Multiloop gas turbine, system, and method of operation thereof

Abstract

The present disclosure relates to a novel gas turbine system having applications, for example, in thermal power generation in an environmentally friendly manner. The multiloop gas turbine system may have multiple functional units each comprising a compressor, a regenerator, a combustion unit, and a turbine. Typically, exhaust flow of a turbine of a preceding loop may be routed to the combustion unit of the next loop, allowing mixing of exhaust flow with hot compressed air of the next loop, and the expanded exhaust from the turbine of the ultimate loop is fed back into the regenerators of each loop to recover exhaust heat.

Claims

1. A power generation method, comprising: providing a plurality of functional units, wherein each of the plurality of functional units comprises a compressor, a regenerator, a combustion unit and a turbine, wherein the number of the plurality of functional units is equal to n, and wherein n is an integer greater than 1; directing, for each of the functional units, compressed gas from the compressor to the regenerator, wherein the compressor utilizes a coolant; directing, for each of the functional units, heated gas from the regenerator to the combustion unit; directing, for each of the functional units, exhaust gas from the combustion unit to the turbine; directing, for each of the functional units other than the nth functional unit, expanded gas from the turbine to the combustion unit of a successive functional unit to provide to the combustion unit of the successive functional unit expanded gas from the turbine; and directing, for the nth functional unit, expanded gas from the turbine to the regenerator of each preceding functional unit, in parallel, to provide to the regenerator of each preceding functional unit expanded gas from the turbine of the nth functional unit, wherein the expanded gas from the turbine of the nth functional unit is split into n portions, each regenerator of the plurality of functional units being provided with one portion of the n portions of the expanded gas from the turbine of the nth functional unit.

2. The power generation method of claim 1, wherein each combustion unit utilizes fuel from a fuel supply to perform combustion.

3. The power generation method of claim 2, further comprising a fuel storage containing the fuel supply, the fuel storage being in operative communication with each of the combustion units.

4. The power generation method of claim 2, wherein the fuel supply is selected from the group: (a) natural gas; (b) methane; (c) kerosene; (d) diesel fuel; (e) gasoline; (f) coal; (g) combustible oil; (h) combustible wood; (i) any combustible material; (j) a liquid hydrocarbon; (k) a gaseous hydrocarbon; (l) hydrogen; (m) jet fuel; and (n) any combination thereof.

5. The power generation method of claim 1, wherein, in each functional unit, the regenerator is configured to receive the coolant from the compressor.

6. The power generation method of claim 5, wherein, in each functional unit, the compressor is configured to receive back recovered coolant from the regenerator.

7. The power generation method of claim 1, wherein, in each functional unit, the coolant is independently selected from the group: (a) water; (b) methanol; (c) ethanol; and (d) any combination thereof.

8. The power generation method of claim 1, wherein the plurality of functional units comprises a first functional unit and a second functional unit, and the second functional unit is the successive functional unit of the first functional unit; wherein the first functional unit comprises a first compressor, a first regenerator, a first combustion unit and a first turbine, and the second functional unit comprises a second compressor, a second regenerator, a second combustion unit and a second turbine; and the first compressor and the second compressors are multistage compressors and the first compressor has more stages than the second compressor.

Description

BRIEF DESCRIPTION OF THE DRAWINGS

(1) These and other features, aspects, and advantages of the present invention will become better understood with regard to the following description, appended claims, and accompanying drawings (some of the drawings may be not drawn to scale and some of the drawings may be drawn at the indicated scale; further, where scale and/or dimensions are provided, they are provided as examples only) wherein:

(2) FIG. 1 illustrates various component icons used in the figures.

(3) FIG. 2 illustrates the Brayton Open Cycle (Prior Art).

(4) FIG. 3 illustrates the Ericsson Closed Cycle (Prior Art).

(5) FIG. 4 illustrates the Regenerative Open Cycle (Prior Art).

(6) FIG. 5 illustrates a Novel Open Cycle scheme according to an embodiment of the disclosure.

(7) FIG. 6 illustrates a Novel Open Cycle scheme according to an embodiment of the disclosure.

(8) FIG. 7 illustrates a Novel Open Cycle scheme according to an embodiment of the disclosure.

(9) FIG. 8 illustrates a Novel Closed Cycle scheme according to an embodiment of the disclosure.

(10) FIG. 9 illustrates a Novel Closed Cycle scheme according to an embodiment of the disclosure.

(11) FIG. 10 illustrates a Novel Closed Cycle scheme according to an embodiment of the disclosure.

