Article pubs.acs.org/EF
Chemical Process Engineering Principles of Combustion Turbines William L. Luyben* Department of Chemical Engineering, Lehigh University, Bethlehem, Pennsylvania 18015, United States ABSTRACT: The lower prices and wider availability of natural gas make it important that chemical engineers become familiar with natural-gas processing and applications. One important use is the generation of power by burning natural gas in a combustion turbine. Traditional systems use air for the oxygen source, but the use of high-purity oxygen with stack gas recirculation is an alternative when sequestration of carbon dioxide is desired. The limiting design parameter in a combustionturbine system is a maximum turbine inlet (combustor) temperature because of metallurgical constraints. In air-fired systems, enough air must be used to keep below this limit. In oxygen-fired systems, enough CO2-rich stack gas must be recycled to keep below this limit. The purpose of this paper is to review the chemical engineering process design principles involved in setting up combustion-turbine systems.
1. INTRODUCTION
oxygen to keep temperatures in the combustor below this maximum limit. The mechanical design of combustion turbines is a very complex fluid mechanics problem that includes issues of heat transfer, ignition, flame speed, flame stability, stress analysis, and materials of construction. This is the province of mechanical and combustion engineering. However, the design of the chemical process system for operating a combustion turbine is also not a trivial exercise. The chemical engineering design issues in air-fired systems involve finding the amount of air to be fed and the optimum combustor pressure because of the trade-off between compressor work and turbine power. The chemical engineering design issues in oxygen-fired systems involve finding the amount of oxygen to be fed, the amount of stack gas to be recycled, and the optimum combustor pressure. There are two basic types of combustion-turbine systems. In the conventional system, air is used as the source of oxygen. Air is free, but it contains a large amount of nitrogen; thus, there is a large flow rate of the resulting stack gas after combustion that has a low concentration of carbon dioxide (3 mol %). If sequestration is desired, the use of high-purity oxygen can produce a much smaller amount of stack gas with a much higher CO2 concentration (66 mol %). Of course, oxygen is not free. The maximum combustor temperature limitation requires compressing and recycling of some of the stack gas, which adds an additional parasitic load. A vast amount of literature exists that deals with all of the complex issues associated with gas turbine and energy systems. Some general references include studies by Cohen et al.,1 Walsh and Fletcher,2 and Putman.3 A recent review paper by Habib et al.4 deals with oxygen-fired systems and contains 170 references. Flowsheets of the basic air- and oxygen-fired combustionturbine systems are discussed in the following sections. For background and comparative purposes, we start by looking at a conventional air-fired furnace/boiler system that burns natural
Exploration for crude oil and natural gas has been revolutionized in recent years by the technology of hydraulic fracturing (fracking), which has been a “game-changer” in the energy field. Predictions made only 3−4 years ago that forecast a tight supply of natural gas have, like most attempts to guess future events, been proven incorrect. Exactly the opposite has occurred. Natural gas prices have dropped, and there is even talk of exporting instead of importing natural gas. Domestic crude oil production has also increased because of fracking technology. Lower energy and raw material prices are producing a renaissance in the U.S. chemical industry. The economics have shifted in favor of domestic production. The wider application of natural gas makes it increasingly more important that chemical engineers become familiar with its processing and uses. One of the most important applications is the generation of power by burning natural gas in a combustion turbine. Combustion turbines have higher efficiencies than traditional steam-generating boilers, particularly when applied in a combined-cycle configuration (use of the hot exit gas from the combustion turbine to generate steam to drive a steam turbine, which generates more power). A combustion turbine combines fuel (methane) and an oxygen source (air or high-purity oxygen) to produce a hightemperature, high-pressure, large-volume gas stream that is fed to a turbine. The work extracted from the turbine as the gas flows from the high-pressure inlet to the low-pressure outlet is a source of shaft work, which can be used to drive electrical generators, pumps, or compressors. Of course, some of the power generated in the turbine must be used to compress the air (oxygen) feed. The higher the combustor pressure, the higher the parasitic load of the air (oxygen) compressor. However, the higher the combustor pressure, the more power is generated in the turbine. The limiting design parameter in combustion-turbine systems is a maximum turbine inlet (combustor) temperature because of metallurgical constraints. This means that enough inert material must be fed along with the fuel (methane) and © 2013 American Chemical Society
Received: July 12, 2013 Revised: September 12, 2013 Published: September 13, 2013 6316
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gas to generate steam to drive a turbine. Aspen simulations using Peng−Robinson physical properties are used to calculate compressor and turbine energies, assuming polytropic efficiencies of 80%. An Aspen RGibbs chemical-equilibrium reactor model is used for the combustor.
