Analysis of a reheat gas turbine cycle with chemical recuperation using Aspen

Pergamon Energy Convers. Mgmt Vol. 38, No. 15-17, pp. 1671-1679, 1997

© 1997 Elsevier Science Ltd. All rights reserved Printed in Great Britain

PII: S0196-8904(96)00208-7 0196-8904/97 $17.00 + o.00

ANALYSIS OF A REHEAT GAS TURBINE CYCLE WITH CHEMICAL RECUPERATION

USING ASPEN

SIMON HARVEY and N'DIAYE KANE Laboratoire d'l~nerg6tique des Syst~mes de Nantes ISITEM, BP 90604, 44306 Nantes Cedex 03, France

Abstract--Gains in the power output and thermal efficiency of industrial gas turbines have occurred in the past, primarily from increased firing temperatures and operating pressures. More recently, there is growing interest in investigating advanced cycle concepts that make use of one or more of the following performance enhancement modifications: compression intercooling, reheat expansion and exhaust heat recovery. Recent attention has focused, in particular, on the chemical heat recovery concept. The "waste" heat in the turbine exhaust is used to convert a methane-steam mixture into a hydrogen-rich fuel in a methane-steam reformer. The potential benefits of such cycles include high conversion efficiency, ultra-low NO, emission levels (less than 1 ppm) and high power density per unit of land. However, such cycles require high turbine exhaust temperatures, which may be achieved effectively by staging the turbine expansion and including a reheat combustor. ABB recently unveiled its new GT26 series stationary gas turbines using staged expansion with reheat combustion, allowing high thermal efficiencies with relatively low turbine inlet temperatures. This type of turbine appears particularly well-suited for chemical heat recovery. In this paper, we present a CRGT cycle based on a reheat gas turbine with key design features similar to those of ABB's GT26 machine. The cycle analysis is performed using Aspen Technology's ASPEN + process simulation software. The paper includes a detailed first and second law analysis of the cycle. © 1997 Elsevier Science Ltd.

Chemical recuperation Reheat gas turbine Aspen Exergy Methane-steam reforming

Cp ex

e~.h=

L = h =

LHV = ,h=

pCb = pRI = P D =

Q= R = S =

S t = T =

TRIT = AT~q =

X :

r]l = ~=

N O M E N C L A T U R E

Molar availability (J mol -~) Specific heat capacity (J kg-~ K --~) Exergy (J kg-~) Chemical exergy (J kg -~) Exergy flow (MW) Enthalpy (J kg-~) Lower heating value (J kg -~) Mass flowrate (kg s -~) Compressor bleed pressure (Pa) Rotor inlet stagnation pressure (Pa) Compressor bleed fractional pressure drop Heat (J) Ideal gas constant (J mol -~ K ~) Entropy (J kg -~ K -~) Stanton number Temperature (K) Turbine rotor inlet temperature (K) Chemical approach to equilibrium (K) mole fraction Blade cooling efficiency Turbine isentropic efficiency Coolant characteristic parameter.

I N T R O D U C T I O N

Chemical recuperation is one of several innovative concepts applicable to natural gas-fired gas turbine-based power generation cycles. The concept has made considerable progress over the past few years w i th the s u p p o r t o f the C a l i f o r n i a E n e r g y C o m m i s s i o n [1]. T h e idea o f i m p r o v i n g h e a t

eng ine p e r f o r m a n c e by us ing a chemica l r eac t i on to r e cove r was te h e a t was first d i scussed in

ECM 38/I 5-17--F- 1671

1672 HARVEY and KANE: ANALYSIS OF A REHEAT GAS TURBINE CYCLE

Ref. [2]. The chemically recuperated gas turbine is an extension of the steam-injected gas turbine concept, in which exhaust heat is used to raise steam which is subsequently injected directly into the gas turbine combustor. Figure 1 illustrates the CRGT concept. Exhaust heat is first recovered in a methane-steam reformer (MSR) followed by a heat recovery steam generator (HRSG). The natural gas fuel is mixed with the generated steam and fed into the MSR. In the reformer, the mixture of natural gas and steam is heated by the combustion turbine exhaust, and an endothermic reaction occurs between the methane and the steam. The reaction requires the presence of a nickel-based catalyst, and results in the production of hydrogen (H2), carbon dioxide (CO2) and carbon monoxide (CO). Complete conversion of methane in this manner could, potentially, increase the effective fuel heating value by approximately 30%. Thus, the methane-steam mixture absorbs heat thermally (as it is heated), and chemically (as the endothermic reaction proceeds), resulting in a larger potential recuperation of exhaust energy than can be obtained by conventional recuperation, which recovers energy by heat alone. The reformed fuel, containing CO, CO2, H2, excess steam and unconverted methane, is then fed into the turbine combuster.