(12) FIG. 11 illustrates a graph of Novel Cycle performance for coal according to an embodiment of the disclosure.

(13) FIG. 12 illustrates a graph of Novel Cycle performance for solar thermal energy according to an embodiment of the disclosure.

(14) FIG. 13 (referred to in Appendix B as FIG. B-1) shows the Variation of Specific Power ω (Theta≡θ).

(15) FIG. 14 (referred to in Appendix B as FIG. B-2) illustrates the Variation of Thermal Efficiency η≡Et, Theta≡θ.

(16) FIG. 15 (referred to in Appendix B as FIG. B-3) illustrates the Sensitivity of CO.sub.2 and H.sub.2O Efflux with η.

(17) FIG. 16 (referred to in Appendix C as FIG. C-1(a)) illustrates the Variation of Specific Power for Theta≡θ=3.

(18) FIG. 17 (referred to in Appendix C as FIG. C-1(b)) illustrates the Variation of Specific Power for Theta≡θ=6.

(19) FIG. 18 (referred to in Appendix C as FIG. C-2(a)) illustrates the Variation of Thermal Efficiency for θ=6η≡Et.

(20) FIG. 19 (referred to in Appendix C as FIG. C-2(b)) illustrates the Variation of Thermal Efficiency for θ=3(η≡Et).

(21) FIG. 20 (referred to in Appendix C as FIG. C-3) illustrates the Relative Specific Power ω/ω.sub.ERC with n.

(22) FIG. 21 (referred to in Appendix C as FIG. C-4) illustrates the Relative Thermal Efficiency η/η.sub.ERC with n.

(23) FIG. 22 (referred to in Appendix C as FIG. C-5) illustrates the Variation of Relative Air to Fuel Ratio.

APPENDIX LISTING

(24) Appendix A identifies various nomenclature used herein.

(25) Appendix B provides an analytical basis for various cycles described herein. Appendix B also discusses Sensitivity of CO.sub.2 Efflux.

(26) Appendix C provides a discussion of performance of various Novel Cycles according to embodiments of the disclosure. Appendix C also discusses results of limited simulation and principles of component commonality according to various embodiments of the disclosure.

(27) Appendix D provides a detailed derivation of equations used herein.

(28) Appendix E identifies various references mentioned herein.

(29) Appendix F discusses some additional historical perspective.

DETAILED DESCRIPTION OF THE DISCLOSURE

(30) Detailed embodiments of the present disclosure are described herein; however, it is to be understood that the disclosed embodiments are merely illustrative of the compositions, structures and methods of the disclosure that may be embodied in various forms. In addition, each of the examples given in connection with the various embodiments is intended to be illustrative, and not restrictive. Further, the figures are not necessarily to scale, some features may be exaggerated to show details of particular components. Therefore, specific structural and functional details disclosed herein are not to be interpreted as limiting, but merely as a representative basis for teaching one skilled in the art to variously employ the compositions,

(31) structures and methods disclosed herein. References in the specification to “one embodiment”, “an embodiment”, “an example embodiment”, etc., indicate that the embodiment described may include a particular feature, structure, or characteristic, but every embodiment may not necessarily include the particular feature, structure, or characteristic. Moreover, such phrases are not necessarily referring to the same embodiment.

(32) For the purpose of describing and claiming the present invention, the term “regenerator” is intended to refer to a type of heat exchanger (for example, a parallel-flow heat exchanger, a counter-flow heat exchanger, or a cross-flow heat exchanger). In one specific example, a regenerator may be an air-to-gas heat exchanger (wherein, for example, air leaving a compressor is heated by gas leaving a turbine such that there is minimal mixing of the air and gas).

(33) For the purpose of describing and claiming the present invention, the term “internal combustion unit” is intended to refer to a combustion unit (such as a burner) which is internal relative to a gas flow path (e.g. a gas flow path of a gas turbine engine system comprising: inlet-compressor-combustion unit-turbine-exhaust, all connected in series in an open cycle configuration).

(34) For the purpose of describing and claiming the present invention, the term “external combustion unit” is intended to refer to a combustion unit (such as a burner) which is external relative to a gas flow path (e.g. a gas flow path of a gas turbine engine system comprising: inlet-compressor-combustion unit-turbine-exhaust, all connected in series in an open cycle configuration).

