REHEAT GAS TURBINE

Abstract: Apparatus for generating useful power and process for generating useful power by reheating gas turbine exhaust gases to directly generate power, or alternatively, by reheating the exhaust gas to directly generate power and further reheating superheated steam by heat exchange for combined gas reheat and steam superheating with optional steam reheating as well. Steam-cooling with a steam thermal barrier for vanes and blades in a gas generator (36, 30A) or power turbine (42) offers advantages over cooling with air or water at the higher temperatures characteristic of the high temperature turbine technology gas turbines coupled with the reheat steam cycle. Consequently, higher cycle efficiency is obtainable for turbine (42) initial firing temperatures of about 2500`F (1371`C) and 2050`F (1121`C) reheat temperatures where over-all combined cycle degradation resulting from steam cooling is below about 1%. A turbine (42) having steam-cooled components can operate at a cycle pressure ratio of about 38 without intercooling and can include steam superheating in the reheat combustor (90).

Description: Description Reheat Gas Turbine Technical Field The invention relates generally to pro- duction of useful power. More particularly, improved power generation results from reheating of exhaust gas formed in the first stage of a reheat cycle. Steam cooling of high temperature components permits high temperature operation. Specific apparatus having high over-all operational efficiency in a reheat cycle is included within the technical field of the invention. Background of the Invention Research and development is currently being directed toward many configurations of power systems involving gas turbines because of the growing awareness of impending world energy short¬ ages. The present invention relates to the need to focus technical attention to the reheat cycle and with use of the apparatus and processes of the present invention, the reheat gas turbine cycle and combined gas reheat and steam reheat cycle can appreciably increase power plant thermal efficiency to approximately an over-all 50% efficiency level. The reheat gas turbine cycle itself is well-known and has received considerable attention over the years, particularly in Europe. A regene¬ rator has been used to heat a compressor's discharge air to improve cycle efficiency, and intercooling has been suggested for the same purpose. It is known that reheating can increase power output by 35 to 40%, but without use of regeneration, over¬ all cycle efficiency is degraded. FO » Another example of a reheat gas turbine cycle presently employed is the afterburner of a jet engine for aircraft use, such as in military aircraft and in certain commercial supersonic planes. The jet reheat cycle has been developed to get reliable service in applications where augmented power output is required for a short or limited time. The greatly increased power output comes at the expense, however, of markedly increase.d fuel consumption, owing to which, commercial ap ' plications had gone to efficient high bypass ratio fan jets for subsonic flight. Technology developed therefrom has made available second generation high-ratio high- firing temperature gas generators for industrial appli¬ cations. Yet, despite the existence of known technology, combined cycle power plants, such as those which utilize the processes and apparatus of the present invention, have not been developed. Attention has not been given to reheat gas turbine combined cycle power plants because it has been thought that the degradation and efficiency of the reheat cycle would not offer an advantage, and also that increased fuel consumption would result. However, higher level heat available in the exhaust and the over-all entropy changes as well as the concept of extracting maximum work at the high working fluid temperature levels is possible with the present invention to obtain highest efficiency. Another explanation of the failure of others to utilize a reheat cycle process is the feeling that greater complication and cost of controls, and additional burner, compatibil¬ ity of nozzle area, start-up procedures, and the like, would offset any advantage in specific power output, particularly at a higher fuel consumption. OMPI IPO Yet another reason can be cited for the failure of others to point in the direction of present invention. Aircraft derived gas turbines offer an advantage in potential physical arrangement for the reheat cycle in that a reheat combustor can be readily added between the gas generator and the powe-r turbine, while conventional industrial or heavy duty gas turbines are ordinarily single shaft units for power generation, such conventional units not readily lending themselves to addition of reheat burner. With the present invention, however, an arrangement is disclosed which utilizes a reheat burner for heavy duty applications. Moreover, second generation aircraft gas turbines fire at elevated temperatures and utilize high compression ratios suitable for reheat cycles, while industrial units, because of a single shaft configuration, are limited to lower ratios. With the appearance of an increasing number of high-temperature and high-pressure-ratio gas turbines which lend themselves to reheat cycles, use of such equipment according to the pro¬ cess and teachings of the present invention becomes technically feasible, affording a practical utility for the apparatus and processes of the present invention. In apparatus for generating useful power by reheating gas turbine exhaust gases, with or without further reheating of super-heated steam for combined gas reheat and steam superheating, condi¬ tions of temperature and pressure having relatively high cycle efficiencies typically involve suffi- OMPI /«, WIPO - ciently high temperatures and high pressures to necessitate special cooling of gas turbine compo¬ nents. Conventionally, blade cooling is accom¬ plished with air to extract heat from the blade surface. The physical properties of air limit its effectiveness as a coolant medium, and impose an upper limit to the temperatures attainable in a gas turbine.. Moreover, air must be sparingly and very carefully used due to its dearness and its effect on cycle degradation both with respect to power and efficiency. At all cycle pressure ratios, and particularly the higher cycle pressure ratios encountered in apparatus involving reheating gas turbine exhaust gases, very little pressure is available to utilize in the region between the compressor discharge and the nozzle and blade pressures encountered. Other approaches to gas turbine component cooling being studied, researched and developed include transpiration air cooling, water cooling and alternative materials of construction capable of withstanding the conditions encountered. Many additional problems incident to prior approaches to component cooling include corrosion of the metal surfaces at high temperatures, low- cycle metal fatigue, radiant heat absorption, and first-stage nozzle and blade surface fouling during start-up and operation. Disclosure of Invention The invention contemplates a process and apparatus for generating compressed, heated gas in a gas generator arrangement, and then reheating the exhaust gas for generation of power in a second turbine, followed by extraction by heat exchange " of the useful energy in the exhaust gas from the second turbine to indirectly power a steam turbine through generation of steam. A combustor accepts a heated and compressed gas produced by a conventional gas generator, adds fuel and delivers reheated gas to a power turbine for directly, generating power. The exhaust gas from the po'wer turbine then indirectly drives the steam turbine by heat exchange. Alternatively, the combus¬ tor for reheating the gas generator exhaust gas incorporates heat exchange means for superheating steam therein prior to delivery of reheated gas to the power turbine and optionally for production of reheated turbine steam for return to the steam turbine to furnish additional power thereto. In order to permit increasing of the operating temperatures and pressures at which a gas turbine is to be operated for maximum efficiency, steam is used as a coolant and as a thermal barrier to optimize the gas turbine cycle pressure ratio and to simultaneously take advantage of the cold rehea ' t steam formed from the steam turbine in order to provide vane and blade cooling for the gas generator first stage, and extracted steam is utilized to provide cooling for second-stage gas generator blading and the power turbine blading. Furthermore, the invention contemplates use of film cooling employing steam as the coolant, whereby steam can be used as a coolant to advantage without sacrificing blade surface roughness, which can be critical where high Reynolds numbers are encountered. Brief Description of Drawings Figure 1 is a schematic view of a gas generator with a reheat gas power turbine driving an electric generator, along with a steam turbine driving a second electric generator in a combined reheat gas power turbine and steam turbine cycle. Figure 2 is a side elevational view, partly in diagrammatic section, of a gas generator with an associated second reheat combustor. Figure 3 is a schematic view of a gas generator with a combustor cavity associated with a reheat gas power, turbine and provided for steam superheating and steam reheating. Figure 4 is a longitudinal sectional view of an axial flow reheat combustor cavity for effect¬ ing gas turbine reheating, steam superheating and steam reheating. Figure 5 is a transverse sectional view of the combustor of Fig. 4 showing in addition details of the mounting support and gib key. Figure 6 is an enlarged transverse sec¬ tional view of a tube for superheating steam and/or reheating in the combustor cavity of Fig. 4. Figure 7 is a diagrammatic view of the steam path heat balance in a conventional reheat steam turbine power plant cycle. Figure 8 is a diagrammatic view of the heat balance in a reheat bottoming cycle for the same steam conditions used in calculations presented herein. O Figure 9 is a graph showing cycle effi¬ ciency as a function of net power output for various cycle pressure ratios and firing temperatures for each of three conditions of no reheat, partial reheat and full reheat. Figure 10 is a graph showing power turbine expansion ratios as a function of cycle pressure ratios for various firing temperatures of a gas generator combined with a power turbine. * Figure 11 is a graph showing gas generator and power turbine exit temperatures as a function of cycle pressure ratios for different firing tempera¬ tures of a gas generator and associated power turbine. Figure 12 is a graph showing the fuel ratio of the first combustor and the reheat combustor as a functional cycle pressure ratio for various firing temperatures. Figure 13 is a graph showing combined reheat gas turbine and steam turbine cycle efficiency as a function of firing temperature for a full reheat cycle and for a simple cycle. Figure 14 is a graph showing the combined output as a functional firing temperature for both a simple cycle and a full reheat cycle gas turbine. Figure 15 is a graph showing cycle effi¬ ciency showing the percentage conversion of heat to steam under the optimum pressure ratio for steam cycle efficiency and output, for three firing tem¬ peratures in the left-hand portion of Fig. 15, and showing the cycle efficiency for optimum pressure ratios for steam cycle efficiency and steam cycle output as a function of reheat temperature rise for OMPI . WIPO ..-. three firing temperatures in the right-hand portion of Fig. 15. Figure 16 is a graph showing the ratio of gas turbine output to steam turbine output as a function of reheat temperature rise for the optimum pressure ratio for efficiency and the optimum pres¬ sure-ratio for output, for various firing tempera¬ tures and for full and partial reheat. Figure 17 is a graph showing the relation- sh ' ip of temperature and enthalpy of both the steam being generated and the exhaust gases for three conditions: superheated and reheated steam, super¬ heated steam only, and thirdly saturated steam only with no superheat. This graph also presents the effect the boiler pinch point has on final stack temperature. Figure 18 is a graph showing the tempera¬ ture differentials between the gas generator bypass gas around the combustor cavity and the steam being superheated. It also shows the temperature differen¬ tials between the combustion cavity gas and the steam being super-heated. Figure 19 is a fragmentary sectional view of a modified form of the combustion cavity of Fig. 3, showing its adaptation for burning of finely divided solid fuel, such as powdered coal. Figure 20 is a side elevationa-1 view, partly in diagrammatic section, of a gas generator and a reheat power turbine, such as can be used in a combined reheat gas power turbine and steam turbine cycle, the components of which can be cooled by steam to permit high temperature and high pressure opera¬ tion according to the present invention. Figure 21 is a graph showing cycle effi¬ ciency where the percentage conversion of heat to steam under the optimum pressure ratio for steam cycle efficiency and output is shown as a function of three firing temperatures in the left-hand portion, and cycle efficiency for optimum pressure ratio for steam cycle efficiency and steam cycle output is shown as a function of reheat temperature rise for three firing temperatures in the right-hand portion. Figure 22 is a schematicized sectional view of. a first-stage nozzle vane showing regions of typical radiation and convection heat flux around the vane, bounded by combined heat transfer coefficients for 2000°F (1093°C) turbine inlet temperature. Figure 23 is a graph of e issivity and absorptivity as a function of the natural logarithm of pressure in PSIA. for steam and for combustion products. Figure 24 is a transverse sectional view of a first-stage nozzle vane adapted for use of steam as a coolant. Figure 25 is a sectional view of the trailing end of the nozzle vane of Fig. 24, taken substantially upon a plane passing along section line 25—25 of Fig. 24. Figure 26 is an enlarged view of the leading edge of the head of the nozzle vane of Fig. 24, showing flow of steam therethrough. Figure 27 is a longitudinal sectional view of the leading edge of the nozzle of Fig. 26, taken substantially upon a plane passing along section line 27—27 on Fig. 26. Figure 28 is an enlarged sectional view of the nozzle portions of the vane of Fig. 24, showing velocity profiles for the gas generator first stage nozzle vane steam nozzle. Figure 29 is a perspective view of a steam-cooled