Heat recovery in the steam reforming section of the cycle is facilitated by high turbine exhaust temperatures, and the CRGT concept, therefore, appears particularly well-suited to run in conjunction with a reheat gas turbine. The purpose of this paper is to investigate a chemically recuperated reheat gas turbine cycle.

REHEAT CHEMICALLY RECUPERATED GAS TURBINE CYCLE

The proposed cycle is shown schematically in Fig. 2. The turbine characteristics are similar to those of the ABB GT26 series heavy duty reheat gas turbine, i.e. a pressure ratio of 30:1, a turbine rotor inlet temperature (TIT) of 1508 K, and an air inlet flow of 530 kg/s, as presented in Ref. [3]. Air (stream l) is compressed in the compressor. Turbine blade cooling air is bled from various sections of the compressor to cool both the high pressure turbine blades (stream CA1) and the low pressure turbine blades. Cooling air for the first stage of the low pressure turbine (stream CA2) is intercooled. Cooling air for the other LP turbine stages is ducted direct from the compressor (stream CA3). The compressed air (stream 2) enters the high pressure combustor together with part of the reformed fuel stream F5. The hot gases leaving the combustor (stream 3) are expanded in the high pressure turbine section. The partially expanded gases (stream 4) then enter the reheat combustor together with the remaining reformed fuel F4. The reheated gases (stream 5) are expanded down to atmospheric pressure in the low pressure turbine. The turbine exhaust (stream 6) contains hot gases. Heat is recovered from these gases to first drive the methane-steam reformer, then the heat recovery steam generator. The exhaust gases are finally discharged into the atmosphere through the stack (stream 8). On the fuel side of the cycle, clean water (stream W1) is pumped up to pressure (stream W2), then heated and vaporized in the HRSG. The steam (stream

METHANE

REFORMEDFUEL

~J REFORMER AIR

STEAM p~

1 GENERATOR

Fig. 1. Chemically recuperated gas turbine concept.

WATER

PUMP

EXHAUST

AIR

HARVEY and KANE: ANALYSIS OF A REHEAT GAS TURBINE CYCLE

rS I "l

2

C

I

' L _ _ 4 f

C~I ~

F3

MSR

I CA3

Fig. 2. Reheat CRGT cycle.

1673

PUMP

W2 - - ~ _ ~ I

1,, I I _r2 wa_~: HRSG

MIXER F1 METHANE

18 I

EXHAUST

W3) is mixed with the methane fuel F1 in the MIXER, and the methane-steam mixture F2 is fed into the reformer. The reformed gas stream F3 is then split. Part of the stream flows through an expansion valve, before being fed into the reheat combustor (stream F4). The remainder of the fuel (stream F5) is fed into the HP combustor. From a thermodynamic viewpoint, better results would obviously be achieved by adopting a dual pressure HRSG and a dual pressure reformer. However, a recent study [4] indicates that the steam reformer will be a sizable and costly component of this type of cycle, and for this reason, a simplified single pressure chemical heat recovery set-up was retained. Further study is necessary to judge whether the adoption of a dual pressure chemical recuperation system is justifiable from an economic viewpoint.

CYCLE SIMULATION

The cycle simulation presented in this paper was undertaken using Aspen Technology's ASPEN + process simulation software. Before presenting the simulation results, we first review the main assumptions used to simulate the different cycle components. Unless otherwise noted, all parameter values are taken from Ref. [5].

Compressor

The compressor section is assumed similar to that of the ABB GT26 machine, with an overall compression ratio of 30:1, an air inlet flowrate of 530 kg/s, and 0.8% leakage.