(35) For the purpose of describing and claiming the present invention, the term “burner” is intended to refer to either: (a) an “internal combustion chamber,” such as associated with all

(36) open cycle configurations; or (b) to an “external combustion chamber,” such as used as an “external heater” and such as associated with all closed cycle configurations.

(37) Embodiments described below are sometimes referred to as Novel Closed (NCn) and Novel Open (NOn) Cycle schemes, where n is the number of Compression/Expansion Stages, n

(38) =1, 2, 3, . . . . In one specific example, n is equal to 2. In another specific example, n is equal to 3. In another specific example, n is equal to 4.

(39) The embodiments described herein are applicable to gas turbine engines in general, such as Aero, Stationary, Locomotive, Marine. It is possible to use any gas turbine software modeling tool known in the art.

(40) Referring now to FIGS. 5-10, various Novel Open and Novel Closed cycle schemes are illustrated for n=1, 2 and 3 compression/expansion stages with Isothermal compression, regeneration and reheating. These differ from conventional reheat scheme which works in series and will have restriction from the depleting Air to Fuel ratio for open cycle. Besides, there is a possibility of increasing specific power substantially by adding a compressor in the loop. Cooled exhaust recompression is, in one example, used for the closed cycle, although valid for open cycle.

(41) For simplicity only, the various Novel Cycle concepts are illustrated under ideal conditions of perfect processes and constant properties. Complete derivation of the performance expressions is provided in the Appendix C.

(42) As depicted in FIG. 5, the baseline Novel Open Cycle (n=1) utilizes compressor 501, regenerator 503, combustor 505 and turbine 507. Also depicted in FIG. 5 is a fuel 509. The baseline Novel Open Cycle (n=1) is similar to Open Regenerative Cycle (RGO) of FIG. 4, except with respect to compression, which is ideally isothermal instead of adiabatic. An additional feature added is the use of water for internal cooling of compression to the extent possible (for example, under hot ambient and up to saturation) and water recapture from the regenerator exhaust gases. In the process 1 kg of Methane generates 2.25 kg of water.

(43) Referring now to FIG. 6, the illustrated two stage Novel Open Cycle (n=2) utilizes compressors 601A and 601B, regenerators 603A and 603B, combustors 605A and 605B, and turbines 607A and 607B. Also depicted in FIG. 6 is a fuel 609. Further, with respect to the two stage Novel Open Cycle (n=2) of FIG. 6, the pressure ratio of each expansion stage is π.sup.1/2. The two compression stages have pressure ratios of π and π.sup.1/2 respectively. The exhaust of the first turbine 607A is routed to the burner 605B of the second stage where it mixes with the fresh air from the second compressor 601B after regenerative heating by regenerator 603B. The exhaust from the second turbine 607B is split in two equal parts and routed to the two regenerators 603A, 603B and further to stack. Compressors 601A, 601B are internally cooled by spraying water.

(44) Water is recaptured from both the regenerator exhaust flows. The air to fuel ratio in the second burner 605 needs to be within limits of stable and complete combustion. Further, each compressor may be single stage or multi-stage (in the example shown in this FIG. 6, compressor 601A is at least two stages and compressor 601B is at least a single stage).

(45) Referring now to FIG. 7, the illustrated three stage Novel Open Cycle (n=3) utilizes compressors 701A, 701B and 701C, regenerators 703A, 703B and 703C, combustors 705A, 705B and 705C, and turbines 707A, 707B and 707C. Also depicted in FIG. 7 is a fuel 709. Further, with respect to the three stage Novel Open Cycle (n=3) of FIG. 7, the pressure ratio of each expansion stage is π.sup.1/3. The three compressions stages have pressure ratios of π, π.sup.2/3 and π.sup.1/3 respectively. The exhaust of the first turbine 707A is routed to the burner 705B of the second stage where it mixes with the fresh air from the second compressor 701B after regenerative heating by regenerator 703B. The exhaust of the second turbine 707B is routed to the burner 705C of the third stage where it mixes with the fresh air from the third compressor 701C after regenerative heating by regenerator 703C. The exhaust from the third turbine 707C is split in three equal parts and routed to the three regenerators 703A, 703B, 703D and further to stack. The remaining operation is similar to the above and hence not repeated. The air to fuel ratio in the second and third burners 705B, 705C needs to be within limits of stable and complete combustion. Further, each compressor may be single stage or multi-stage (in the example shown in this FIG. 7, compressor 701A is at least three stages, compressor 701B is at least two stages and compressor 701C is single stage).