gas generator first-stage rotating blade. Figure 30 is a transverse sectional view of the rotating blade of Fig. 29, taken substantially upon a plane passing along section line 30—30 on Fig. 29, and showing circulation patterns of steam therethrough. Figure 31 is an enlarged view of the front portion of an alternative embodiment of rotating blade incorporating leading edge transpiration steam cooling and thermal barrier steam flow. Figure 32 is a schematicized sectional view of a gas generator incorporating steam cooling. Figure 33 is a graph showing the relation¬ ship of adiabatic expansion efficiency as a function of firing temperature for various methods of cooling. Figure 34 is a graph showing -the specific heat of steam and air as a function of temperature. Figure 35 is a graph describing the abso- lute viscosity of steam and air as a function of temperature. /λr WIF Figure 36 is a graph showing the conductiv¬ ity of steam and air as a function of temperature. Figure 37 is a schematic flow diagram for cooling air in simple-cycle and reheat-cycle gas turbines. Figure 38 is a graph showing power loss as a function of cycle efficiency reduction for simple- cycle gas turbines at different percentages of cooling air used. * Figure 39 is a graph showing power loss as a function of efficiency reduction for reheat gas turbines for different percentages of cooling air used. Figure 40 is a temperature-entropy diagram for the reheat steam turbine cycle with cooling steam applied to a reheat gas turbine. Figure 41 is a cooling steam schematic flow diagram for a reheat steam turbine and reheat gas turbine. Figure 42 is a graph showing cycle effi¬ ciency of cooling steam as a function of the natural logarithm of cycle pressure ratio for various heating temperatures. Figure 43 is a graph showing the net output of cooling steam as a function of the natural loga¬ rithm of cycle pressure ratio for various heating temperatures. Figure 44 is a graph showing cooling steam output components as the function of final steam inlet temperature in the gas turbine with steam turbine output components included. Figure 45 is a graph showing expansion temperatures as a function of the natural logarithm of cycle pressure ratios for a 38 cycle-ratio reheat gas turbine. Figure 46 is a graph showing cycle tempera¬ tures of a reheat gas turbine, and also the cycle temperature of a reheat steam turbine, as a function of the natural logarithm of cycle pressure ratio. Alternate paths of 3 and 7 stages of feedwater he'ating are shown. Figure 47 is a graph showing power turbine blade-path temperature as a function of the natural logarithm of pressure ratio. Detailed Description of the Invention In Fig. 1, gas generator 20 receives air through inlet line 22, producing compressed air by compressor 24, which is driven through shaft 26 by gas generator turbine 28, which is powered by gas produced in gas generator or first combustor 30 from air entering combustor 30 through line 32 and fuel entering combustor 30 through fuel line 34. Reheat or second combustor 36 receives exhaust from gas generator turbine 28 through reheat inlet line 38 and discharges reheated gas through line 40 to power turbine 42, which drives first electric generator 44 directly by shaft 46. Reheat gas output leaves power turbine 42 through exit line 48 and passes into heat exchanger 50 prior to discharge through stack line 52. Exit gases through exit line 48 pass through three stages of heat exchanger 50, the first OM being superheater 54, where superheated steam is produced through superheat line 56; the second being evaporator 58, where water from storage drum 60 and combined recycled water is evaporated in line 62; and the third being economizer 64, where water from storage drum 60 and recycled water from line -66 is evaporated in line 62. In economizer 64, recycled water entering from line 66 is warmed for evaporation in line 59 before entering evapo- rator 58. Superheated steam leaving superheater 54 through line 56 enters and drives steam turbine 68, which directly powers second electric generator 70 through shaft 72. Condensate from steam turbine 68 is formed and collected in condenser 76 and pumped by condenser pump 78 through line 80, along with steam formed directly in turbine 68, through line 82, to heater 86, the output of which is fed by boiler feed pump 88 to recycle line 66. It is to be particularly noted that a tandem arrangement of gas generator and turbine combination can be utilized, with first electric generator 44 being powered by two such arrangements, the reheat exhaust gas output of which feeds into exit line 48. It should further be noted that each of the individual components of the system shown in Fig. 1 with the exception of the reheat combustor is conventional and only the combined arrangement including the reheat combustor leading to the advantages and efficiencies disclosed in the present invention, are intended to be described as new. OMPI /,. WIFO « In Fig. 3, gas generator 20 functions in the same manner as the gas generator described for Fig. 1, and steam turbine 68, condenser 76, heater 86, and pumps 78 and 88 function in the same manner as in Fig. 1. However, the. output of gas generator 20 enters cavity 90 through line 38, cavity 90 being, a combined reheat combustor and superheater, reheat¬ ing exhaust gas from line 38 of gas generator 20 and discharging reheated gas through line 40 to drive power turbine 42, which drives first electric gene¬ rator 44 by means of shaft 46. Cavity 90 also superheats the output from evaporator 58, entering cavity 90 through line 92 and leaving through line 56, for driving steam turbine 68. Reheating of steam also occurs in cavity 90, the steam entering cavity 90 through line 94 and leaving cavity 90 through line 96 to enter steam turbine 68. Steam turbine 68 drives second electric generator 70 through shaft 72, and recycling of condensed output through line 66 occurs through economizer 64 and evaporator 58. Although the reheat feature is optional, Fig. 2 is a representation of gas generator 20 in association with cavity 90 for performing both a reheat and superheating function. Gas generator 20 is made up of combustor 30 and air compressor 24, which has stages 100, 102 and 104, 106 and 108 of a five-stage low pressure section, as well as stages 110 and 112, which are representative stages of a 14-εtage high pressure section. Combustor 30 dis- charges compressed heated air to stages 114 and 116 of a two-stage high pressure section and stage 118 of OMPI " IPO a one-stage low pressure section. Shaft 26 connects compressor 24 with gas generator turbine 28. A number of high temperature and high pressure ratio gas turbines are now on the market, gas generator 20 in Fig. 2 illustrating the second generation LM 5000 Model, other model designations currently available commercially including the LM 2500, JT 9, RB-211, Spey, and the Mars. Gas generator 20 in Fig. 2 is coupled with reheat combustor 36, showing fuel line 35, fuel nozzle 120, annular combustion region 122, diffuser 124, and the combustion region of power turbine 42. It is to be understood that a plurality of fuel nozzles 120, arranged concentrically produce the annular flow of reheat gas which drives power turbine 42 -- Alternatively, the output of gas generator 20 can feed cavity 90 shown in Fig. 4, for performing both a reheat and a superheating function. Air from diffuser 130 passes around struts 132 and is heated in combustion region 134 as fuel nozzle 136 dis¬ charges fuel for combustion therein. Fuel enters cavity 90 through fuel line 138 and is ignited by spark plug 140. Superheat and reheat helical coils 142 are shown somewhat schematically in Fig. 4, having the configuration of Fig. 6 in enlarged detail. Boiler steam header 144 furnishes steam through line 92 to cavity 90, and after boiler steam has traversed its helical path through cavity 90, it leaves cavity 90 through line 56 for collection in boiler steam header 148. Steam leaves cavity 90 through reheat steam outlet line 96, entering reheat steam header 146. Boiler steam enters cavity 90 through line 92 and header 144, while boiler steam exits cavity 90 through line 56 and header 148. Reheat steam enters cavity 90 through reheat steam inlet line 94 from header 150. Insulation 152 surrounds cavity 90, diffuser 130, headers 146, 148, 144 and 150, and also surrounds inlet power turbine nozzle 154. Exit gases pass through inlet nozzle 154 to * drive turbine 42, passing next to power turbine first stage nozzle 156 as shown in Fig. 4. The inside surface 157 of inlet nozzle 154 is in the shape of a nose of a bullet, while outside bell mouth 196 has the shape shown. Fig. 6 shows stream- lined fabricated superheat tube 142 containing perforated sheath 158 surrounding three tubes 160, 162 and 164 for containing surperheated and/or reheated steam, the respective diameters varying according to the flow rates and throughput demanded of each component. The advantages of the cycle of the present invention, such as is accomplished with the combined superheating and reheating apparatus of Figs. 4, 5 and 6, is illustrated by the established and industry accepted data of Figs. 7 and 8, which data are derived from reference No. 14 above, showing the steam path heat balance in Fig. 8 used in the present invention, compared with a conventional reheat steam turbine in Fig. 7. Certain assumptions have been made in developing the data of Figs. 7 and 8, for the sake of concreteness. A tandem arrangement of two gas generator 20 and power turbine 42 arrangement of the type illustrated in Fig. 2, along with a single turbine arrangement is used in conjunction with the data of Fig. 8. Both steam turbine heat balances take into account the boiler feed pump input, but a loss of 3% should be applied to each to take into consideration steam leakage, mechanical losses and the hydrogren-cooled electric generator losses. By calculating the net output, Fig. 7 shows 457 BTU'.s, and by determining the net input from the indicated data, a net input of 1165 BTU's is seen to result, giving a net efficiency of 42.7% and applying an established boiler efficiency of 89% along with the 3% losses given above results in a conventional power plant cycle efficiency of 36.86%. Similarly, from Fig. 8, a gross output of 581 BTU's is obtained from a net input of 1437 BTU's, giving a net efficiency of 40.43% and applying the 3% losses given above results in a cycle efficiency of 39.22% where no boiler efficiency is applicable. Utilizing these data, and making the 15 assumptions listed in Table I, the above results are obtained. TABLE I - Assumptions in Efficiency Calculations Gas Generator and Steam Reheat No. Item Value Units 1. Inlet Pressure 14.7 psia 2. Ambient Temperature 60 degrees F 3. Power Turbine Efficiency 87 percent 4. Reheat Gas Turbine Combustion Loss 3 percent 5. Steam Superheater Combustion Loss 1 percent 6. Steam Reheat Combustion Loss 1 percent 7. Re"heat Combustor Pressure Drop 3 percent 8. Power Turbine & Generator Loss 2 percent 9. Steam Turbine & Generator Loss 3 percent 10. Boiler Radiation & Blowdown Loss 2 percent 11. Constant Reheat Firing Temperature 1800 degrees F 12. Inlet & Back Pressure Loss 4/6 inches of water 13. Condenser Pressure 2 inches Hg absolute 14. Liquid Fuel with 18,400 BTϋ/# LHV 15. Steam Heat Balance, Fig.8. The cycle analysis can be carried out for three modes of operation, namely, continuous, electric base and peak operation, giving the results shown in Table II for half plant capacity. In this arrangement, two 5 model LM 5000 gas generators are mounted to a single hydrogen-cooled electric generator of 100,000 W capacity, the power turbines having oppositely rotating power shafts. The steam turbine is also a 100,000 KW capacity, with the steam output for the two boilers and 10 the two gas turbine superheaters feeding the one steam turbine. OMP - TABLE II - Steam Reheat Cycle Efficiency Gas Gene¬ Reheat Fuel Gas Tur- Steam Total Net rator Fuel Input LflV bine Net Turbine Net Cycle Input LHV BTU/HR Output Net Output Eff % Mode BTU/HR x 10 x 10' KW Output KW KW Continuous 296.24 337.75 43,608 48,234 91,842 49.44 Electric Base 320.16 334.14 46,601 49,313 95,914 50.03 Peak 336.72 328.22 48,750 49,561 98,311 50.46 Results of Table II are to be compared with a similar arrangement in which steam reheat is absent but a reheat gas turbine, such as is shown in the schematic arrangement of Fig. 1, is utilized. Table III presents the assumptions applied in the calculations of the gas turbine reheat cycle efficiency: TABLE III - Assumptions in Efficiency Calculations Gas Generator Reheat Only. No. Item Value Units 1. Ambient Temperature 60 degrees F. 2. Inlet Pressure 14.7 psia 3. Steam Pressure 1250 psig 4. Steam Temperature 900 degrees F. 5. Feed Water Temperature to Boiler 250 degrees F. 6. Steam Turbine Efficiency 80 percent 7. Power Turbine Efficiency 87 percent 8. Reheat Gas Turbine Combustor Loss 3 percent 9. Reheat Combustor Pressure Drop 3 percent 10. Power Turbine _ Generator Loss 2 percent 11. Steam Turbine & Generator Loss 3 percent 12. Boiler Radiation _ Blowdown Loss 2 percent 13. Constant Reheat Firing Tempera¬ ture 1800 degrees F. 14. Inlet & Back Pressure Loss 4/10 inches of water 15. Condenser Pressure 2 inches Hg absolute TABLE IV - Gas Turbine Reheat Cycle Efficiency. Half Plant Capacity Gas Gene¬ Reheat Fuel Gas Tur- Steam Total Net rator Fuel Input LHV bine Net Turbine Net Cycle Input LHV BTU/HR Output Net Output Eff % Mode BTU/HR x 10 ( x 10" KW Output KW KW Continuous 296.24 171.57 43,002 23,442 rt H 66,444 48.69 cr Φ H CO Electric Base 320.16 162.57 45,969 23,814 69,783 49.34 Φ c Φ ti* Peak 336.72 155.90 48,083 23,734 rt 71,817 49.76 3 i o QJ O Φ Hi ω rt o r Hi φ The three modes of operations for which the calculations above are presented are defined by the new International Standard ISO/DIS-3977 for gas turbines, and are as follows. The continuous mode of operation represents continuous or base load operation typical of pipeline service, process applications or electric power generation up to 8700 hours per year with infrequent starting greater than 100 hours per start. The electric base mode of operation refers to longer duration of mid-range electrical power generation, with process applica¬ tions up to 4000 hours per year and with up to eight hours per start. The peak mode of operation refers to short duration intermittent type of operation as occurring typically in electrical power generation for peak load demands, where operation should normally be limited to 500 hours per year and up to two hours per start. OMPI The assumptions in the calculations above and used in the cycle are based on figures for the Model LM 5000 gas