Combustors

Conventional values are assumed for the pressure losses (3%) and the heat losses (0.4%). For flow control reasons, fuel injection pressure losses are higher, and are estimated at 30% (see Ref. [6]). It should be pointed out that for the proposed reheat CRGT cycle, major combustor redesign is necessary. This is due first of all to the low heating value of the fuel gas with typical steam-to-fuel ratios, resulting in low flame temperatures (which in turn can cause flameout problems and high CO emissions) and high fuel flowrates. Combustor cooling also requires substantial combustor redesign, due in part to the low heating value of the fuel (which results in a large fraction of combustion air entering the primary zone, leaving less air available for combustor liner cooling), and in part to the high fuel gas temperature. However, current development work is well advanced on such combustors, as discussed in Refs [1] and [7].


Turbines The GT26 turbine has a five stage expansion section, with a single stage located between the

HP and reheat combustors, and four stages in the low pressure turbine section. Equal pressure ratios are assumed for the five turbine stages. A simplified model was assumed to account for blade cooling and to estimate the necessary blade cooling flows. For the cooled expansion model, each turbine stage is considered separately. For each stage, the nozzle and rotor cooling flows are estimated. The nozzle cooling flow is then assumed to mix perfectly with the main flow prior to the flow through the turbine. The rotor blade cooling flow, on the other hand, is mixed with the stage flow at the stage exit. The cooling requirements are based on the temperature of the main gas stream at the stage inlet and outlet for the nozzle and rotor, respectively. The appropriate compressor bleed point location for each turbine stage cooling flow is estimated using a compressor-bleed fractional pressure drop (PD) and the following criterion (see Ref. [8]):

pcb _> pRI(1 q- PD) (1)

where pCb is the compressor bleed pressure and pR~ is the turbine rotor inlet stagnation pressure. For this preliminary study, strict equality was assumed in equation (1). PD was taken equal to 0.2 for stages 2-4 (stage 5 does not require cooling). For stage 1, the cooling air is taken directly from the compressor discharge. We also note that stage 2 cooling air is intercooled using an off-board intercooler, according to ABB's presentation of the GT24/26 in Ref. [3]. The intercooler air discharge temperature is set at 150°C.

The blade cooling flow for each turbine stage is estimated, using a simplified model proposed in Ref. [9]. For internal convection cooling, the coolant flow can be estimated using:

• tn Cp.g stgQbl q~ mcoo, = g Cp.cool Qgeb 1 -- ~b (2)

where subscripts g and cool refer to the main gas stream and the blade cooling stream, respectively. Standard values for Sty, Qb/Qg and eb are 0.005, 4 and 0.3, respectively (see Ref. []0]). ~b is defined by:

T~-Tb ~b = T g - Tcoo, (3)

in which Tb refers to the turbine blade metal temperature. If full or partial film cooling is associated with internal convection on the external blade surface,

the overall cooling flowrate requirements are reduced, typically by 10-30%, as discussed in Ref. [9]. For the case considered, we assume that stages 1 and 2 are film-cooled, with a maximum allowable blade temperature of 900°C, and coolant flowrate reduced by 15% compared to an internal convection cooled blade. Stages 3 and 4 are cooled by internal convection, with a maximum allowable blade temperature of 850°C.

Methane-steam reformer (MSR) The methane-steam reforming section contains the steam generator, consisting of a water heater

(economizer) and an evaporator, and the reformer itself, in which the chemical reforming reaction occurs between the fuel and steam, according to the following:

CH4 + 2H20 = CO2 + 4H2 (4)

CO2 + H2 = CO + H20. (5)

Reaction (4) is rate-limited by reaction kinetics, whereas the shift reaction (5) can be assumed to be at equilibrium for the conditions considered. The steam-to-methane ratios must be chosen sufficiently high to prevent carbon coking. Typical steam-to-methane ratios are in the 3-5 range. Reaction (4) proceeds in the presence of a catalyst, usually nickel-based. For the low temperature reforming considered in this case, it is necessary to use catalysts that are active at low temperatures. According to Ref. [1], catalysts can be found that are sufficiently active for temperatures above 600 K. For an overall cycle analysis, the chemical non-equilibrium effects due to reaction kinetics