(46) As depicted in FIG. 8, the baseline Novel Closed Cycle (n=1) utilizes compressor 801, regenerator 803, combustor 805 and turbine 807. The baseline Novel Closed Cycle (n=1) is similar to the Ericsson Cycle (ERC) of FIG. 3, except for expansion, which is adiabatic instead of ideally Isothermal. The internal combustion burner of FIG. 3 is replaced by either external combustion heater or renewable heat source, such as concentrated solar thermal (CST) or biogas burner or geothermal. In this embodiment, the exhaust from the turbine comprises hot air and is recompressed after passing through the regenerator. Water can be recaptured, if feasible and desirable, by cooling the external combustion burner exhaust, in the case that the fuel contains a large portion of methane. In another embodiment, the closed cycle may use air or any other gas, such as CO.sub.2, for its operation.

(47) Referring now to FIG. 9, the illustrated two stage Novel Closed Cycle (n=2) utilizes compressors 901A and 901B, regenerators 903A and 903B, combustors 905A and 905B, and turbines 907A and 907B. In addition, each compressor may be single stage or multi-stage (in the example shown in this FIG. 9, compressor 901A is at least two stages and compressor 901B is at least a single stage). Further, the two stage Novel Closed Cycle (n=2) of FIG. 9 is similar to the Open cycle scheme of FIG. 6, except the fact that the turbine exhaust after similar splitting and regenerative cooling goes to respective compressors for recompression.

(48) Referring now to FIG. 10, the illustrated three stage Novel Closed Cycle (n=3) utilizes compressors 1001A, 1001B and 1001C, regenerators 1003A, 1003B and 1003C, combustors 1005A, 1005B and 1005C, and turbines 1007A, 1007B and 1007C. In addition, each compressor may be single stage or multi-stage (in the example shown in this FIG. 10, compressor 1001A is at least three stages, compressor 1001B is at least two stages and compressor 1001C is single stage). Further, the three stage Novel Closed Cycle (n=3) of FIG. 10 is similar to the Open cycle scheme of FIG. 7, except the fact that the turbine exhaust after similar splitting and regenerative cooling goes to respective compressors for recompression.

(49) It is apparent that the thermodynamics of both the “Closed” and “Open” schemes of the Novel Cycles is precisely same. As before, the cycles operate with the same compression pressure ratio π and the same cycle temperature ratio θ.

(50) For fair comparison, performance of the Novel Cycles must be compared with the sum of n Ericsson cycles of similar complexity, cycle pressure ratio π and cycle temperature ratio θ. As n increases the Novel Cycle is expected to approach the Ericsson cycle but with fast diminishing improvement in both specific power and thermal efficiency. Therefore, in one example, n=2 or 3 is considered reasonable and sufficient for practical implementation. Appendix C provides a discussion of performance of various Novel Cycles according to embodiments of the disclosure.

(51) In various embodiments, mechanisms are provided to address the changing global requirements, such as least specific CO.sub.2 efflux, ability to work efficiently with renewable thermal energy (e.g., solar), adaptive designs for distributed generation/consumption patterns, minimization or elimination of use of water and acceptability through affordability to the global emerging markets.

(52) In various embodiments, a compressor can be cooled externally by a coolant such as water or air. In one such embodiment, the cooling may be through a heat exchange across compressor external wall (the so-called inter-stage coolers or inter-coolers). In other embodiments, a compressor can be cooled internally by evaporation of a liquid, such as water or water-methanol mixtures. In other embodiments, the cooling may be external cooling and/or internal evaporative cooling. In another embodiment, the coolant may be one which consumes large energy for evaporation or Latent heat.

(53) In other embodiments, renewable sources of energy may include: (a) Solar Thermal (e.g., concentrated solar thermal); (b) Geo-Thermal; (c) Biomass, in its numerous forms from animals and plants; and (d) Ocean temperature differences.

(54) In another embodiment, concentrated solar thermal may be integrated with Internal/External Heating, the former obviating the need for a heat exchanger.

(55) As described herein are mechanisms for, in essence, bridging the gap between the Brayton cycle and the Ericsson cycle by utilizing: (a) quasi-isothermal internally cooled compression; (b) staged expansion with parallel reheat loops; (c) successive expansion mass flow compounding at burners/heaters; and (d) exhaust splitting, recirculation, regeneration (recompression-closed cycle). In one embodiment, provided is networking of 3 regenerative, quasi-isothermal compression open cycles in series-parallel loops. In another embodiment, provided is networking of 3 regenerative, quasi-isothermal compression closed cycles in series-parallel loops.