generator obtained from the manufacturer after the first second generation unit was thoroughly tested. Characteristics of the LM 5000 gas generator are shown in Table V for base load- ratings of 60°F. and 14.7 pεia inlet pressure. TABLE V - Characteristics of Gas Generator. Base Load Ratings - 60° F. and 14.7 Psia Inlet LM5000 Cycle Pres-sure Ratio P^/ -, 29 Exhaust Flow #/sec. 272 IGHP x IO 3 * 50.00 Cycle Eff. (IGHP)* LHV % 43.0 Firing Temp. F 2,100 Gas Generator Ex Pressure Psia 56.10 Gas Generator Ex Temp. F 1,215 *Isentropic Gas Horsepower for 100% Expansion Efficiency. It is clear from comparison of the results of Table II and Table IV that steam reheat generates an over-all power plant efficiency approaching or exceeding 50% LHV when burning the distillate fuel in the apparatus of Fig. 4. Even without incorpora¬ tion of the superheater in the second combustor, it is possible with an arrangement such as that of Fig. 2 or a process shown in Fig. 1 to obtain net cycle efficiencies approaching or exceeding 49%. These efficiencies represent improvements over the effi¬ ciencies obtainable with use of conventional electric power generating equipment, and can lead to consider¬ able cost savings in actual operation. A simplified economic evaluation for half plant capacity and one power turbine making certain reasonable assumptions as to unit size, fuel cost, and the results of earlier calculations, the economic data of Table VI provide a measure of the degree of commercial success expectable with both the reheat gas turbine H Us TABLE VI - Φ 3 cr Q. Φ r_ - rt ti- Comparative Economic Analysis of Units Utilizing Reheat Cycles, rt m ff rt Cycle % % Fuel Fuel Cost* Savings* Savings* Savings* φ rt us ff Eff. Inc. Consump. $/yr.x $/yr. $ xlO / $xl0 / a as rt H Unit LHV Eff. 10^ 10-$ 5 yrs 5 yrs . ~ φ ti cr σ φ ti- s Conven¬ ti rt tional s vQ Combined us i Cycle 44 Base 100 12.41 Base Base rt C Reheat H cr G.Turb./ ti- S.Turb. ti Φ (Fig. 2) 48.5 10.2 90.7 11.26 1.15 5.75 Base o o Reheat 3 tr G.Turb./ ti- ti Reheat as S.Turb. Table VI indicates that there would be generation of 5.75 million dollars in five years for amortization of the capital cost of the reheat power turbine, while an additional 1.15 million dollars would be made available in five years of operation to amortize the capital cost of a reheat gas turbine combined with a reheat steam turbine system. It should further be noted that a savings in the heat recovery boiler would result, since the boiler would be " simpler and have less surface and greater mean temperature difference. Moreover, the two super¬ heaters would not be duplicated in the heat recovery boiler, effecting even further savings. These figures which are based on the standard industry accepted combined cycle available today having a cycle efficiency of 44% LHV, provide a clear indica¬ tion of the degree of commercial success expectable from the two embodiments of the present invention, but the existence of uncertainties, such as rising fuel costs, limited future oil supply, coal fuel development, high interest rates, power rate in¬ creases, and other uncertainties, could result in a shift favoring the embodiments of the present invention to an even greater extent in the future. OMPI ,. WWIIPPOO Considerations incident to the reheat turbine combustor of the present invention are the following. Combustor 36 incorporates an annular liner 121 flaring outwardly to an exit plane 170 at power turbine 42. Fuel nozzles 120 are disposed in liner 121 and have orifices 123 equally spaced on an orifice plane and separated by a chord width having a chord width diameter with respect to axis 172 of the entire arrangement. Fuel nozzles 120 are directed outwardly along a pitch line forming a pitch angle with respect to axis 172, the intersec¬ tion of the pitch lines with exit plane 170 having a pitch line diameter in a certain preferred ratio to the chord diameter. This ratio is determined by calculating the volume ratio of gases at the exit plane and at the orifice plane by applying Boyle's law, PV=WRT. Under typical conditions, the output gas at the exit plane of the gas generator combustor 30 has a temperature of approximately 2100°F. (1149° C) and a pressure of approximately 30 atmospheres, while the exhaust gas at the orifice plane of the reheat combustor 36 has a temperature of approxi¬ mately 1800°F. (982°C) and 3.75 atmospheres. Accord¬ ingly, a volume ratio of 6.9 is obtained, and this value is preferably increased slightly to produce a lower pressure drop combustor 36. Moreover, the length of diffuser 124 and combustor 36 should be about 60% greater than the diffuser of a conventional power turbine not incorpo- rating combustor 36. Preferably, diffuser 124 and in the alternative apparatus, diffuser 130, has a length of at least about four feet to prevent separation and resulting turbulence of incoming gas. Of course, diffuser 124 or 130 will be constructed of proper angular change and cross-sectional area change to accomplish the above. The power turbine diameter can be slightly greater, and the nozzle chord width adjusted to give approximately 1.35 times the first stage power turbine nozzle area. Referring now to Figs. 4 to 6, the required diameter of cavity 90 is about 10 feet, which takes into account a gas velocity of about 100 feet per second and the increase in area required to accommo¬ date superheater tubes 142, assuming an average gas temperature of about 2000°F. (1093°C). The increase in area to accommodate tubes 142 is about 70%. The length of cavity 90 can vary somewhat according to the surface area required for superheating, and a workable length is about 18 feet. Outer shell 176 is fabricated in sections about two feet long and bolted together at joints 178 so that superheater coils 142 can be inserted properly. The outer shell 176 and associated pipes 56, 92, 94, 96, etc., can be disconnected from headers 144, 146, 148 and 150 at flanges 57, 93, 95, 97, etc., for removal of the two foot sections of outer shell 76. Superheater tubes 142 are wrapped around in a manner to form helical coils of specific configurations, such as that illustrated in Fig. 6. Coils 142 form an annular combustion cavity and control air flow to burner 136. Downstream from burner 136 coils 142 then temper hot combustion gases and distribute them through more coils 142. Combustion cap 155 can also be seen. OMP /,, WVv'IIPF In the arrangement shown in Fig. 5, cold steam enters at the bottom of cavity 90, circulates through tubes 142 and then exits at the top for each superheater. Shell 176 can be center-line supported by supporting mounts 180 to control expansion and to maintain a reasonable center-line elevation with respect to foundation 182 and to satisfy axial alignment of gas generator 20 and power turbine 42. Furthermore, gas generator 20 can be quickly and readily removed to service cavity 90. Superheater coils 142 are individually configured so that each tube length is balanced for steam pressure drop and for equal radiant and convection heat absorption area. Equal pressure drop is needed for flow distribution and the balance of radiation and convection heat absorption provides good superheater temperature control, particularly at partial load conditions. The gas side pressure drop of conventional superheaters is about two inches of water per superheater, amounting to less than a one percent drop for the two superheaters in an atmospheric aspirated boiler. This percent pressure loss can be applied to superheater coils 142 operating at approximately 3.75 atmospheres. Tubes 142 are fabricated with three different sizes 160, 162 and 164, and are streamlined with perforated leading and trailing edge overlays 158 to minimize pressure drop to control the hot gas flow as illus¬ trated in Fig. 6. The perforations allow proper temperature gas to enter and reach the interior surfaces of tubes 160, 162 and 164, thereby increasing considerably the effective heat transfer area. Diffuser 130 from gas generator 20 to cavity 90 has an extra length, and the gas generator assisted by the angle of the struts 132 imparts a swirl to gas passing therethrough to give the gas a longer flow path to assist in diffusion of the gas entering cavity 90. Power turbine 42 has a nozzle consisting of an outer cone or bell mouth 196 and inner bullet- shaped nose 157 to reduce inlet loss from cavity 90 to power turbine 42, where velocity of gas is increased. Data are shown in Table VII comparing a non-reheat steam turbine cycle and a reheat turbine cycle, where pertinent data on the gas generator power turbine and boiler are included. As in previous comparisons, calculations have been made for three modes of operation, namely, continuous, electric base and peak. It is seen that under all modes of operation, the cycle efficiency is improved by use of steam reheat, and total net output is increased substantially. TABLE VII Half Plant Capacity Non-Reheat Steam Turbine Steam Conditions 1250 pisg - 900°F. Reheat Steam Turbine Steam Conditions 2400 psig - 1000/1000°F. Average Gas Generator Figures Presented Continuous Electric Base Peak No STRH STRH No STRH STRH No STRH STRH 1. GG IHP x 10 50.0 1 55.0 59.1 2. GG Firing Temperature °F 2100 1* 2250* 2400* 3. GG Exhaust Flow 4/ Sec . 272 283 289 4. GG Exhaust Temperature °F 1215 1262 1295 5. GG Exhaust Pressure psia 56.10 58.60 61.0 6. GG Fuel Cons.BTU/HR x,10 L HV 296.24 320.16 336.72 7. Fuel Cons.BTU/HR x 10 LHVg 169.55 337.75 162.57 334.14 155.90 328.22 8. Tot.Fuel Cons.BTU/HR x 10 LHV 465.79 633.99 482.73 654.30 492.62664.94 9. PT Ex.Flow to Boiler #/Sec .274.5 277.2 285.4 288.1 291.3 294.0 10. PT Ex. Temp, to Boiler °F 1269 1281 L253 1266 L240 1252 11. PT Expansion Ratio 3. 702 3.867 4.025 12. PT Net Output BTϋ/# Flow 151.3 152.2 155.8 156.5 159.7 160.4 13. Steam Flow #/HR x 10 197.5 292.1 200.5 298.6 199.9 300.1 14. Boiler Stack Temp.°F 328 300 335 300 341 300 15. Air/Fuel Ratio #Air/#Fuel 38.40 28.21 38.42 28.34 38.32 28.39 16. PT Net M Output 43.00 43.61 45.97 46.60 48.08 48.75 17. ST Net MW Output 23.44 48.23 23.81 49.31 23.73 49.56 18. Total Net MW Output 66.44 91.84 69.78 95.91 71.81 98.31 19. Cvcle Efficiency LHV% 48.69 49.44 49.34 50.03 49.76 50.46 ^Estimated Values. Notations: STRH - Steam turbine reheat IHP Isentropic horsepower MW Megawatts GG - Gas generator PT Power Turbine LHV Lower heating value It should be particularly noted that in both embodiments of the present invention, axial flow through the second combustor is present, giving rise to lower pressure loses which translate into higher over-all efficiency. Moreover, in both embodiments, a longer diffuser is possible between the gas generator and second combustor. In the case of the reheat gas combustor, a two-shaft design leads to high efficiency. In the case of the cavity for gas reheat, combined with steam reheat, the long diffuser provides for maximum velocity head recovery. Referring again to Fig. 4, the cylindrical construction of cavity 90 is apparent, leading to equal and even loop, that is, pressure vessel, stress. Coils 190 about the periphery of the inside of cavity 90 define and in effect form the combustion chamber thereof, and inside coils 190 serve as an insulator to shell 176. Moreover, inner coils 142 are arranged about the combustion region 134 to control gas flow and mixing. Plenum 192 serves to stabilize eddy currents before gas enters through inlet nozzle 154 of power turbine 42. Inlet nozzle 154 is configured to reduce pressure loss, through the shape of outside bell mouth 196 and bullet nose 157. Gib keys 198 permit control and maintenance of axial alignment during operation, and flex legs or trunnions (not shown) allow for axial expansion. Annular combustion caps 155 when used in combination with burner fuel nozzles 136 provide for even O /,. WI temperature distribution. The cylindrical con¬ struction of cavity 90, with axial flow of gas therethrough, allows use of a plurality of gas generator and reheat assemblies to permit adaptation to large scale power plants on the order of 1000 megawatt size. In view of the lower intermediate pressure generated in reheat cavity 90, lower grade liquid fuels or solid fuels in pulverized or powdered form ca'n be used in burner nozzles 136, such as low energy content gas from coal, coal derived liquids, shale oil, or other liquid fuels, such as Bunker "C" type oil, residual oil, crude oil or coal powder of relatively poor quality and low energy content per unit volume or weight. Further consequences of the lower intermediate pressure of reheat cavity 90 are lower parasitic compreεion power losses, capability of using larger power turbine blades for better cooling and more rugged service due to lower abra- sion, erosion and corrosion losses, and a larger over-all power turbine for increased expansion efficiency. In the case of burning pulverized or powdered coal in burner 136, diffuser 130 will be equipped as shown in Fig. 19 with circumferential divider 200 to separate incoming gas into an inner stream 202 and an outer stream 204, with outer stream 204 being given centrifugal motion by conven¬ tional streamlined turning vanes 206 to impart a greater spinning motion to outer stream 204. The combustion products from combustion of the coal inside cavity 134 are then directed into the outer spinning stream 208 by the pattern of coils 142 for the solid matter to be collected on the outer circumference of shell 176 into collector 210 by centrifugal force and bled off of plenum 192 through line -212. Combustion gas free of solid matter then enters nozzle 154 as described above. . Fig. 9 shows a map of cycle efficiency as a function of net work output, illustrating the increase of both cycle efficiency and net output with increasing firing temperature, but also showing the dropoff of cycle efficiency while net output increases as reheat proceeds from no reheat to full reheat. In Fig. 9, the subscript designates the pressure ratio, r fi representing a pressure ratio of 6, and r__. representing a pressure ratio of 22, for example. It is important to note that for full reheat, the optimum cycle ratio for net output is also the optimum ratio for cycle efficiency. Both rise to a maximum and then fall off sharply. It is the increase in net work output while maintaining a higher cycle efficiency which gives rise to the advantages of the present invention, utilizing the particular structure taught herein to attain higher firing temperatures of reheated gas. Fig. 10 discloses the effect of firing temperature and cycle pressure ratio, since expansion ratios available for reheat cycles are important in determining combustor size and pressure drop. Intersection of the optimum pressure ratio for efficiency with the curves for three