2. CONVENTIONAL AIR-FIRED FURNACE WITH STEAM TURBINE Figure 1 gives a greatly simplified flowsheet of a conventional power plant. Fuel (methane) is burned in a furnace with air to
Figure 2. Combustion turbine using air.
oxygen, water, and carbon dioxide. The required air flow rate shown in Figure 2 (32 890 kmol/h) is 3.45 times higher than the stoichiometric amount. The turbine produces 138.1 MW of power as the gas flows from inlet conditions of 8 atm and 1343 K to outlet conditions of 1 atm and 923 K. The pressure in the combustor is a design optimization variable, which we will explore later in this paper. The pressure used in Figure 2 is 8 atm. The air must be compressed from 1 atm up to this pressure, which consumes 93.54 MW of power (80% polytropic efficiency assumed). In a combined cycle system, the hot gas from the combustion turbine is sent to a steam boiler that generates additional power. The total energy recovered in cooling it to 422 K is 150.1 MW. Assuming an efficiency of 30%, the energy recovered from the boiler/steam turbine is 45.0 MW. Therefore, the net production of power (shaft work) in this air-fired combustion-turbine system is 138.1 + 45.0 − 93.54 = 89.56 MW. This should be compared to the conventional fired furnace steam-turbine power generation of only 63.0 MW for exactly the same amount of fuel. This numerical example demonstrates clearly the advantages of a combustion-turbine combined cycle system.
Figure 1. Conventional air-fired boiler.
produce heat. For the numerical example considered in this paper, the methane flow rate is set at 1000 kmol/h for all cases. At the high temperatures existing in the furnace, the combustion reactions completely convert methane into carbon dioxide and water. We ignore minor reactions of nitrogen to form nitric oxide. CH4 + 2O2 → CO2 + 2H 2O
Every mole of methane requires 2 mol of oxygen. Because the oxygen composition of air is 21 mol %, every mole of oxygen brings in 79/21 = 3.762 mol of nitrogen. If just the stoichiometrically required oxygen were fed, a methane feed flow rate of 1000 kmol/h would require 2000 kmol/h oxygen, which would require an air feed flow rate of (2000)(100/21) = 9523 kmol/h. However, an excess amount of air must be used to ensure complete combustion. Assuming a stack gas composition of 5 mol % O2, the required air feed is 12 810 kmol/h. A complex heat-exchanger system is used to recover as much of the heat as possible. The limitation is a minimum stack gas temperature of 422 K (300 °F) to prevent condensation of water in the stack. The heat boils and superheats water that drives a steam turbine. In the figure, only the heat released and captured is shown (210.4 MW), with a fuel (methane) flow rate of 1000 kmol/h. The thermal efficiency of a furnace/boiler system is about 30%, with most of the energy going out in the cooling water used to condense the low-pressure steam leaving the steam turbine. The net power generated is about 63.0 MW.
4. MULTI-STAGE COMPRESSION A single-stage air compressor is shown in Figure 2. With a compression ratio of 8, the discharge temperature is very high (664 K, 735 °F) and may be above a maximum temperature limitation because of materials of construction (shaft seals). Walas6 recommends a maximum discharge temperature of 450−477 K (350−400 °F). Luyben7 discusses various criteria for the conceptual design of multi-stage compression systems, including an economic analysis that finds the optimum trade-off between compressor power and capital investment. Using a multi-stage air compression system reduces air compressor power and the required amount of air to achieve a specified combust temperature, but this results in a reduction of the power generated by the combustion turbine. Figure 3 gives a flowsheet with a two-stage air compressor system. Using the heuristic that the compression ratio should be equal in each stage, the discharge pressure of stage 1 is 2.828 atm and the discharge pressure of stage 2 is 8 atm. Note that the discharge temperatures are greatly reduced (465 and 471 K) to below the Walas limit. A water-cooled heat exchanger cools the gas from the first-stage compressor to 322 K.