HARVEY and KANE: ANALYSIS OF A REHEAT GAS TURBINE CYCLE 1675

can be modeled using the chemical approach to equilibrium AToq. For a typical reformer using a standard catalyst, ATeq is given by the following equations in Ref. [7]:

AToq = 0 if Tox, > 923 K (6)

ATeq= 43.33(1.0 T , , ~ 2 7 3 ) if Tex~,<923K (7)

where Text, refers to the temperature of the reformed fuel stream at the exit of the reformer. Tox~t can be computed by imposing a hot-side temperature approach (taken to be 25 K in this study). Another design constraint placed on the system involves the minimum allowable pinch point A T, which occurs at the water economizer exit, which is set at 10 K in this study.

S u m m a r y o f m a i n p a r a m e t e r v a l u e s

All main parameter values for the different cycle components are listed in Table 1.

S I M U L A T I O N R E S U L T S

All simulation results presented in this paper were obtained using Aspen Technology's ASPEN + process simulator. The thermodynamic properties were calculated using the Redlich-Kwong-Soave pVT equation of state with the ASPEN+ default binary interaction coefficients. The results presented include the results of an exergy analysis. The exergy flows of the material streams in the process were computed according to Ref. [11]:

ex = (h - ho) - To(s - so) + e~hm, x (8)

Table 1. Cycle data

Inlet conditions Air: 288 K, 101,325 Pa, 60% rel. humidity Fuel: 288 K, pressurized CH4, LHV = 50.056 MJ/kg Water: 288 K, 101,325 Pa

Compressor Mass flowrate 529.3 kg/s Inlet pressure loss (filter) 1% Air leakage 0.8% Pressure ratio 30.0 Polytropic efficiency 0.895

Combustors Heat losses 0.4% Pressure drops 3% Fuel injection pressure drop 30%

Turbine Firing temperature (TRIT) 1508 K qp, cooled stages 0.89 r/p, uncooled stages 0.925

Heat recovery steam generator Heat losses 0.7% Minimum pinch point AT 10 K Hot-side pressure drop 2% Cold-side pressure drop 10%

Reformer Heat losses 0.7% Hot-end temperature approach 25 K Hot-side pressure drop 2% Cold-side pressure drop 10%

Electro-mechanical losses Electro-mechanical efficiency 0.982


where subscript 0 refers to the reference conditions, chosen as 288 K, 101,325 Pa for this study. H20 is in the liquid state at reference conditions. The last term refers to the mixture chemical exergy. For a hydrocarbon fuel C~H,, the chemical exergy can be computed according to:

E.X~h = MW-----~ ao + x + a °2 - - x~tCo °2 - - -~ aHo ~°~') (9)

where superscript f refers to the fuel, and a0 to the molar availability at reference conditions, defined by:

do ~- ( h - Tos)lpo.ro. (10)

It is to be noted that ASPEN + does not have a built-in exergy function and, thus, computations of stream exergies and component exergy losses are extremely tedious. In this paper, the authors used an external Fortran subroutine to perform exergy calculations of all streams in the ASPEN + simulation. This subroutine was written by researchers at the Vrije University Brussels, Belgium (see Ref. [12]).

Table 2 shows the key results of the simulation. All results correspond to an inlet air flow of 529.3 kg/s, which is the design flow rate of the ABB GT26 machine. Table 3 lists the stream data for the different streams shown in Fig. 2. The two values for temperature and pressure for stream CA3 correspond to the flows in stage 2 and 3 of the LP turbine, respectively. The relatively low air flowrate for blade cooling of the LP turbine is due to the presence of the intercooler resulting in cold air being available for cooling of the hot stage nozzle and rotor.