firing tempera¬ tures reveals that at 1600°F. (871°C), the expansion ratio available is only about 2.5, while at 2000°F. (1093°C), the expansion ratio increases to 3.75, and at the still higher firing temperature of 2400°F. (1316°C), the ratio increases further to 5.25. Inasmuch as the power turbine expansion ratio is a measure of the efficiency over-all of a power generating system, and the gas generator 20 described as second generation equipment now available commer¬ cially and firing at 2000 to 2100°F. (1093 to 1149°C) continuous load, the reheat cycle of the present invention is decidedly more practical than with first generation gas generators firing at 1600°F. (871°C), which are seen to produce a rather low power turbine expansion ratio. Fig. 11 shows the effect of another factor to consider in a reheat cycle, namely, the tempera- ture level of gas generator exhaust. Fig. 11 shows the dependence of exit gas generator tempera¬ ture as a function of cycle pressure ratio, showing under conditions of full reheat for three firing temperatures the exit gas temperature at the gas generator in dashed lines, showing the strong dependence on firing temperature and the appreciable dropoff as the cycle ratio is increased. * Dashed lines in Fig. 11 refer to gas generator exit tempera¬ tures, while solid lines for full reheat refer to power turbine exit temperatures. Fig. 11 illustrates that power turbine exit temperatures are not as sensitive to cycle pressure ratios as gas generator exit temperatures. With second generation gas generators, such as that described above in connec- tion with the present invention, the cycle pressure ratios available range from 18 to 30, and the gas generator has a 2000°F. (1093°C) or higher firing temperature, giving exit temperatures within practi¬ cal limits for reheat cycle. It should be noted that when considering heat recovery, the power turbine exhaust temperature is important. The power turbine exhaust temperature ranges in Fig. 11 are ideal for heat recovery in the combined cycle. It is well-known that the best range of exit tempera- tures for heat recovery boilers is between 1200 and 1400°F. (649 and 760°C) . When considering a reheat cycle power turbine, the materials of construction for the exhaust plenum 192 must be carefully considered. High temperature materials, however, are within the state of the art, as is cooling for power turbine blading for the reheat firing temperatures of the present invention. Fig. 12 discloses another factor to be considered, namely, the amount of fuel to be burned in the reheat combustor compared with the first combustor. In general, the reheat combustor requires less fuel than the first combustor. The proper fuel ratio can be seen from Fig. 12 for the range of cycle pressure ratios of 18 to 30 and for firing temperatures of 2000° F. (1093°C) to fall between about 1.2 and 1.8. Firing temperature has a decided effect on combined cycle efficiency and simple cycle efficiency as a function of firing temperature as shown in Fig. 13. Increasing the firing temperature from 160 °F. (871°C) to 2400°F. (1316°C) raises the efficiency from a little over 40% to about 53.5% for the combined simple cycle and from about 45% to a little over 56% for the combined reheat cycle. At 2000°F. (1093°C) firing temperature, the reheat cycle has an efficiency point advantage of about 3.5%, as illustrated in Fig. 13 by the vertical arrows. This represents a 7.4% over the simple cycle efficiency value, which is the equivalent to an increase of about 200°F. (111°C) in firing temperature, as indicated by the horizontal arrows in Fig. 13. This gain is significant and is worthy of note, considering the high cost of fuel today and projected cost in future years. It should be noted that the reheat firing temperature of 1800°F. (982°C) was selected for two basic reasons: first so that lower cost and readily available power turbine blading incorporating so called internal convection-blade cooling can be used instead of more expensive film-cooled blading, and secondly so that the exhaust gases exiting from the power turbine will be at a preferred temperature of about 1250°F. (677°C) for optimum steam generation, that is, minimum stack temperature, and a lower cost of materials and construction required for the power turbine exhaust plenum. In the future, as the state of the art permits gas turbine firing temperatures and pressure ratios to be increased to provide higher power turbine expansion ratios, the reheat firing temperature would be increased to maintain the -optimum 1250°F. (677°C) temperature range to the heat recovery boiler. At such time the more expen¬ sive film-cooled blades would be employed. Such future developments would further increase output and improve cycle efficiency as shown in Figs. 9 through 15. - An over-all picture of the two cycles in terms of cycle efficiency is presented in the left portion of Fig. 15, showing on the left side the simple cycle gas turbine efficiency for the two optimum conditions of compression ratio for output and efficiency. The efficiency rises as the heat is absorbed by the boiler and converted to work, where 100% of the heat obtainable is recovered. At this point, the simple cycle gas turbine is slowly changed to a full reheat cycle by adding increasing amounts of fuel in the second combustor until both firing temperatures are equal. The cycle efficiency rises as the efficiency lines run from left to right. Fig. 14 illustrates the effect on output for the combined cycle as a function of firing temperature. Output increases substantially as firing temperature rises, and moreover, the reheat output has a slightly sharper rise than the simple cycle output. Fig. 14 is useful in evaluating size and potential cost of equipment. For example, with a 2000°F. (1093°C) firing temperature, the combined reheat cycle is shown to develop an increase of about 95 BTU per pound of air flow over the combined simple cycle, equivalent to approximately 56% more output. This figure is an indicator of the relative physical size of the steam and gas turbines and heat recovery boilers of the two cycles and points to potential cost advantages of the reheat cycle. This statement is made in light of the relatively small change in the output ratio of the gas turbine to the steam turbine as shown in Fig. 16 as reheat tempera- ture rises. A further point apparent from Fig. 14 is that the reheat cycle fired at 2000°F. (1093°C) is equivalent to a simple cycle fired at nearly 2600°F. (1427°C) to generate the same output. Incorporating the superheat and reheat functions in the gas turbine reheat combustor 90 shifts the heat load away from the heat recovery boiler economizer 64 and evaporator 58 and makes it possible to provide a simpler and less costly heat recovery boiler with less tube surface area with large temperature differentials between the exhaust gas and the water/steam. A lower cost heat recovery boiler is apparent with less pressure drop possible. Fig. 17 illustrates how this is accomplished by showing three conditions: first for superheating and reheating (condition A), second for superheating only (condition B) and third for no superheating or reheating (condition C) . This significant heat transfer change and the temperature differential advantages of the third case without superheating or reheating for the heat recovery boiler which is part of the invention are illustrated by comparison of the characteristics of the steam, represented by lines D, E, F and G, as well as the gas characteristics in each respective condition represented by lines A, B and C, respectively. Note that condition C is not subjected to the boiler approach temperature, which is near the assumed practical limit of 50°F. (28°C) shown in Fig. 17. Fig. 18 illustrates particality of the disclosed arrangment by showing the advantageous temperature differentials for high heat transfer that takes place in the reheat combustor 90 between the reheat gas and the steam coils 142 which is predicated on the specific optimum cycle pressure ratio selected of about 29 and the associated 2100°F. (1149°C) gas generator firing temperature and the 1800°F. (982°C) reheat firing temperature. It can be noted that the bypass gas entering at about 1250°F. (677°C) has a mean effective tempera- ture difference with respect to the steam of about 400°F. (222°C) at a typical location shown at arrow H, whereas the comouεtor cavity gas exiting at about 2450°F. (1344°C) from cavity 134 has a mean effective temperature difference of about 1200°F. (649°C) as shown at arrow I. The two vertical OMP arrows H and I illustrate these two areas of heat transfer. It should also be noted that the flame in cavity 134 provides a source of radiant heat that further increases the heat flux available to minimize the surface of tubes 142 and thus minimize pressure drop. The considerations outlined in discussion of Figs. 9 to 19 show the value of gas reheat in improving over-all cycle efficiency, thereby demon- strating and substantiating the advantages of the first embodiment of the present invention, providing for gas reheat. The still greater advantages of the second embodiment of the present invention, providing for gas reheat and steam reheat, are apparent from the discussion and comparisons drawn in connection with discussion of the apparatus illustrated in Figs. 4 to 6. Accordingly, the process of the present provides for significantly increased reheat cycle gas turbine output, while the reheat gas turbine efficiency is only slightly degraded over the simple cycle for equal firing temperatures. Moreover, the combined cycle incorporating a reheat gas turbine offers significant cycle efficiency improvements for equal firing temperatures, and the output per unit air flow is significantly greater for the combined cycle utilizing the reheat gas turbine, leading to potential cost savings for such a cycle. The possibility of use of a combined reheat gas turbine and reheat steam turbine cycle, as disclosed above, leads to the prospect of using extracted steam from a conventional steam cycle. Such a cycle can optimize the gas turbine-cycle pressure ratio, while at the same time taking advan- tage of the cold reheat steam generated by the steam turbine for first-stage blade cooling of the gas generator. Extracted steam can also be used for cooling the second-stage gas generator blading and the power turbine blading. Under typical reheat steam turbine steam conditions of 2400 psig-1000/1000°F (16.56 MPa and 538/538°C) , such as the conditions disclosed in applicant's copending Ser. No. 954,838, steam possesses superior physical characteristics as a coolant when substituted for conventionally used air. With use of conventional methods of nozzle and blade casting, together with metallurgical techniques within the present state-of-the-art, base-load firing temperatures of the gas turbine can be increased according to the methods and apparatus of the present invention from the present level of about 2100°F (1149°C) to about 2400°F (1315°C) or higher. These physical characteristics, including specific heat, viscosity, conductivity, radiation absorptivity, and Prandtl number, lead to substantial advantages of steam for cooling, even taking into account the specific volume and sonic velocity. The conventional coolant, air, must be used sparingly and very carefully, primarily due to its effect on cycle degradation with respect to power and efficiency. Nevertheless, despite the fact that a very small :. / f W pressure drop exists between the compressor dis¬ charge and the nozzle and blade, and despite the fact that at higher cycle pressure ratios, the coolant temperature increases to its detriment, highly developed convection and film-cooling tech¬ niques have been developed using air as a coolant, initially for aircraft use, and later for use with industrial gas turbines. Turbine nozzle and rotating blade design components can be steam cooled and thermal barrier protected to achieve about 55% Lower Heating Value (LHV) combined cycle efficiencies according to the teachings which will be hereafter outlined. Fig. 20 shows a gas generator 41A, gas generator blading 43A, a reheat combustor 30A, including fuel line 32A, fuel nozzle 34A, annular combustion region 36A, diffuser 38A, and power turbine combustion region 40A and power turbine blading 42A. Ordinarily, a plurality of fuel nozzles 34A are arranged concentrically to produce the annular flow of reheat gas to drive the power turbine located rightwardly of Fig. 20. Such gas generator and reheat combustor comprises components, such as blading 42A and 43A, which must withstand the temperatures generated by heated gases in annular gas generator combustion region 35A and power turbine combustion region 36A. Accordingly, the temperature of combustion gases obtainable is limited by the ability of internal components to maintain structural integrity and be free from corrosion for prolonged periods of operation. With the present invention, cooling of these components by use of steam permits higher operating temperatures to be attained with minimum cycle degradation and with greater efficiency than if such components were air cooled. An over-all increase in efficiency can result over and above the power plant efficiency approaching or exceeding 50% LHV (Lower Heating Value), as disclosed above. In fact, an efficiency approaching or exceeding 55% LHV is possible with use of steam cooling and thermal barrier protection as taught by the present invention. An over-all picture of reheat cycles in terms of cycle efficiencies is presented in the left portion of Fig. 21, showing the simple cycle gas turbine efficiency on the left side thereof for the two optimum conditions of compression ratio for output and for efficiency. Efficiency rises as the heat is absorbed by the boiler and converted to work. At the point where 100% of the heat obtainable is recovered (vertical axis inside the graph of Fig. 21), a new parameter is plotted along the horizontal axis, namely reheat temperature rise. With no reheat, but 100% conversion of heat to steam, points along the intermediate axis can be read. However, cycle efficiency rises as reheat temperature rises, in the manner shown in the right portion of Fig. 21. Point A in Fig. 21 represents a firing temperature of 2500°F (1371°C) and reheat temperature of 2050°F (1121°C). This temperature is obtainable due to use • STJR ____ -45- of steam cooling. The theoretical efficiency obtainable is shown from Fig. 21 to be 56.7% LHV-as denoted by point A. This curve does not compensate for blade cooling, but summarizes the theoretical maximum obtainable. If steam is used for blade cooling and if the steam can be heated to