(56) Reference will now be made to some example applications of the Novel Cycles according to various embodiments of the disclosure.

(57) The first of these applications is the use of coal as a gas turbine fuel. In this regard, the Novel Closed/Open Cycle Schemes may be applied to various futuristic requirements, such as converting an existing gas fired plant in an environmentally acceptable and commercially competitive manner by operating on cheaper and locally available fuel, such as coal.

(58) Referring to FIG. B-3 (of Appendix B), it is seen that for such conversion we have to increase the thermal efficiency 1.667 times, essentially by using the same turbo-machinery hardware of the existing plant. Compressor blades of the existing Gas-Fired plants can be utilized to construct a two or three stage compressor layout. Multiple units of the original turbine can be used. Relative values of specific power and pressure ratio for various values of cycle temperature ratio θ are presented in FIG. 11, when the relative thermal efficiency is 1.667 times the original baseline gas turbine. This is because burning of coal produces 1.667 times more CO.sub.2 per unit heat of combustion as compared to methane, the main constituent of natural gas.

(59) It may be noted that the relative specific power is doubled or tripled for n=2 and 3, which would justify the added complexity.

(60) Another application is in the context of gas turbines for solar power. In this regard, the Novel Open and Closed Cycle Schemes may be utilized, for example, in concentrated solar thermal power applications. Currently, many solar or solar-fuel hybrid designs are being explored around existing gas turbine hardware or otherwise [2, 4, 5]. For open cycle schemes [4], the typical cycle temperature ratio is about 3 to 4.5 and pressure ratios are moderate at about 9 to

(61) 15. For supercritical CO.sub.2 based closed cycles [5], the typical cycle temperature ratio is about 3 and pressure ratios are low at about 3 to 6. In one example, the Novel Open and Closed Cycle schemes may be applied to such applications. The simulated performance of these Novel Cycles for the modest cycle temperature ratio of θ=3 and cycle pressure ratios from π=3 to 15 is presented in FIG. 12. Here the specific power and thermal efficiency of the Novel Cycles as referenced to the baseline Brayton cycle for same π and θ is denoted by ‘RW’ and ‘REta’ respectively. The data is presented for n=2 and 3.

(62) As seen from FIG. 12, the Novel Open and Closed Cycle Schemes give high specific power for all pressure ratios, albeit at increased complexity. The thermal efficiency of these cycles is far superior (150-240%) at low pressure ratios (3 to 6) and is about 11-15% higher at high pressure ratio of 15 for n=2 and 3. Thus, the Novel Cycles are well-suited to the supercritical CO.sub.2 based closed cycles. In this example, compressors can be cooled only externally.

(63) As mentioned above, for simplicity only, the various Novel Cycle concepts are illustrated under ideal conditions of perfect processes and constant properties. In this regard, it is believed that since the ideal cycle analysis with constant properties and mass flow rates correctly captures the performance trends for specific power and thermal efficiency, these trends will be consistent under real cycle conditions. Based upon the analysis and limited simulation results presented herein the following four major conclusions can be drawn:

(64) (1) The Novel Open and Closed Cycle Schemes, as analysed, satisfy the basic object of simultaneously achieving high specific power and thermal efficiency without the bottoming Rankine cycle.

(65) (2) The Novel Cycle concepts provide the necessary flexibility for choosing the design values of cycle pressure ratio π and cycle temperature ratio θ, while optimizing the cycle for specific requirements, such as preferred energy source (e.g., coal, natural gas or concentrated solar thermal), water conservation (such as for arid areas) and/or water generation.

(66) (3) The Novel Cycles are suited, for example, for small/medium plants with low pressure/temperature ratios as preferred options under a distributed generation/consumption model.

(67) (4) Principles of Component Commonality to reduce overall acquisition costs and improve affordability have been established and may be considered for future plant configurations.

(68) In one embodiment, the multiloop gas turbine has each loop comprising at least one compressor to draw air or other gas for combustion purpose, a coolant to cool the compressed air or other gas, a regenerator to admit compressed air or other gas as well as hot expanded air or other gas from a turbine of the ultimate loop in separate passages with heat conducting walls, a combustion unit for combustion of compressed air or other gas and a fuel, and at least one turbine, wherein exhaust flow of the turbine of the preceding loop is routed to the combustion unit of the next loop, allowing mixing of exhaust flow of the first loop with the hot compressed air or other gas of the next loop, wherein the hot expanded exhaust from the turbine of the ultimate loop is fed back into the regenerators of each loop to recover exhaust heat thereby enhancing efficiency, and wherein the coolant is recaptured from the regenerator of each loop.