Table 4 presents the detailed results of the exergy analysis of the cycle, i.e. the exergy losses associated with the different cycle components. The losses are presented in order of decreasing magnitude. The results are presented both in absolute values of the exergy flow losses, and as percentages of the fuel exergy inlet flow to the cycle. As expected, the major source of losses are the two combustors. The turbine losses include the losses associated with the mixing of the cooling air streams into the main expansion stream. The losses are greater in the LP turbine as a result of the much higher expansion pressure ratio in this turbine (15:1 compared to 2:1 in the HP turbine). We also note the significant losses associated with the fuel/steam mixer, which are of the same magnitude as the losses associated with the reformer! The losses associated with the LP fuel expansion valve are not particularly high, on the other hand. We also note that the losses in the steam generator are more than three times those associated with the reformer, whereas the heat duty of the steam generator is only 1.6 times that of the reformer.

Table 2. Cycle results summary

Net power output 383.0 MW Air flowrate 529.3 kg/s Fuel flowrate 14.0 kg/s Water flowrate 78.7 kg/s Stack temperature 408.0 K Thermal efficiency [LHV] 54.8%

Heat duties Reformer 132.3 MW Steam generator 236.8 MW Cooling air intercooler 4.7 MW

Table 3. Cycle stream data

Stream rh [kg/s] E, [MW] T [K] p [MPa]

1 529.32 0 288.0 0.101 2 461.10 246.84 823.5 3.040 3 524.61 712.17 1531.7 2.949 4 554,96 563.17 1298.2 1.400 5 584.18 785.60 1526.0 1.358 6 617.81 249.30 882.9 0.106 7 617.81 163.58 719.1 0.103 8 617.81 49.49 407.9 0.101 Wl 78.75 0 288.0 0.101 W2 78.75 0.48 288.2 4.733 W3 78.75 90.13 526.0 4.259 F1 13.98 724.12 288.0 4.259 F2 92.72 807.41 483.5 4.259 F3 92.72 885.76 857.9 3.833 F4 29.20 275.76 853.9 1.765 F5 63.51 606.67 857.9 3.833 CA1 30.36 16.25 823.5 3.040 CA2 18.59 6.87 667.7 1.489 CA3 15.04 3.66 565.5/464.1 0.860/0.454


DISCUSSION

The first conclusion to be drawn from the results presented, is the relatively good performance of the reheat CRGT cycle compared to a conventional state-of-the-art combined cycle. ABB announces a thermal efficiency of 58.5% and a power output of 365 MW, for a triple pressure reheat combined cycle based on the 50-Hz GT26 reheat gas turbine. The reheat CRGT cycle presented in this paper achieves a thermal efficiency which is slightly lower (54.8%), with a slightly higher power output (383 MW). It is, furthermore, important to recall that the heat recovery equipment used in the considered CRGT cycle is much simpler (single pressure, without reheat) than the triple pressure reheat heat recovery boiler used in the combined cycle. Also, the CRGT cycle does not require a condenser or a steam turbine. As discussed in the introduction, even better results would most certainly be achieved for the CRG T cycle with a dual pressure steam generator and reformer heat recovery equipment, but the resulting efficiency will, nevertheless, remain lower than the efficiency of the conventional combined cycle. The reason for this lower efficiency is discussed in detail in Ref. [7]: the reduced exergy losses due to the improved heat recovery system are more than offset by the increased losses associated with the fuel/steam mixing operation at the reformer inlet (these losses could, however, be reduced by preheating the fuel stream) and with the injection of large quantities of excess steam with the reformed fuel stream into the combustor. Further study is necessary to obtain more detailed results necessary for this comparison. Table 5 gives the component mass flows of fuel and air streams entering the high pressure combustor.

Table 5 shows that the combustible fuel is a relatively small fraction of the fuel stream, whereas a significant fraction of the fuel stream is steam. Thus, the CRG T cycle suffers from the same problem as the steam injected gas turbine cycle (STIG), i.e. the high irreversibilities associated with mixing steam into the gas cycle. This can be seen in Table 4: the combined exergy losses in the HP and reheat combustors represent 27% of the fuel exergy flow supplied to the cycle! A significant portion of these losses can be attributed to the mixing of the steam in the fuel stream into the gas stream. However, due to the inherently better nature of the heat recovery process, the efficiency of a CRGT cycle will always be superior to that of a STIG cycle. This is clearly apparent in the results of the exergy analysis shown in Table 4: the exergy losses in the reformer (7.4 MW) are lower than the losses in the steam generator (24.4 MW) despite a larger heat duty (see Table 2). The superior heat recovery associated with the chemical recuperation process can be seen in Fig. 3, which shows the temperature profiles of the hot and cold streams in the heat recovery section of the cycle. The cold-side profiles show water heating in the economizer, followed by constant temperature vaporization in the evaporator. At the evaporator exit, the methane fuel at ambient temperature is mixed with the steam, resulting in stream cooling. The methane-steam mix then enters the reformer section. Two cases are considered in this section. The first case considers catalytic steam reforming occurring in the reformer, resulting in the continuous line profile marked "with chemical recuperation". The endothermic nature of the reforming reactions can be seen from the curvature of the temperature profile. The profile also illustrates that the process is feasible, and that there is no temperature profile cross-over within the reformer. The other option considered involves simple heating of the methane-steam mix, resulting in the dashed line profile marked