within 300°-F (167°C) of the firing temperature in the process of blanket cooling or thermal barrier cooling, then the combined cycle efficiency would be slightly degraded to a value of about 56% LHV, and if a safety factor is applied for the steam (to be heated to a lower value of approximately 600°F (333°C)), then 55% LHV can reasonably be selected as a target efficiency value. Considerable incentive exists to maximize blade cooling and thermal barrier protection from a cycle efficiency standpoint by use of steam. If cycle degradation is greater with liquid water as a coolant, then it is possible and likely that higher firing temperatures offer little or no advantage over somewhat lower firing tempera¬ tures. Modern-day gas turbines which fire at elevated temperatures and at higher cycle pressures encounter greater heat flux to the combustor and blading components, including far more radiant heat transfer as well as greater convective heat transfer, With respect to the combustion liner, flame radia¬ tion is a dominant factor. A number of variables are involved, including flame temperature, combus- tion pressure, flame luminosity, flame dimensions, combustor geometry and the emissivity and absorptiv¬ ity of both the gases and the liner. Combustion conditions, such as type of fuel, development of OMPI soot or smoke, and the like, affect the radiation produced from the three primary sources, namely, non-luminous CO,,, non-luminous water vapor and hot glowing sub-micron carbon particles, often described in the art as radiant clouds. Radiant heat transfer for known first-generation 70 psia gas turbine liners have been reported to be upwards of 175,000 BTU per Sq Ft-Hr. A second-generation 25 ratio combustor, on the other hand, has a radiant heat transfer rate of 230,000 to 280,000 BTU per Sq Ft-Hr. At these rates and a 1500°F (833°C) tempera¬ ture differential between the flame and liner, the effective heat transfer coefficient is about 167 BTU per Sq Ft-Hr per °F. Expressed in terms of average firing temperature, this coefficient becomes about 400 BTU per Sq Ft-Hr per °F. Fig. 22 illustrates typical heat flux values, where the total transfer coefficients around vane 48A are resolved into radiation and convection values, which are indepen- dent of each other. Leading edge 50A of vane 48A points toward the combustor flame for most gas turbines and absorbs radiant heat at a level of about 400 BTU per Sq Ft-Hr per °F, the heat flux being related to inlet temperature and not to flame temperature. The leading edge gas film is squeezed to a very thin film due to the stagnation pressure and velocity profile characteristics, which factors cause a very high convection heat transfer rate. The total transfer coefficient is the sum of the two terms, shown on Fig. 22 to be about 740 BTU per Sq Ft-Hr per °F. Heat transfer by radiation varies as the fourth power of the absolute temperature in accordance with Stefan-Boltzmann's Law. Fig. 23 provides values of emissivity and absorptivity as a function of the natural logarithm of pressure in PSIA. These values can be applied to the above mentioned radiation law -to calculate radiant heat flux. The pressure side of vane 48A has a rather constant heat transfer coefficient of about 160 BTU per Sq Ft-Hr per °F, as shown in Fig. 22. However, the suction side develops a hump two-thirds of the way downstream, at which point the boundry layer is thought to thicken and become more turbulent. At this point, the heat transfer coefficient doubles to about 320 BTU per Sq Ft-Hr. per °F. The radiant heat portion of the total heat transfer coefficient is reduced where the blade surface is exposed only to the radiation of the carbon particle cloud and the CO„ and H„0. This portion might be about 25% of the total for a gas temperature of around 2100°F (1149°C). In conventional air-cooled first-stage nozzles for gas turbines constructed according to present technology, many cooling weep holes on the vane surface and leading edge are provided to permit cooling air to exit outwardly and stream over the vane nozzle surface in the form of a thin film. If steam is substituted for air, first-stage nozzle design of the gas generator is modified to the form shown in Figs. 24, 25, 26 and 27. Fig. 24 suggests that leading edge 56A of vane 54A be constructed separately from vane body 58A and allowed to float as leading edge 56A expands radially inwardly from a fixed peripheral anchor point 60A. It should be noted that leading edge 56A receives the greatest heat flux from radiation and convection, and separa¬ tion of leading edge 56A will keep heat away from body 58A of vane 54A. Leading edge weep-holes (not shown in Fig. 24) can additionally be employed on leading edge 56A to assist in absorption of radiant heat. Built-in steam nozzles are incorporated in leading edge 56A, as is best seen in the enlarged view of Fig. 26. Leading edge 56A can be made of solid ceramic material or coated with ceramic thermal heat barrier 62A and body portion 58A has extended surface horizontal fins 64A, best seen in Fig. 27, for convective heat removal. Fins 64A also provide steam-nozzle guides or ducts. Floating loading edge 56A can be made removable from body portion 58A so that replacement or recoating is easy during overhaul of vane 54A. Preferably, steam exits along the surfaces of vane 54A through leading edge nozzles 66A tangentially, or as closely as possible to tangentially, with a very small reactive velocity. The steam exit velocity through nozzles 66A should be equal to or slightly higher than the entrance velocity of the gas stream. An average velocity might be about 700 feet per second (213 m per second) . Steam can also exit through trailing edge nozzle 67A on trailing edge portion 69A and will preferably emerge with a velocity nearly equal to the gas velocity near nozzle 67A. Velocity profiles of the steam and gas are shown in Fig. 28 at the nozzle slot exits. The laminar films are shown on the vane surfaces. The amalgamation of the nozzle steam velocity profile, both at the leading edge nozzle and at the trailing edge nozzle, is apparent. Where the steam and gas meet laminar flow exists for both gases, leading to a healing effect of the wake and a smooth transition of velocity between the two fluids until the steam boundary layer is established. The pressure inside the nozzle cap, made possible by the selection of the gas turbine cycle- pressure ratio of about 38, as well as by using the cold reheat line from the steam turbine, is about 60 psi (.414 MPa) greater than the surrounding gas pressure when allowance is made for the pressure drop in the steam line to the nozzles. This pressure differential then provides an exit velocity of about 700 feet per second. The desired result is for the steam tangential velocity to match the gas stream velocity in order to disturb the flow as little as possible, and in the process, a blanket or boundary layer of steam can be established over the surfaces of vane 54A to serve both as a radiant heat absorbant and a retardant to conduction of heat to the blading. A thin laminar flow is maintained on the surfaces of vane 54A similar to the oil film wedge of a journal bearing. This boundary layer should be stable on the pressure sides of the vane (concave) surface 68A, but on the suction (convex) side 70A, there is a thicken¬ ing of the boundary layer where film effectiveness might be reduced. Therefore, horizontal extended surface fins 72A (best seen in Fig. 25) form trailing edge nozzle 67A for convection cooling. Steam flow to the trailing edge can reasonably be assumed to be approximately 1/4% of the main gas flow. The steam exit velocity, will be about equal to the nozzle exit velocity to reduce exit losses. The steam nozzle edge should be as sharp as practical to minimize eddy current disturbances at tήe nozzle exit . point near orifice 66A. The width of the nozzle should preferably be rather small and is calculated to be about 0.010 inches (.254 mm) for 1-1/2% steam flow and one inch radius between nozzle vanes. However, a 60 psi (.414 MPa) pressure drop should keep the slots blown clean and prevent plugging by ash deposits. Nozzle side-wall cooling is also important when high inlet temperatures are involved. Side slots and weep-holes on both the outer and inner walls are necessary to film-cool the surfaces with steam. The general arrangement shown in Figs. 24 to 27 points to other advantages of steam cooling. The steam can be used to start the gas generator as there is ample flow of steam. Considering a conventional LM 5000 gas generator and 1-1/2% of the gas flow as steam, the rate of steam generation required is about 15,000 lbs. per hour (6804 Kg/hr). At a water feed rate of 20 lbs./KWHr (9.07 Kg/KWHr) the power made available would be about 1005 HP (750KW). With blade-cooling steam, the starting horsepower would be about twice this value. Another advantage of steam cooling is the nozzle and blade-cleaning ability of steam when the gas turbine is shut down. Steam at about 30 psia (.207 MPa) can reach sonic velocity out orifices 56A. Any ash deposit can be cleaned off the blading. A built-in form of steam lance is accordingly present. When steam is applied, the general concept of cooling as suggested for the nozzle vanes can be used for rotating blades, such as blade 80A illus- trated in Fig. 29, in a perspective pictorial view, and shown in section in Fig. 30. Overlapping slots 82A are substituted for the single long slot of the blade of Fig. 24 in order to give blade 80A added strength to withstand centrifugal forces. Slots 86A are used at the trailing edge of blade 80A for the same reason. It is important that the air foil surface be as clean as possible. Air cooling as presently used produces surface roughness which is detrimental, such surface roughening being capable of being reduced considerably, particularly with respect to the leading edge 88A if adequate convective cooling can be realized with steam. Weep-holes (not shown) in leading edge 88A are to be avoided if possible for aerodynamic reasons; however, such holes could be helpful in reducing the radiant heat absorp¬ tion, but in any event should be used sparingly, if at all. A second form of rotating blade adapted for steam cooling is shown in Fig. 31, where another alternative to the high leading edge heat flux involves incorporating transpiration cooling at a sacrifice in surface roughness. Blade 90A shows how such a process can be accomplished using transpiration cooling with steam as the coolant. Steam is metered from plenum chamber 92A to transpiration chambers 94A through small orifices 96A. The steam under controlled pressure is then allowed to flow through porous wire braiding 98A. Transpiration cooling using air as the coolant is presently under active investigation, as is well-known by those skilled in the art, with extensive research and development work presently under way. However, attention has not yet been focused on the advantage of steam for transpiration cooling or to the question of whether extraction steam is purer than extracted air. When appropriately controlled vaporization processes are used, the steam introduced into plenum chamber 92A is purer than air presently used, and deposits which would otherwise form or plug the pores in wire braiding 98A are less with the steam than with the conventional air cooling. Transpiration cooling can also be applied if required to the rear suction side of blade 96A, where greater turbulence and greater heat flux exist, such method again being at the sacrifice of surface roughness and its effect on heat transfer and aerodynamics. The cooling steam is introduced through the turbine wheel in a manner similar to that in which air is introduced with present conventional air cooling techniques. However, in order to obtain a higher pressure inside the blades, a modified approach is desirable, as shown in Fig. 32. Steam enters steam header 102A of gas generator 100A. Gas feed line 104A furnishes combustible gas to fuel nozzle OM 106A, and heated gases are formed in'combustion region 108A to pass through turbine wheel 110A. Cooling steam is introduced through turbine wheel 110A where steam lines are brought through bearing struts 112A to a labyrinth packing or seal 114A at the small diameter of shafting 116A. Steam then flows- internally in annular cavity 118A up to entrance holes 120A of turbine wheel 110A. No internal vanes in the cavity are present in order to minimize pumping losses and steam-head buildup. The blades are preferably sealed off with flexible metallic-type gaskets at the roots of the blades so that steam pressure can be maintained. In addition, power turbine blading also must be cooled in view of the 2000°F (1093°C) inlet temperature of gases contacting the power turbine blading. Although the same approach to cooling can be used as for cooling turbine wheel 110A, the power turbine blades are larger and have a greater surface area, necessitating careful utiliza- tion of steam in order to avoid degrading cycle efficiency. Low-pressure extraction steam is used for this purpose, in a quantity of approximately 1-1/2%. A significant advantage in using steam as a coolant comes in the area of low-cycle fatigue and thermal shock experienced by metal nozzle vane and blade components during start-up. More importantly, however, low-cycle fatigue and thermal shock resist¬ ance are important considerations on full-load trip-out or load dumping from full load to no load. The effect of low-cycle fatigue is perhaps ten times more deleterious on emergency shut-down than on start-up, inasmuch as the compressor discharge air quenches the hot vanes and blades as gas temperature drops suddenly. With use of steam as a coolant where a small start-up boiler and a steam accumulator for shut-down are provided, it is possible to program the cooling steam on start-up and shut-down. Proper control valves and process controllers enable use of such a start-up boiler and steam accumulator. After shut-down, steam is admitted at a programmed rate of decay with built-in time constants to keep the blades under a controlled temperature drop. A further consideration with use of steam concerns hot corrosion (sulfidation) , which becomes more critical at the elevated temperatures encountered in the proposed process. Various coatings have been developed to help protect blade surfaces where corrosive products of combustion come in direct contact with the blades and the protective coatings. If a blanket of steam is maintained over the blade surfaces, contact of corrosive elements is substan¬ tially, if not entirely, avoided at the blade surfaces, Only the heavy particles of corrosive combustion products can penetrate the steam