressure ratio from the optimum value of 20 to a value of 30, then the cycle efficiency


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would drop from 45.3 percent to 44.9 percent, and the regenerator heat ratio would

decrease by 30 percent. A 30 percent decrease in the regenerator heat ratio implies a

corresponding reduction in size of the heat exchanger that is often the bulkiest component

of a regenerative GTE. Hence the larger compressor and small reduction in efficiency

might represent tolerable design compromise for a compact engine where overall size is a

critical issue. Increasing the turbine inlet temperature obviously results in improved cycle

efficiency, but the regenerator requirements become more severe at the same time.

Fig. 9

Earlier discussion pointed out that the over all pressure ratio for the gas-

generator configuration was feasible even for turbine inlet temperatures of 1500oC,

whereas the simple cycle at the same temperatures requires PRs much higher than modern

compressors can supply. The maximum regenerator pressure is the same as the

compressor outlet pressure, so the PRs required for the gas-generator configuration (30)

are feasible in the heat exchanger as well. However, the principle drawback of the single-

shaft configuration is that optimum PR increases relative to the values required for thegas-generator configuration, with negative consequences on both compressor and heat

exchanger requirements.The chief attraction to the single shaft configuration appears to

be in improving cycle efficiency at lower turbine inlet temperatures where optimum cycle

pressure ratios are more modest.


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CONCLUSIONS

An alternative configuration for a regenerative GTE cycle with numerous

favourable operating characteristics is discussed. For practical ranges of operating

parameters, the alternative configuration always results in a cycle efficiency superior to

either a conventional regenerative cycle or a simple cycle. This performance

improvement is robust and not limited to a narrow range of operating conditions or

component efficiencies. Although the demands on the heat exchanger are severe, the

regenerator temperatures and pressures are well below the limits of existing heat

exchanger designs. The alternative regeneration scheme is particularly attractive at high

turbine inlet temperatures. For turbine inlet temperatures as high as 1500oC, optimum

PRs are only 30, whereas for the same conditions the optimum pressure ratio of a simple

cycle is excessive (>40) for temperatures larger than 1115oC. When a power turbine and

gas generator can be configured on the same shaft, operating at the same speed, then the

alternative regeneration cycle efficiency can be improved even further and this situation is

particularly useful if the heat exchanger is limited by low effectiveness or large pressure

drops.


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NOMENCLATURE

Cpa = specific heat of air

Cpg = specific heat of product gases

Cyc eff = cycle efficiency

HPT = High Pressure Turbine

n = regenerator effectiveness = actual heat transfer/

maximum possible heat transfer

P = working fluid pressure

PT = power turbine

PR = cycle pressure ratio

T = working fluid temperature

Greek

Preg = pressure drop through regenerator

cycle = cycle efficiency

C = isentropic compressor efficiency = isentropic work/actual

work

T = isentropic turbine efficiency = actual work/isentropic work

a = specific heat ratio for air

g = specific heat ratio of product gases

Subscripts

1,2,3. = thermodynamic states identified in Figs. 1 and 2

S = state resulting from an isentropic compression or expansion


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REFERENCES

[1] Cohen, H.,Rogers, G.F.C., and Saravanamuttoo, H.I.H., 1996, Gas Turbine

Theory,4th Ed., Longman Group, Harlow, England.

[2] Bathe, W.W., 1996, Fundamentals of Gas Trubines, 2nd

Ed, John Wiley and Sons,

New York.

[3] Khartchenko, N.V., 1998, Advanced Energy Systems, Taylor and Francis,

Washington, DC.

[4] Wright, I.G., and Stringer, J., 1997, Materials Issues for High-Temperature

Components in Indirectly Fired Cycles, ASME Paper No.97-GT-300.

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An Alternative Regeneration for Gas Turbines