Table 4. Component energy losses HP combustor LP combustor Stack loss Steam generator Compressor LP turbine Reformer Electromechanical Fuel/steam mixer HP turbine LP fuel valve Blade cooling air IC Pump Total losses

141.3 MW 19.5% 53.3 MW 7.4% 49.5 MW 6.8% 24.4 MW 3.4% 20.1 MW 2.8% 19.1 MW 2.6% 7.4 MW 1.0% 7.0 MW 1.0% 6.8 MW 0.9% 6.5 MW 0.9% 3.3 MW 0.5% 2.3 MW 0.3%

~ 0 MW 0% 346.5 MW 47.10%

Table 5. Mass flows entering HP combustor

Air stream rnAi, 461.1 kg/s

Fuel stream rhco 0.33 kg/s m.2 1.26 kg/s m. cm 7.02 kg/s m...:o 48.41 kg/s mco2 6.49 kg/s


1 0 0 0 . 0 , ~ , ,

.Q - i

o. E I -

8 0 0 . 0

6 0 0 . 0

4 0 0 . 0

200.0 0.0

J Without chemical ~ J /

j /

J ~/ Wi(thp~e imioCnal

Economizer i Evaporator ) Reformer : i )i~ > < - - ~ - - >i

, il , i i , i . , i

100.0 200.0 300.0 400.0 Heat exchanged [MJ/$]

Fig. 3. Heat recovery temperature profiles.

"without chemical recuperation". This curve clearly illustrates that considerably less heat recovery is possible in this case.

Table 5 also illustrates the low conversion of methane in the reformer for the operating conditions considered. A detailed analysis of the composition of reformed fuel stream at the exit of the reformer shows that only 27% of the methane feed is converted, and the remaining 73% remain unconverted. Furthermore, we note that the reformed fuel stream contains far more CO2 than CO or H2, suggesting that the reformer operating conditions favor the production of CO2 in the equilibrium shift reaction (5). These results illustrate a well-known property of methane reforming chemistry: methane conversion to CO and H2 is favored by high temperatures and low pressures. In the case considered, the reformed fuel leaves the reformer at relatively low temperature (858 K) and high pressure (3.83 MPa). Thus, the methane conversion could be improved by reducing the pressure ratio of the gas turbine, which would lower the pressure of the gas at the reformer exit and increase the turbine exhaust temperature. The results presented in this study are restricted to the gas turbine pressure ratio of the ABB GT26 machine, i.e. 30:1. Further work is necessary to assess the impact of reducing this pressure ratio while retaining the staged combustion concept. Indeed, further work is equally necessary to assess the advantage of the staged combustion concept compared to a standard gas turbine configuration with a single combustor.

Notwithstanding the good performance of the proposed CRGT cycle compared to a conventional combined cycle, there are other advantages of CRGT cycles. One of the key advantages of such cycles is their potential for ultra-low NOx emissions. Predicted NOx emissions for these cycles are as low as 1 ppm, without using selective catalytic reduction of the exhaust gas stream, as discussed by Ref. [7]. Such low values are projected because of the low flame temperature of the reformed fuel gas, on the order of 2100 K with typical steam/methane ratios. This flame temperature is lower than that which can be achieved with steam injection. As a means to reduce emissions, steam injection is limited by flameout and high CO production considerations. With a reformed gas fuel stream, the fuel contains significant quantities of hydrogen, which has a much lower flameout temperature than methane. Furthermore, the presence of H2 reduces CO emissions at a given flame temperature. However, combustion tests are necessary to confirm the emissions capability of CRGT cycles.