film to reach the blade surfaces because of their mass and momentum. Leading edges 88A or 56A, however, are exposed and protective coatings would be required, such as protective coating 62A shown in Fig. 26. Thermal- barrier ceramics and/or plasma-sprayed coatings can be applied, such as alloys of platinum, chromium, and aluminum. The foregoing approaches and techniques enable full advantage to be taken of the superior OMP properties of steam as a blade coolant and thermal barrier, and enable integration of the coolant with the steam cycle. There are distinct differences in the physical properties of steam and air which affect heat transfer and cooling characteristics of gas turbine vanes, blades and disks. It is necessary to p'rovide for steam cooling due to . the unsuitability of conventional cooling techniques at higher firing temperatures. For example, Fig. 33 provides an indi- cation of the problem faced with current technology where adiabatic expansion efficiency is plotted as a function of firing temperature for various methods of cooling. The curve presents a broad band of efficiency levels,, and shows that there is point of diminishing return in terms of cycle efficiency and gas turbine output with use of air as a coolant, further suggest¬ ing that it is possible that this point has now been reached for industrial gas turbines at about 2100°F (1149°C) . Perhaps the most significant difference between steam and air as a coolant concerns their relative specific heats, or the capacity to absorb heat for a given temperature rise. Fig. 34 graphical¬ ly compares the specific heat of steam with air and shows the distinct difference and advantage of steam over air. While the specific heat of air remains almost constant over a wide range from 500°F to 2400°F (260°C to 1316°C) , the specific heat of steam, besides being about twice that of air, also fluctuates with both pressure and temperature. The specific heat of steam in the region of concern for gas turbine blade cooling varies from about 0.85 at 600 psi (4.14 MPa) and 500°F (260°C) to a value of about 0.63 at 100 psi (0.690 MPa) and 2000°F (1093°C) The specific heat rises from .50 to .60 at lesser pressures for the same temperature range, again reflecting over twice the heat absorbing capacity of air and making steam a far better coolant for gas turbines than air. Further, steam has a lower viscosity than air in the range of pressure and temperature encoun- tered in blade cooling. Fig. 35 shows the relation¬ ship between absolute viscosity and temperature for both steam and air. The viscosity of both air and steam rises with temperature at about the same rate. Although the viscosity of steam is only moderately lower than air, there is an advantage due to friction factors and Reynolds numbers for heat transfer and flow characteristics in the laminar film region and boundary layer near the contact surface, the Reynolds number being inversely proportional to the viscosity. Another characteristic of steam relevant to heat transfer is its thermal conductivity, namely, the ability of heat to be transmitted within or through the fluid itself. Fig. 36 shows the relation between thermal conductivity and temperature for both steam and air. The superior characteristic of steam is particularly apparent at the elevated temperatures encountered in blade cooling, the conductivity of steam being about 50% higher than air at 1500°F (816°C) . Radiation to and from non-luminous gases is another important factor to be considered in heating and cooling of blading, particularly at elevated temperature levels of around 2400°F (1316°C). At such temperatures, 30% or more of the heat transfer is in the form of radiant energy. When a thin layer of relatively cool steam blankets the vanes and blades, the steam is an excellent radiant heat absorber. At elevated temperatures, carbon dioxide and water of combustion, as well as any water present which has been introduced for nitrogen oxides control (both the carbon dioxide and water being present in low partial pressures), the film of pure steam around the vane and blades will absorb much of the combustion radiant heat, and will re-emit this energy at a much lower energy content and heat rate. Much of the radiant heat will be carried away with the continuous flow of steam before it can enter the blades. Cooling air presently used for cooling does not have this characteristic, and almost all of the radiant heat present passes directly into the blades. A further consideration in evaluating heat transfer is the Prandtl number, a higher value indicating significantly better heat transfer, as is well-known in the art. For steam at low pressures, this number drops from 0.98 at 600°F (316°C) to 0.88 at 1200°F (649°C), while for air the corresponding Prandtl number is 0.70 at 600°F (316°C) and 0.65 at 1400°F (760°C) . Other comparisons are also significant in evaluating the relative comparison of steam and air involve the specific volume and sonic velocity relative to piping and passageway sizes and critical area nozzle considerations. These characteristics are given in Table VIII. TABLE VIII Specific Volume Sonic Velocity Cu Ft/Lb 100 psia 600 psia Ft/Sec Temperature °F Air Steam Air Steam Air Steam 400 3 186 4 937 .5310 1431 1764 800 4 667 7 446 .7779 1.189 1717 2113 1200 6 149 9 860 1.025 1.621 1955 2402 I 1600 7 631 12 258 1.272 2.033 2165 2654 o 2000 9 113 14 65 1.519 2.442 2356 2880 I 2400 10 594 17 036 1.766 2.839 2534 3090 SIU Conversion: F = 9/5C+32 Cu Ft/Lb = 62.43 cm /g Ft/sec = .3048 m/ ec It is to be noted that the specific volume for air is less than that of steam for the same pres¬ sure and temperature, but when taking into considera¬ tion the difference in specific heats, the relative volume of steam required is less. For example, at 800°F (427°C) and 600 psi (4.14 MPa), the specific heat- for steam and air are .56 and .26, respectively (see Fig. 34) , and the relative specific volume of steam compared to air is .26 divided by .56 multiplied by * 1.189, giving .5521 cu ft/lb compared to .7779 cu ft/lb for air. The piping cross-sectional area in this example is therefore about 30% less for equal- velocity steam than for air. In order to demonstrate the utility of apparatus and methods heretofore described, reference is made to Fig. 37 and the subsequent analysis, which demonstrates the effect on cycle efficiency of introduction of varying percentages of cooling air for both the simple cycle and reheat cycle, with further analysis comparing the effects of steam cooling . In Fig. 37, schematically showing the flow of cooling air from entrance stream 120A, the cooling air requirements for the first-stage nozzle vanes and Whe « * - 8 th total ( 4) Then the ratio of the change in cycle efficiency, R Re,, ttoo tthhee 'original cycle efficiency is expressed as follows: R E - E. e = ca (5) where E is the original cycle efficiency. Various values of cooling air from 2 to 8% helve been selected and a plot of Gas Turbine Power Loss versus Simple Cycle Efficiency Reduction is presented as Fig. 38. Two different firing tempera- tures of 2000 and 2400°F (1093 and 1316°C) pertinent to this paper have been chosen for the graph. Fig. 38 indicates that power loss is greater than the cycle efficiency loss for a simple cycle air cooled gas turbine. As an example, considering 2000°F (1093°C) inlet temperature and 6% cooling air for an 18 cycle ratio the power loss is shown to be about 8.3% and the efficiency loss 2.6%. In the analysis of the reheat cycle, it can be noted from Fig. 37 that the cooling air to the first and second gas generator stages, that is, flows through lines 122A and 124A, passes through re heat combustor 126A where heat is added and thereafter incremental work ' is developed. A new set of equations can be derived, where the new work develop c ed, wca, after cooling^ air is used,' is expressed as follows: Wwc„a a = Wc(1 - Rca)' + EeRca(hc - h2„s)' + Wpt. - wc which can be reduced to: Wwca = Wpt. + EeRca(hc - h2s) - RcaWc (6) where h is the enthalpy at the compressor discharge temperature h __•__> is the enthalpy at the gas generator exit for constant entropy E is the expansion efficiency of the gas generator turbine , is the work of the power turbine. It should be noted that the first expression of W C{1 - RC3.) is the new expansion work developed by the gas generator where the original work of compression must equal the gas generator work of expansion. Then the ratio of the change in work, Rw, to the orig ~ inal work is the following ^: _ Pt - wca (?) The new cycle efficiency is expressed by the following equation: ca ( F- + F 2> (1 Rca)' Rca(*h1, - h 3 J where F is the Fuel of combustor No. 1 F„ is the Fuel of combustor No. 2 h. is the Enthalpy at the firing temperature of combustor No. 2 h is the Enthalpy of the cooling air at the gas generator exit. Then the ratio of the change in efficiency to the original efficiency is given as: A plot of gas turbine power loss versus reheat cycle efficiency reduction as a function of percent cooling air and cycle pressure ratios is given in Fig. 39. The effect of cooling air on gas turbine output is about half that encountered for the simple cycle. This is explained by the fact that the gas generator cooling air is heated in the second combustor to reduce the effect of the loss of output. Also it can be noted that the change in cycle efficiency is, as can be expected, about the same as for the simple cycle. As an example, considering 6% cooling air for the first-stage gas generator for a 38 ratio cycle, the reheat loss in output is shown to be about 4.2% with a 3.4% efficiency reduction. Fig. 39 is also useful in determining the loss in output and efficiency resulting at the second-stage gas generator due to cooling air for the cycle pressure encountered at the air extraction point. The reduction in fuel consump-tion, F , x can be readily calculated by using the values on Fig, 39 and applying the following formula: = 1 - ( - - ) The power turbine expansion ratios encoun¬ tered are low compared to the over-all high cycle ratios. The cooling air, for this reason, is con¬ sidered to be compressed and then expanded for the sole purpose of cooling; however, the losses of compression and expansion are of concern. For instance, considering the work of compression for a 5.3 ratio to be 90 BTU/Lb, the losses would be about 23 BTU/Lb for 85% and 87% compression and expansion efficiencies respectively. Using 3% cooling air as an example, the loss would be .7 BTU/Lb or .3% of the total work of 206.85 BTU/Lb. This loss would result in a lower efficiency of about .3% also. This illustration shows that the power turbine cooling air does not have much effect on the gas turbine cycle, although it does have a minor effect on the combined cycle efficiency by increasing the gas flow and stack loss. Considering compatible cycle pressure ratios for the optimum simple cycle gas turbine to be 18 and that of the reheat turbine to be 38 for 2400°F (1316°C) initial inlet temperature, data on the various losses can be tabulated and summarized as given in Table IX. O TABLE IX Simple Cycle Reheat Cycle R = 18 T = 2400°F R = 38 T = 2400/2000' ' Power Loss % Eff. Loss % Power Loss % Eff Loss GG 1st Stage 6% CA 7.0 1.8 3.6 2.1 GG 2nd Stage 3% CA 3.5 1.8 1.1 GG 3rd Stage 1% CA 1.0 .2 PT 1st Stage 3% CA .3 TOTAL 11.5 2.7 5.7 3.5 Notations: GG = Gas Generator PT = Power Turbine CA = Cooling Air. The third stage of the simple cycle gas generator with its 18 pressure ratio requires cooling air because of the lower expansion ratio and the higher temperatures encountered whereas the tempera¬ ture drops faster for the higher ratio of 38 for the reheat cycle. Table IX indicates that if there were no cooling air required, the output would increase significantly and also the cycle efficiency would improve appreciably. This point is important and will be used later when considering substituting steam for air as the coolant. It should be noted that the exhaust tempera- ture is about the same or slightly higher. There is no cooling effect; this would raise the temperature but at the same time more work would be extracted which would again lower the temperature. It has been shown that allowing the cooling air to remain in the main air stream will increase both power output and gas turbine cycle efficiency. The effect steam has on over-all combined-cycle performance can now be studied and correlated with air cooling data. The over-all cycle of the steam working fluid in the combined cycle is given in Fig. 40 where cycle temperatures are plotted against cycle entro¬ pies. Tracing the steam turbine cycle, water is pumped from point 130A to point 132A. The water is evaporated from point 132A to point 134A in the heat-recovery boiler and then receives its first superheat from point 134A to point 136A in the second gas turbine combustor. The steam then expands to point 138A providing useful work. The steam is again superheated from point 138A to point 140A in the second combustor before expanding again from point 140A to point 142A. Looking now at the steam cooling for the reheat cycle, steam is extracted at cold reheat point 138A " and is heated in the first stage nozzles and blading to point 144A below the first combustor exit temperature. It can be noted that the gas turbine cycle ratio of 38 allows for 60 to 70 psi steam pressure drop. About 10 psi (68.95 KPa) would be needed to transport the steam to the blading - making about 50 psi (345 KPa) available for cooling which is much more pressure than now realized for air cooling. Extraction steam, not shown in Fig. 40, is used for cooling the second-stage blading. Steam at point 158A is extracted for cooling the first- stage blading of the power turbine where the steam is heated from point 158A to point 152A and expanded to point 154A. It should be noted that the cooling steam previously used for the gas generator is heated in the second combustor all the way to point 148A before expansion. The gas turbine reheat increases the power developed by the gas generator cooling steam. Heat is then recovered in the heat recovery boiler from points 150A and 154A to point 156A, the assumed 300°F (149°C) stack temperature. In accord- ance with the calculated steam cycle heat balance, the heat in the exhaust steam is converted to useful power at an efficiency of 40.43% LHV to complete the cycle. The steam flow can be broken down for purposes of analysis schematically as shown in Fig. 41. The cooling steam is separated from the main stream, a procedure often done in analyzing single and " double automatic extraction/admission steam turbine cycles. In the case of the reheat gas turbine, steam in line 160A is extracted from the cold reheat line 162A