3. AIR-FIRED COMBUSTION TURBINE Figure 2 gives a typical combustion-turbine flowsheet. For the same natural gas flow rate of 1000 kmol/h, enough air must be fed to give a combustor temperature at a maximum limit, which is stated5 to be 1343 K (1950 °F). The combustor is essentially adiabatic; therefore, the exothermic heat of reaction raises the temperature of the gas stream, which contains nitrogen, excess 6317
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combustion and steam turbines is quantified for the numerical example of a fuel flow rate of 1000 kmol/h. Results showing the effects of combustor pressure and combustor temperature are shown in Figure 4. The top left graph shows that increasing pressure requires more air for a constant combustor temperature. The reason for this effect is the increase in the compressor discharge temperature as the compressor discharge pressure increases. More air is required for the same fuel flow rate to keep a constant combustor temperature. At a constant pressure, increasing the limit on the combustor temperature reduces the required amount of air. Therefore, the higher the metallurgical temperature limitation, the less air needs to be compressed. The middle left graph shows that compressor work increases with discharge pressure as expected. Compressor work decreases with the maximum temperature limitation because less air is required. The middle right graph shows that the power generated in the combustion turbine increases with combustor pressure because the expansion ratio increases (discharge pressure is constant at 1 atm). Higher permitted temperatures require less air; therefore, the flow rate of gas through the turbine is smaller, which generates less power. The bottom left graph shows that the power generated in the steam turbine in the combined cycle process decreases as the pressure increases because of the smaller flow rate of the hot gas entering the boiler. As expected, the bottom right graph shows that the oxygen compressor discharge temperatures increase as the compression ratio increases. It is not a function of the combustor temperature because the air flow rate has no effect on the discharge temperature. The key graph is that shown in the upper right. The net power generation is plotted versus combustor pressure for three different maximum temperature values. The net power is the combustion-turbine power plus the steam-turbine power minus the compressor work in the two air compressors. For each temperature, an optimum combustor pressure exists at which net power is maximized. The optimum pressure and net power increase as the temperature limitation is increased.
Figure 3. Combustion turbine using air, for a two-stage air compressor.
Total compressor work is very significantly reduced from 93.54 MW (Figure 1) to 30.47 + 31.76 = 62.23 MW. The required air flow rate to maintain the fixed combustor temperature is also reduced from 32 890 to 26 080 kmol/h because the air going to the combustor is at a lower temperature (471 versus 664 K). However, the smaller air flow rate reduces the power produced in the combustion turbine from 138.1 to 110.5 MW. The power generated in the steam turbine is also smaller (36.42 versus 45.0 MW). Therefore, the net power (the shaft work available to drive a generator or compressor) is slightly smaller (84.69 versus 89.56 MW), despite the large reduction in compressor work.
5. COMBUSTOR PRESSURE OPTIMIZATION The combustor pressure used in the previous studies was 8 atm. In this section, we justify this selection for a two-stage air compressor system. The process engineering trade-off between the power to compress the air and the power generated in the
Figure 4. Effect of the pressure and Tmax, for an air-fed combustion turbine, with two-stage air compression. 6318
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Figure 5. Combustion turbine using oxygen with CO2 recycle, with one-stage O2 and CO2 compression, for an 8 atm combustor.