CONCLUSIONS

In this paper, we have presented a reheat gas turbine cycle with chemical recuperation. The analysis is based on ABB's GT26 heavy duty reheat gas turbine. The results of the analysis show thermal efficiency values that are slightly lower than state-of-the-art combined cycle power


generation systems, confirming the accepted view in the power generation industry that combined cycles remain the preferred solution for efficient base-load power production. However, further work is needed to optimize the proposed cycle in terms of pressure ratio (the 30:1 pressure ratio o f the GT26 machine is in fact optimized for combined cycle operation) and improved heat recovery equipment (in this study, a single pressure steam generator / reformer was considered, and further study is necessary to evaluate the improvement that might be achieved by adopt ing a dual pressure configuration). Preliminary calculations under taken by the California Energy Commission (Ref. [1]) for a reheat C R G T cycle based on aeroderivative gas turbine technology show superior performance, indicating that heavy duty C R G T cycles may not be the optimal configuration for high efficiency values.

Further work is also required to assess the part load performance of C R G T cycles. Part load efficiencies o f combined cycle plants are notoriously poor, and if the efficiency of C R G T cycles can be shown to deteriorate less at part load conditions, C R G T cycles may turn out to be of considerable interest for intermediate load applications. Such is the case for example for the humid air turbine (HAT) cycles currently being proposed, which show efficiencies below those o f combined cycles, but clearly superior part- load performance, together with rapid startup times. Finally, one o f the key considerations fuelling current interest in C R G T cycles is the potential for ultra-low NOt emissions without the need for selective catalytic reduction o f the exhaust stream.

Acknowledgement--The authors would like to thank Professor B. Facchini (Energy Systems Engineering Department, University of Florence, Italy) for a number of suggestions and comments that were of great help in preparing this paper. The authors are also grateful to Professor J. De Ruyck (Vrije Universiteit of Brussels, Belgium) for providing us with the ASPEN+ exergy analysis tool.

REFERENCES

1. Janes, J., Chemically recuperated gas turbine, California Energy Commission Staff Report P500-92-015, 1992. 2. Olmsted, J. H. and Grimes, P. G., Proceedings, 7th lntersociety Energy Conversion Engineering Con/erence, 1972,

pp. 241-248. 3. Farmer, R., Gas Turbine WorM, 1993, Sept.-Oct. 18-23. 4. Adelman, S.T., Hoffman, M. A, and Baughn, J.W., ASME Journal of EngineeringJor Gas Turbines and Power, 1995,

117, 16-23. 5. Macchi, E., Consonni, S., Lozza, G. and Chiesa, P., ASME Journal of Engineering for Gas Turbines and Power, 1995,

117, 489-498. 6. Foster-Pegg, R.W. and Elmasri, M.A., Design of gas turbine combined cycle and cogeneration systems, Seminar notes,

Thermoflow Inc. 1996. 7. Lloyd, A., Thermodynamics of chemically recuperated gas turbines. MSc Thesis, Center for Energy and Environmental

Studies, Princeton University, USA, 1991. 8. Kesser. K. F., Hoffman, J. W. and Baughn, J. W., ASME Journal of Engineering for Gas Turbines and Power, 1994,

116, 277 284. 9. Stecco, S. and Facchini, B., Proceedings, 1989 ASME Cogen-Turbo Symposium, 1989.

10. Elmasri, M. A., ASME Journal of Engineering for Gas Turbines and Power, 1986, 108, 151-159. 11. Moran, M., Availability' Analysis: A Guide to Efficient Energy Use. Prentice-Hall, Englewood Cliffs, NJ, 1982. 12. Bram, S. and De Ruyck, J., Proceedings, ECOS'96 Symposium, June 25-27, 1996, Stockholm, Sweden, 1996,

pp. 217-224.

Analysis of a reheat gas turbine cycle with chemical recuperation using Aspen

Documents

Transcript of Analysis of a reheat gas turbine cycle with chemical recuperation using Aspen