and is heated in the first stage gas generator blading - designated as heat exchanger 162A. Incremental fuel is introduced in line 164A and burned in combustor 166A to heat this steam in line 168A of heat exchanger 170A as it passes through and around the blading as will be shown later. The main fuel for the gas turbine enters through inlet line 172A. The cooling steam, after expanding through gas generator turbine 174A, is again heated in combustor 176A at heat exchanger 178A,the heat being supplied in line 180A. Fuel for the main steam stream enters from line 182A; the cooling steam for the power turbine from line 184A and the main gas turbine air flow from line 186A. The heated gas generator turbine cooling steam then expands through power turbine 188A to produce power after which the steam is exhausted to aseudo heat exchanger 190A, where heat is extracted to heat the feed water line 192A. The final stack temperature of 300°F (149°C) is assumed to be compat¬ ible with the main heat-recovery boiler stack tempera¬ ture. Cooling steam for the power turbine in line 194A is extracted from low-pressure steam turbine 196A and is heated by power turbine first- stage blading designated as heat exhanger 198A. The steam then continues to expand to the exhaust where it is also recovered in pseudo heat exchanger 190A. Extraction steam that flows in lines 200A, 202A and 204A is used to heat the feed water from 101°F (38.3°C) to 240°F (116°C). Boiler feed pump 206A raises the temperature to 250°F (121°C). It is assumed that the make-up water arrives at an enthalpy (h) of 70 to be compatible with the heat balance cycle of Fig. 41. The total cooling steam is now heated in another pseudo heat exchanger 208A by fuel from line 210A to bring the stream back to the initial 2415 psia (16.55 MPa) 1000°F (538°C) condition. Heat exchanger 208A is necessary to determine the fuel required to complete the loop so that the cooling steam-cycle efficiency can be determined and evaluated. Fig. 41 can also be used for the simple- cycle gas turbine by eliminating combustor 176A and by extracting the steam at the appropriate steam pressure to allow for adequate injection pressure and piping pressure drop. Equations to correspond with Fig. 41 can now be presented for the re-heat cycle. The fuel input equation is as follows: F t " f l-2 + f 2-2 + f 2-4 + f 2-5 + f 3 (11) where F is the total fuel added, , ______ 2 i s fuel added at line 164A, f 2 _ 2 at line 184A, f-_ 4 at line 180A, f 2 _ 5 at line 212A, and f_ at line 210A. Fuel f -i represents the fuel required for the main steam stream at line 182A and fuel -- -_ - represents the fuel to heat the gas turbine airflow at line 186A and do not enter into the equation. The work output equation is expressed by giving work components for both the gas turbine and the steam turbine. Wt. = (Wgg 4 t. + pt. )' + ( h,pt. + i.pt, + Wl,pt. )' ( v 12)' where is the total work Wggt, is the g r as g 3 enerator work pt, is the power turbine work W. . is the high-pressure steam turbine work i.pt. is the intermediate steam turbine work Wl-pt, is the low-pressure steam turbine work, The cycle efficiency is expressed by the following general equation: . E s = __ ( 13 ) F t where E is the cycle efficiency of the steam flow LHV. It can be noted that 2% leakage at line 214A is not used in the heat balance. Thus a 2% safety factor is introduced to take into account losses in the heat recovery boiler, leakage, and the generator. Also the adiabatic expansion efficiency is assumed to be 86% instead of 87% to introduce another safety factor. Standard gas tables have been used for data relating to the expansion of the steam. Certain new steam tables extend to only 1500°F (816°C). However, to obtain close heat balances, the steam tables were used in exchangers 190A and 208A for enthalpies below 1500°F (816°C) and the standard gas tables for water vapor were not used for lower temperatures. Close heat balances were obtained using this approach. Equations (11) and (12) can be used for the simp ^ le- cycle gas turbine by extracting steam at the proper pressure and using the proper steam enthalpies. The results of the above procedures are presented in Fig. 42 for cycle efficiency and Fig. 43 for output. The efficiency and output figures have been plotted against the natural logarithm of the cycle pressure ratio to reduce the abscissa for readability and to make the plots compatible with further graphs to be presented. Two reheat cycles were investigated: cycle ratio of 26 with 2000/1800°F (1093/982°C) firing temperatures and cycle ratio 38 with 2400/2000°F (1316/1093°C) firing temperatures. Fig. 42 indicates that the over-all cycle efficiency of the first-stage gas generator cooling steam for the reheat cycle is quite high and is about 45% LHV, and depends upon the temperature level to which the steam is heated. Considering that the steam is heated to 2000°F (1093°C) the cycle efficien¬ cy is still approximately 44% LHV. Cycle efficiencies for the 26 cycle ratio are shown to range from 39% to 42% LHV. -^URE O.VIPI Λ WIPO « The power turbine cooling-steam cycle efficiencies for the two reheat cycles cited above (considering power turbine expansion ratios of 5.302 and 3.721 respectively) are also shown in Fig. 42. The efficiencies for the 5.302 expansion ratio (ER) range from 36% to 39% LHV and the efficiencies for the -3.721 ER vary from 35% to 38% LHV. These values, coupled with the cycle improvements available for eliminating the air cooling, can be used to arrive at an over-all effect on combined-cycle efficiency which will be discussed later. Fig. 43 shows net output in terms of BTU/Lb of steam for the various two sets of conditions. The output per pound of steam ranges from a low of 550 to a high of 825 BTU/Lb. Considerably more work per pound of steam is obtained by the cooling. The steam that expands through the steam turbine is 581 BTU/Lb. This averages about 31% greater specific work and ultimately reduces water rejection to the atmosphere whether indirectly from the condenser or the stack. Figs. 42 and 43 include curves for the simple-cycle gas turbine. Considering an 18 cycle ratio with 2400°F (1316°C) inlet temperatures, the cycle efficiencies range from 36% to 43% LHV - depending upon the temperature level to which the cooling steam is heated. The outputs vary from 570 to 720 depending upon the steam-heating effect. As an indication of the breakdown in the output, work components made available are plotted -2 against the final steam inlet temperature in Fig. 44. The various components can be added together to give the total output per pound of cooling steam. Then by properly determining realistic cooling steam-flow values based on specific heat ratios, temperatures, other factors cited in Figs. 34 through 36, and known cooling-air requirements, the total output can be determined. Likewise, an over-all cycle-efficiency level for the calculated cooling steam flows can be determined as will be presented later. Figs. 45 and 46 give an insight to blade- path temperatures for both steam and air. These curves also show how the extraction steam of the steam turbine can be integrated with the gas turbine. Fig. 45 shows a plot of temperature versus the natural logarithm of the cycle pressure ratio of the 38 cycle-ratio gas turbine. The natural loga¬ rithm is used as the abscissa. The expansion lines therefore are represented as straight lines and equal spacing can be made for the various gas turbine stages as shown. Steam expansion lines, shown dashed, are at a lesser slope than those of the gas. This curve is useful in determining what happens to the exhaust mixture of gas and steam in terms of the mixture temperature. Fig. 46 is a plot of both the ' gas turbine and steam turbine temperatures versus the natural logarithm of the two cycle-pressure ratios. Note that 14.7 psia has been used for Pa and that the steam turbine expansion is inverted to the right OMPI where the logarithm values are negative for P^/F ratios less than one. Seven successive heaters are plotted in Fig. 46 from left to right to show at what pressure they fit in the cycle. Also, the steam reheat point and the crossover point for the steam turbine are shown on the graph and illustrate quite clearly why it is necessary to properly integrate the gas turbine pressure with the steam cycle. Evaluation of the effect of substituting steam cooling for air cooling requires consideration of change in power turbine exit temperature. Fig. 45 has already shown the effect steam has on exhaust temperature. However, the data already presented on evaluating steam takes into account the flow and exhaust temperature and can be considered as a separate stream. Moreover, the elimination of cooling air must be considered and dealt with. When cooling air is eliminated, the power turbine exhaust tempera- ture will rise because the cooling air bypasses the combustor and tempers the hot gas stream as it passes through the blading. Fig. 47 shows this effect. Considering 400% theoretical air, 87% expansion efficiency, and 3% cooling air, the exhaust tempera- ture is calculated to be 32°F (17.8°C) lower. Conversely, removing the cooling air will raise the exhaust temperature 32°F (17.8°C) equivalent to 9.1 BTU/Lb (21.2 KJ/Kg) of air. This heat is recoverable as work at 40% steam-cycle efficiency. If the gas turbine is fired 45°F (25°C) hotter, this extra heat then balances the heat needed to compensate for the cooling air as can be noted from Fig. 47. ^&U O Considering 4% cooling air, the exhaust temperature calculates to rise 42°F (23.9°C) and the firing temperature 60°F (33.3°C). Table X presents pertinent data which has been used in the calculations on the 38 cycle-ratio reheat gas turbine firing at 2400/2000°F (1316/ 1093-°C) . Table X 8 Cycle Ratio - 2400/2000°F Reheat Gas Turbine Data 1. Inlet Press PSIA 14.70 2. Inlet Temp °F 60 3. Cycle Press Ratio 38 4. Compression Eff % 85 5. Work of Comp BTU/Lb Air 265.68 6. No.l Comb Firing Temp °F 2400 7. No.l Comb Press Loss % 4 8. GG Inlet Press PSIA 536.26 9. GG Exit Press PSIA 81.44 10. GG Exit Temp °F 1514 11. GG Exp Ratio 6.580 12. GG and PT Exp Eff % 87 13. GG Theo Air % 400 14. No. 2 Comb Firing Temp °F 2000 15. No. 2 Comb Press Loss % 3 16. PT Inlet Press PSIA 78.997 17. PT Outlet Press PSIA 14.9 18. PT Theo Air % 400 19. PT Exp Ratio 5.302 20. PT Exit Temp °F 1297 21. PT Net Work BTU/Lb Air 206.85 22. Air Leakage % .5 23. Comb Eff % .97 24. No.l Comb Fuel BTU/Lb Air 370.97 25. No.2 Comb Fuel BTU/Lb Air 145.53 26. Total Fuel- BTU/Lb Air 516.50 27. Cycle Eff % LHV 40.05 Notatons: GG - Gas Generator PT - Power Turbine LHV - Lower Heating Value OMPi / J wδr Using the data developed and curves pre¬ sented, Table XI summarizes the effect steam substitu¬ tion has on over-all combined-cycle efficiency. This table indicates that the over-all combined cycle will be slightly degraded by about 0.7% (0.4% Points). It is doubtful, however, if 2400/2000°F (1316/1093°C) levels for base-load operation can be accomplished without resorting to steam cooling or some other type of cooling beyond what is currently in use today. As will be explained in the summary, the assumed over-all combined cycle efficiency is deter¬ mined to be 56% LHV based on the data developed above when considering 2500/2050°F (1371/ 1121°C) firing temperatures to reflect incremental fuel added to heat the cooling steam. A significant increase in output of 36.2% can be noted when compared to the original gas turbine work. This calculates to be 6% increase for each 1% steam. This added bonus should improve the installed cost per unit of power output of the combined cycle. TABLE XI Coolant Without Cooling Air Steam Cooling Air Steam Work Fuel Work Fuel % % BTU/Lb Air BTU/Lb Air BTU/Lb Air BTU/Lb Air First-Stage GG GT 10.59 10.81 21. 16 60.60 ST 6. 17 Second-Stage GG 4 1.5 GT 4.01 4.29 9.66 28.85 ST 3.09 First-Stage PT 4 1.5 GT .80 4.50 27.34 ST 4.52* 5.89 Sub-Totals 16.0 6.0 19.92 15.1 50.47 116.79 Cycle Eff % 131. .9* 43. .21 Change in PT Exit Temp °F 42 Original GT Work BTU/Lb Air 206.85 Credit for Exit Temp Inc in GT Work % 24.62 ' 40% work BTU/Lb 4.52* Inc in Total Work(Vs GT Work)% 33.64 Increase in GT Work BTU/Lb Air 50.72 Original Comb Cycle Eff % 56.00 Increase in ST Work BTU/Lb Air 19.67 Over-all Cycle Eff 7. LHV Total Increase in Work BTU/LB 70.39 50% GT & 50% ST 55.62 Total Fuel BTU/Lb Air 131.89 Overall Cycle Eff % LHV Over-all Eff of Inc Work % LHV 53.37 60% GT _ 40% ST 55.56 *-Assuming steam being heated to 300°F of the 1st and 2nd combustor temperatures. This approach is not very critical and does not change the over-all cycle efficiency very greatly. Notations: GG = Gas Generator GT = Gas Turbine PT = Power Turbine ST = Steam Turbine Gas Turbine Cyucle Pressure Ratio = 38 First Combustor Firing Temp °F = 2500 (Includes fuel added for steam heating) Second Combustor Firing Temp °F = 2050 (Includes fuel added for steam heating). -$U E . __OMPI v Λ,. " ~WΪPθ " . There could be another penalty and considera¬ tion in using steam as a blade coolant in that about 20% (of the steam flow) make-up could be required to replace the loss in steam out the stack. Abbreviations used throughout the applica¬ tion are as follows: HP - Horsepower MPa - Megapascals ER - Expansion Ratio LHV - Lower Heating Value h - Enthalpy. SIU conversion factors are as follows: BTU = 1.055 KJ BTU/Lb = 2.326 KJ/Kg PSI = 6,895 KPa F = 9/5C + 32. Throughout the specification and claims, unless otherwise specified, cycle efficiencies are expressed in terms of fuel lower heating value abbre¬ viated LHV) , temperatures are given in degees Fahrenheit, pressures in pounds per square inch absolute, costs in U.S. dollars, power in kilowatts (KW) , energy in British Thermal Units (BTU), and parts and proportions in percent by weight. The foregoing is considered as illustrative only of the principles of the invention. Further, since numerous modifications and changes will readily OMPI occur to those skilled in the art, it is not desired to limit the invention to the exact construction and operation shown and described, and accordingly, all suitable modifica tions and equivalents may be resorted to, falling within the scope of the invention. Similarly, as the state of the art of gas turbines advances through improved higher temperature metallurgy, higher temperature construction materials, higher firing temperatures, higher compressor efficiency, higher turbine efficiency, and higher compressor ratios, the general temperature and pressure relationships between the first combustor, second combustor, gas generator and power turbine are considered to move upward accordingly from the levels presented heretofore and fall within the scope of the invention.