fired single-stage combustion-turbine system (compare Figure 2 to Figure 5). Therefore, the principle operating cost difference between the air- and oxygen-fired combustion-turbine systems is the cost of oxygen. The air and recycle compressors both have very high discharge temperatures if single-stage compression is used. Figure 6 gives an oxygen-fired flowsheet, in which two stages of
6. OXYGEN-FIRED COMBUSTION TURBINE Notice that the stack compositions leaving both the conventional fired furnace (Figure 1) and the air-fired combustion turbine (Figures 2 and 3) show low concentrations of carbon dioxide (7.24, 2.95, and 3.69 mol % CO2) and very large flow rates (13 810, 33 890, and 27 080 kmol/h). These make the recovery of CO2 for sequestration difficult and expensive. If sequestration is desired for environmental, political, or economic reasons, the use of an oxygen-fired combustion turbine can be considered. As shown in Figure 5, the stack gas composition is much richer in CO2 (65.75 mol %) and the stack gas flow rate is an order of magnitude smaller (1499 kmol/h). Thus, CO2 recovery should be much less expensive. However, this system incurs an operating cost for supplying high-purity oxygen; therefore, this must be considered in the economics of carbon sequestration. What we consider in this section is the effect on power generation. The combustion of methane in an oxygen/CO2 mixture is reported by Habid et al.4 to be considerably different from an oxygen/nitrogen mixture. The composition of oxygen must be about 26 mol % in the CO2 system to obtain adequate flame and combustion results. Therefore, the flowsheet given in Figure 5 specifies the composition of the “Mix” stream to be 26 mol % O2. This is achieved in the Aspen simulations using a Flowsheeting Options Design Spec that varies the fresh oxygen feed flow rate. The combustor temperature is held at 1343 K by varying the flow rate of the “Recycle” stream. For a combustor pressure of 8 atm, the power required for a single oxygen compressor is only 6.27 MW. Because pure oxygen is used, there is no nitrogen to serve as a heat sink. Therefore, a large recycle of stack gas is required to keep the combustor temperature at its 1343 K maximum limit. The 17 870 kmol/h recycle gas must be compressed and requires 45.98 MW of work if a single-stage compressor is used. Note that the compression cost of the recycle gas is much larger than the compression cost of the oxygen feed. The turbine generates 92.36 MW of power. The steam boiler produces 160.6 MW of energy, which is converted to 48.18 MW of power, assuming 30% efficiency. The net power generation in this single-stage compression system is 88.29 MW, which is only slightly lower than the air-
Figure 6. Combustion turbine (8 atm) using oxygen with CO2 recycle, with two-stage O2 and CO2 compression.
compression are used in both the air and recycle compressors. Compressor discharge temperatures are reduced from 649 and 556 K in the single-stage process to 467 and 434 K in the twostage process, which is below the Walas limitation. It is interesting to note that compression of oxygen results in higher temperatures than the compression of the recycle gas, which is rich in carbon dioxide. For the same compression ratio of 8 (compressing from 1 to 8 atm) and starting with the same inlet temperature (322 K), the discharge temperature of the oxygen compressor in the single-stage system is 649 K, while that of the recycle compressor is only 556 K. The difference in molecular structures between oxygen and carbon dioxide causes this result. The heat capacity ratio of oxygen is higher than the heat capacity ratio γ of the carbon-dioxide-rich recycle stream. 6319
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Figure 7. Effect of the combustion pressure, for a turbine using oxygen with CO2 recycle, with two-stage O2 and CO2 compression, for a 1343 K combustor.
Figure 8. Combustion turbine (15 atm) using oxygen with CO2 recycle, with two-stage O2 and CO2 compression.
As γ increases, the adiabatic isentropic compression work increases. The two-stage system reduces oxygen compressor work from 6.27 to 5.20 MW and the recycle compressor work from 45.98 to 33.27 MW. However, the combustion-turbine power drops from 92.36 to 80.66 MW because of the smaller gas flow rate (17 380 versus 20 100 kmol/h). Steam-turbine power is also smaller (42.96 versus 48.18 MW). The net result is a reduction in shaft work from 88.29 MW in the single-stage compressor process to 80.66 MW in the two-stage system.
the pressure was varied over a range from 5 to 20 atm using two-stage oxygen and recycle compressors and maintaining a 1343 K combustor temperature. Figure 7 gives results for the oxygen-fired system. As pressure increases, more recycle gas is required and more power is produced in the combustion turbine. However, less power is produced in the steam turbine because of the lower temperature of the gas leaving the combustion turbine (more power extracted). For example, at 8 atm, the temperature is 1037 K (Figure 6). At 15 atm, the temperature is 961 K. The upper right graph in Figure 7 shows an asymptotic rise in net power with increasing pressure. There is no maximum in the curve. However, the lower right graph shows how the discharge temperature of the recycle compressor increases with increasing
7. OPTIMUM PRESSURE IN AN OXYGEN-FIRED COMBUSTION TURBINE The combustor pressure used in the previous section was 8 atm. Is this the optimum pressure? To explore this question, 6320
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(4) Habib, M. A.; Nemitallah, M.; Ben-Marisour, R. Recent developments in oxy-combustion technology and is application to gas turbine combustors and ITM reactors. Energy Fuels 2013, 2−19. (5) http://www.skybrary.aero/bookshelf/books/1621.pdf. (6) Walas, S. M. Chemical Process Equipment, Selection and Design; Butterworth-Heinemann: Woburn, MA, 1990. (7) Luyben, W. L. Compressor heuristics for conceptual process design. Ind. Eng. Chem. Res. 2011, 50, 13984−13989.
pressure. Using the Walas limit of 470 K, the optimum design is a 15 atm combustor pressure. Figure 8 gives the flowsheet of an oxygen-fired two-stage compression process operating with a combustor pressure of 15 atm. A comparison of the air- and oxygen-fired systems can be made by looking at Figures 3 and 8. For the same flow rate of fuel, there is not a whole lot of difference between the two processes. The net shaft works are 84.69 MW for the air-fired process versus 83.78 MW for the oxygen-fired process. The combustion turbine in the air-fired process is slightly larger (110.5 versus 102.9 MW), but the steam boiler and steam turbine are slightly smaller (121.4 versus 127.6 MW). The total compression work of the two air compressor in the air-fired process is 62.23 MW, while the total compression work of the four compressors in the oxygen-fired process is 56.86 MW. The only significant difference between the two processes is the operating cost of purchasing the oxygen feed. Capital costs have not been considered in this analysis. The oxygen-fired process requires more equipment (four compressors, two drums, and two heat exchangers), but the compressors are smaller (56.86 total MW) than the two compressors in the air-fired process (62.23 MW). The combustion turbine is also smaller (102.9 versus 110.5 MW), but the boiler and steam turbine are larger. Using the Economic Analysis function in version 8 of Aspen Plus predicts a total installed capital investment of $24 000 000 for the air-fired process and $20 000 000 for the oxygen-fired process. Utility costs are reported to be about the same ($26 000 000 per year). The basis for the capital investment calculations is “under the hood” in Aspen Plus; therefore, the reliability of these results has not been confirmed.
8. CONCLUSION This paper has attempted to provide chemical engineering practitioners and students with a clear understanding of the basic process principles in the design and operation of naturalgas combustion-turbine systems. Turbine metallurgical hightemperature limitations require air feed flow rates that are significantly above the stoichiometric amounts needed for combustion of the methane fuel in air-fired systems. This metallurgical high temperature requires recycling of CO2-rich stack gas to serve as a heat sink in oxygen-fired systems. Air-fired processes have an optimum combustor pressure that balances air compression work with combustion-turbine power produced. The optimum pressure in oxygen-fired systems corresponds to running at a maximum compressor discharge temperature. Multi-stage compression setups should be used in both processes.
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AUTHOR INFORMATION
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[email protected]. Notes
The authors declare no competing financial interest.
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REFERENCES
(1) Cohen, H.; Rodgers, G. F. C.; Saravanamuttoo, H. I. H. Gas Turbine Theory, 4th ed.; Longman: Harlow, U.K., 1995. (2) Walsh, P. B.; Fletcher, R. Gas Turbine Performance, 2nd ed.; American Society of Mechanical Engineers (ASME): New York, 2004. (3) Putman, R. E. Industrial Energy Systems; American Society of Mechanical Engineers (ASME): New York, 2004. 6321
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