Claims: CLAIMS 1. A process for production of useful power compris¬ ing the following steps: (a) generating in a gas generator means a high temperature, high pressure first gas from ambient air and gas generator fuel; (b) reheating said first gas in a reheat combus¬ tion cavity by injection and ignition of reheater fuel to form a high temperature, high pressure second gas; (c) driving with said second gas a power turbine for generation of useful power in a first generat¬ ing means and production of an exhaust third gas; (d) exchanging heat in heat exchange means from said third gas with liquid water for generation of input steam for production of useful power in a second generating means by a steam turbine. 2. The process of claim 1 wherein said heat exchange means comprises a superheater, an evaporator, and an economizer, the superheater receiving said third gas for exchange of heat therefrom and production of a fourth gas, the evaporator receiving said fourth gas for exchange of heat therefrom to produce a fifth gas, and the econo- mizer receiving said fifth gas for exchange of heat therefrom and discharge into ambient surround¬ ings, the economizer receiving liquid water condensate from said steam turbine, receiving heat from said fifth gas and producing warmed y --- r V. : -- condensate for passing to said evaporator, the evaporator receiving heat from said fourth gas for converting the water from said economizer to an evaporator steam, the superheater receiving said evaporator steam and transferring heat from said third gas to superheated steam for discharge to said steam turbine for production of useful power in the second generating means. 3. The process of claim 2 wherein said reheater fuel is a lower grade fuel than said gas generator fuel, and is selected from the group consisting of coal gas, coal derived liquid, shale derived liquid, Bunker "C" type oil, residual oil, crude oil, and powdered coal. 4. The process of claim 3 wherein said first generat¬ ing means is a first electric generator and said second generating means is a second electric generator. 5. The process of claim 3 wherein said first generat- ing means includes a pair of oppositely extending, oppositely rotating axial driving shafts, each driving shaft being drivingly connected to one of said power turbines. 6. The process of claim 2 wherein said gas generator means comprises a generator turbine drivingly connected by a generator shaft to a turbine gas compressor, said compressor receiving said ambient air for producing high pressure air and discharge thereof into a first combustor, said first combustor receiving said gas generator fuel and forming a combustion gas for driving said generator turbine, said generator turbine driving said compressor and forming said first gas for discharge to said reheat combustor. 7. The process of claim 6 wherein said gas generator fuel and said reheater fuel are common and are supplied from a common source. 8. The process of claim 6 wherein said compressor comprises a five-stage low pressure section followed by a 14-stage high pressure section and said gas generator turbine comprises a two-stage high pressure section driving the 14-stage compressor followed by a one-stage low pressure section driving the five-stage compressor and wherein said power turbine comprises an axial flow annular reheat combustor wherein said reheater fuel is injected thereinto through a plurality of axially symmetric fuel nozzles. 9. The process of claim 6 wherein said first combus¬ tor is operated with a rich fuel mixture and said reheat combustor is operated with a lean fuel mixture to provide control of nitrogen oxide pollutant emissions. 10. The process of claim 6 wherein said first combustor is operated with a lean fuel mixture and said reheat combustor is run with a rich fuel mixture to control nitrogen oxide pollutant emissions. -^U Λ 11. The process of claim 6 wherein water or steam is injected into said reheat combustor to control nitrogen oxide pollutant emissions. 12. The process of claim 6 wherein said first combustor generates a temperature of at least • about 2100°F. . (1149°C) and a pressure ratio of said first gas to said ambient air of about 29. 29. 13. The process of claim 12 wherein the temperature of said second gas is about 1800°F. (982°C). 14. The process of claim 1 wherein step (b) is carried out in a combustion cavity wherein said input steam generated in step (d) is superheated for production of useful power in said second generating means by driving said steam turbine, and wherein said liquid water for generation of said input steam is liquid water condensate from said steam turbine. 15. The process of claim 14 wherein said heat exchange means comprises an evaporator and an economizer, said evaporator receiving said third gas for transfer of heat therefrom to produce a fourth gas, said economizer receiving said fourth gas for extraction of heat therefrom to produce a discharge gas, said economizer receiving liquid water condensate from said steam turbine and receiving heat from said fifth gas for warming said water and discharging said warmed water to said evaporator, for receiving heat from said fourth gas for producing input steam from said warmed water, said input steam being discharged to said cavity. 16. The process of claim 15 wherein said gas generator means comprises a generator turbine drivingly connected by a generator shaft to a turbine gas compressor, said compressor receiving said ambient air for producing high pressure air and discharge thereof into a first combustor, said first combustor receiving said gas generator fuel and forming a combustion gas for driving said generator turbine, said generator turbine driving said compressor and forming said first gas for discharge to said reheat combustor. 17. The process of claim 16 wherein the firing temperature of said first combustor to produce said first gas is at least about 2100°F. (1149°C) and the pressure ratio of said first gas to said ambient air is about 29. 18. The process of claim 17 wherein said second gas has a temperature of about 1800°F. (982°C). 19. The process of claim 15 wherein said steam turbine generates an exhaust steam which is passed into said combustor cavity for reheating, producing a reheated steam which is returned to said steam turbine for assisting in driving thereof. 20. The process of claim 15 wherein said steam turbine produces exhaust steam which is passed into said combustion cavity for reheating in helical coils and returned therefrom to said steam turbine for assisting in driving thereof. 21. The process of claim 14 wherein steps (a), (b) and (c) are carried out utilizing said steam as a coolant for internal components of said gas generator, combustion cavity, and power turbine. 22. The process of claim 21 wherein said internal components are blanketed with a flow of steam whereby said components are protectively insu lated against radiant and convection heating. 23. The process of claim 14 wherein said first gas and second gas are characterized by a predeter¬ mined pressure ratio such that over-all efficiency is maximized. 24. The process of claim 23 wherein said gas turbine cycle pressure ratio is about 38, said first gas is produced at a temperature of about 2400°F and is reheated in said combustion cavity to about 2000°F to form said second gas. 25. ' The process of claim 24 wherein said power turbine includes nozzles cooled by coolant steam introduced through a steam cold reheat line and wherein said coolant steam is supplied at steam extraction points along the combined reheat cycle. 26. The process of claim 23 wherein the first gas produced by said gas generator is at least about 2500°F, said internal components of the gas generator, combustion cavity and power turbine being steam cooled so as to result in less than 1% loss in cycle efficiency. 27. The process of claim 14 wherein said power turbine is steam-cooled and comprises stationary vanes and rotating blades, each vane having a leading edge portion for receiving leading ledge steam and a trailing edge portion connected to the leading edge portion for receiving trailing edge steam wherein the process is started by the following steps: 17- (a) introducing the leading edge steam into the leading edge portion of the vane and introducing the trailing edge steam onto the trailing edge portion of the vane, thereby allowing steam to strike the blades to heat the blades and rotate the blades; (b) feeding combustion air onto the power turbine; (c) injecting combustion fuel and igniting the fuel; and (d) accelerating the blades to full operating speed. 28. In the production of useful power by gas reheating of an input gas, the improvement comprising a reheat combustion cavity having a shell and a plurality of axially symmetrically disposed fuel nozzles for gas reheating of an input first gas to form a reheated second gas. 29. The improvement of claim 28 wherein an annular liner has a plurality of fuel nozzles disposed therein and flaring outwardly to an exit plane at a gas powered turbine, the nozzles having orifies equally spaced on an orifice plane and separated by a chord idth having a chord width diameter with respect to said axis, the nozzles being directed outwardly along a pitch line forming a pitch angle with respect to said axis, the intersection of said pitch lines with said exit plane having a pitch line diameter with respect to said axis, the ratio of said pitch line diameter to said chord diameter being approximately proportional to the volume ratio calculated from Boyle's law of said output gas at said exit plane to said exhaust gas at said orifice plane. 30. The improvement of claim 29 wherein ' said volume ratio is about 6.9 or slightly greater than 6.9 with respect to the gas generator combustion cavity. 31. The improvement of claim 30 wherein said gas generator is removably connected to said reheat combustion cavity by a diffuser, the diffuser having a length of at least about four feet, and wherein said combustion cavity is made up of a plurality of sections removably attached together. 32. The improvement of claim 28 wherein said reheat combustion cavity is adapted for superheating input steam, the combustion cavity having a substantially cylindrical shell, helical heat exchange coils, a pair of steam headers outside said shell and a cavity supporting means, the coils being arranged inside the shell for conducting an input steam from one of said headers through a reheated gas axially flowing through said shell, the shell receiving input gas annularly for reheating by said fuel nozzles, V " «IFO- the axially symmetrically disposed fuel nozzles annularly discharging fuel in the direction of said input gas and throwing said reheated gas produced from said fuel and said input gas expandingly into said shell to direct said reheated gas upon said helical coils, the helical coils being arranged with diameters and axial spacing to form a combustion region fee of helical coils and a surrounding region of spaced helical coils. 33. The improvement of claim 32 wherein each of said helical coils comprises a perforated overlay and three tubes of different diameters, the overlay being streamlined with perforated leading and trailing edges to control the hot gas flow therethrough and around said tubes and wherein each of said fuel nozzles has a combustion cap for even temperature distribution of said combustion gases. 34. The improvement of claim 33 wherein said useful power is generated by means of a plurality of said power turbines in a power plant having a capacity of at least 1000 megawatts. 35. The improvement of claim 33 wherein said steam headers are cylindrical and extend substantially along the axial length of said combustion cavity, the headers comprising an inlet header and an oppositely disposed outlet header, the inlet header receiving steam for superheating and discharging said steam to said combustor cavity, the outlet header receiving superheated steam from said combustion cavity for use in generation of said useful power. 36. The improvement of claim 35 wherein said combustion cavity includes helical coils for receiving exhaust steam from a steam turbine, reheating steam within said helical coils, and discharging reheated steam to said steam turbine for assist- ing in generation of power therein and wherein a second pair of steam headers is disposed outside said combustion cavity for introducing said exhaust steam into said combustion cavity and for receiving reheated steam from said combustion cavity for return to said steam turbine, said second pair of steam headers being oppositely disposed in a plane passing through the axis of said combustion cavity and perpendicular to the plane of said first steam headers. 37. The improvement of claim 36 wherein said cylindri¬ cal shell comprises a plurality of sections removably attached together and is constructed of cylindrical sections to facilitate assembly and tube insertions therewithin. 38. The improvement of claim 36 wherein said cylindri¬ cal shell sections and associated internal tubing can be disconnected by means of outer tubing flanges for assembly and disassembly OM and wherein said cylindrical shell has a plenum on the exit side thereof for stabilizing eddy currents and for discharge of reheated gas into a power turbine for generating useful power. 39. The improvement of claim 38 wherein said power turbine has an inlet nozzle having an outside bell mouth and an inside bullet-shaped nose, wherein said combustion cavity has horizontal center line supporting means for control of expansion and alignment of said shell and wherein said cylindrical shell has gib keys on the bottom thereof to maintain axial alignment of said combustion cavity, and said cylindrical shell has flex legs or trunnions to allow for axial expansion of said combustor cavity. 40. The improvement of claim 39 wherein the gas discharged from said combustion cavity has a temperature of about 1800°F. (982°C). 41. The improvement of claim 32 wherein said combus¬ tion cavity is removably connected to a gas generator by a diffuser to transmit incoming gas to the combustion cavity, wherein said diffuser has a length of at least about four feet and wherein said diffuser and combustion cavity together have a length of at least about 18 feet. 42. The improvement of claim 41 wherein a powdered coal fuel is introduced through said fuel nozzles, for burning in said combustion cavity, said diffuser having a plurality of turning vanes and a circumferential divider to separate said incoming gas into an inner stream and an ' outer stream, the outer stream being given sufficient centrifugal motion by said turning vanes to permit collection on the outer circum- ference of said combustion cavity of solid matter produced from combustion of said fuel. 43. The improvement of claim 28 wherein a gas generator means produces said input gas, the gas generator means having at least one nozzle vane comprising a leading edge portion and a body portion wherein the leading edge portion is attached at a fixed periphery anchor point to the body portion and is expandable radially inwardly therefrom, the leading edge portion being separated from the body portion to form a cavity having a leading edge nozzle from which a first supply of steam blankets the body portion for protective blanketing thereof, whereby separation of the leading edge portion from the body portion and said steam keeps heat away from the body of the vane. 44. The improvement of claim 43 wherein the body portion has a trailng edge portion including a plurality of fins and ducts forming a trailing edge nozzle from which a second supply of steam emerges, the second supply of steam serving to absorb heat from the suction side of the trailing edge portion and wherein said leading edge portion is provided with a nozzle cap comprising a protec¬ tive coating selected from the group consisting of thermal barrier ceramics and plasma-sprayed metal alloys and wherein said nozzle cap is a plasma- sprayed alloy of platinum, chromium, aluminum, or mixtures thereof. 45. The improvement of claim 44 wherein said leading edge portion is provided with weep holes for allowing steam to issue therefrom, whereby the influx of radiant heat on the leading edge portion is substantially reduced and wherein said leading edge portion is removable and replaceable and is provided with a woven wire covering for steam transpiration cooling of the leading edge portion, and wherein said leading edge portion and said body portion are provided with a smooth surface, whereby aerodyna¬ mic efficiency is improved. 46. The improvement of claim 44 wherein the separa¬ tion between said leading edge portion and said body portion defines a steam nozzle for impart¬ ing a tangential component to the velocity of steam issuing therefrom and wherein said steam nozzles are in the form of two rows of alternately staggered slots. OMPI ° 47. The improvement of claim 44 wherein said nozzle vane is provided with internal con¬ vection cooling fins and steam nozzle ducts for introducing steam into said nozzle vane and conducting said steam to said orifices and wherein said steam emerges at high speed for sonic cleaning of said body portion when start¬ ing the gas generator. 48. The improvement of claim 44 wherein the leading edge nozzle directs the first supply of steam tangentially to the body portion, and wherein the trailing edge nozzle directs the second supply of steam parallel to the gas flow along the trailing edge portion, whereby the first supply of steam is dischargeable at approximately the same velocity as the main gas stream, and the second supply of steam is dischargeable at approximately the same velocity as the exit gas stream velocity, and whereby eddy currents and leaving losses are minimized. 49. The improvement of claim 28 wherein a gas generator means produces said input first gas and a power turbine is deiven by said reheated second gas, the gas generator means or power turbine having a rotating blade comprising a head portion and a body portion, the head portion having an internal steam cavity for conducting steam to the outer surface of the OMPI rotating blade through nozzles on the surface, the body portion having a trailing edge portion, whereby the blade is blanketed with steam for protection against convective and radiative absorption of heat and whereby the body portion has a trailing edge portion which contains fins • and ducts forming said nozzles for cooling and accelerating cooling steam. 50 * . The improvement of claim 49 wherein said nozzles are in the form of steam nozzles for imparting a tangential component to the steam velocity issuing therefrom, wherein said nozzles comprise a pair of rows of parallel staggered slots for steam blanketing of the body portion of the rotating blade, wherein said cavity is provided with cooling fins and steam nozzle ducts for conducting steam to said orifices and wherein said steam is dischargeable at the head portion and the trailing edge portion at high velocity on start-up to create a reactive force, whereby the gas generator is rotatable, and whereby the eddy currents and leaving losses are minimized, 51. The improvement of claim 49 wherein said cooling steam is admitted to the interior of said power turbine blade, the turbine being supported by shafting having a small diameter and the turbine being provided with entrance holes for admission of cooling steam, wherein said cooling steam is introduced by steam lines leading to a labyrinth packing seal at the small diameter of the shafting, an annular cavity connecting said labyrinth packing seal to the turbine at the turbine entrance holes, said cooling steam flowing internally in the annular cavity through the entrance holes into the turbine wheel. 52. The improvement of claim 51 wherein said turbine blades are sealed with flexible metallic-type gaskets at the shafting, whereby steam pressure can be maintained and wherein said annular cavity is provided with orifices for cooling the first-stage disk. 53. The improvement of claim 48 wherein said steam is dischargeable at the leading edge nozzle and at the trailing edge nozzle at high velocity on start-up to create a driving force on the blade, whereby the gas generator means is rotatable, and whereby eddy currents and leaving losses are minimized.

Inventors: RICE I
Applicants: RICE I
Application year: 1979
Main classification: F02C6
IPC codes: F01D5, F01K23, F02C3, F02C6, F02C7