(69) In various embodiments, the multiloop gas turbine is an open cycle or closed cycle turbine. In an embodiment of the open cycle multiloop gas turbine, the combustion unit is, but not limited to, an internal combustion burner. In an embodiment of the closed cycle multiloop gas turbine, the combustion unit is, but not limited to, an external combustion heater or a renewable heat source, such as solar or geothermal. The compressors may be internally cooled by the coolants, such as by spraying water. The water may be recaptured from regenerator exhaust flow.

(70) In one example, the multiloop gas turbine is a three-loop gas turbine and includes three-stage compressors at the first loop, two-stage compressors at the second loop, and a single compressor at the third loop.

(71) In another example, the multiloop gas turbine is a two-loop gas turbine and includes two-stage compressors at the first loop and a single compressor at the second loop.

(72) Also disclosed herein is a method of operating a multiloop gas turbine by: (i) producing an exhaust from a turbine of a preceding loop, (ii) routing the exhaust of a preceding loop to a heating unit of a next loop, allowing mixing of exhaust flow of the first loop with the hot compressed air or other gas of the next loop, (iii) feeding back the hot expanded exhaust from the turbine of the ultimate loop back into the regenerators of each loop to recover thermal energy and thereby enhancing efficiency, and (iv) recapturing water from each loop for recycling the same for cooling after purification, if required. In an open cycle operation with natural gas as the fuel there is typically enough water generated, which can be captured and used, especially in arid areas.

(73) In one embodiment, all of the compressors operate with same mass flow and inlet conditions. In this embodiment, the compressor in the first loop (of a three loop configuration) may be a three-stage compressor, each compressor stage having pressure ratio of π.sup.1/3. The compressor in the second loop of this embodiment may be a two-stage compressor, each compressor stage having pressure ratio of π.sup.1/3. In the third loop of this embodiment, there is only one compressor stage. The turbines of this embodiment operate with same pressure ratio and same entry temperature. All regenerators of this embodiment have identical hot-side flow conditions. On the cold-side, the mass flow rates and temperatures are identical; however, the pressures are different as πP.sub.o, π.sup.2/3P.sub.o and π.sup.1/3P.sub.o respectively. All the heating units of this embodiment operate between the same temperature limits but differ in mass flow capacities and pressure levels. The heating unit in the first loop of this embodiment operates with mass flow w at pressure πP.sub.o. The number heating units of this embodiment is same as the number of turbines.

(74) In an embodiment, the turbine blades (e.g., rotor blades, stator blades) described herein of any one or more compression/expansion stages may be substantially the same (e.g., in terms of dimensions such as size, shape and weight and in terms of materials).

(75) In an embodiment, the turbine blades may be the same or different.

(76) In an embodiment, the compressor blades (e.g., rotor blades, stator blades) described herein of any one or more compression/expansion stages may be substantially the same (e.g., in terms of dimensions such as size, shape and weight and in terms of materials).

(77) In an embodiment, the compressor blades may be the same or different.

(78) In an embodiment, the regenerators described herein of any one or more compression/expansion stages may be substantially the same (e.g., in terms of dimensions such as size, shape and weight and in terms of materials).

(79) In an embodiment, the regenerators may be the same or different.

(80) In an embodiment, the combustion units (e.g., burners) described herein of any one or more compression/expansion stages may be substantially the same (e.g., in terms of dimensions such as size, shape and weight and in terms of materials).

(81) In an embodiment, the combustion units may be the same or different.

(82) The Novel Closed/Open Cycles substantially enhance the performance in terms of both efficiency and specific power.

(83) The present invention is cost-effective, especially if the power generation system utilizes component commonality. The major cost component of gas turbine hardware is the number of different types of rotor and stator blades of axial flow compressors and turbines, especially for the internally cooled turbine. Using the same or substantially similar components such as similar compressor blades, similar turbine blades, similar regenerators, similar combustion units and the like will reduce the overall cost as the onetime “NRE” cost of masters, dies, jigs and fixtures, machine-tools and balancing rigs is shared. This feature is illustrated herein for the major components.

APPENDIX A

(84) Nomenclature

(85) A—Flow area in m.sup.2 B—Burner BRO—Brayton Open Cycle C—Compressor CC—Combustion Chamber C.sub.p—Specific Heat at constant pressure in kJ/kg/K C.sub.v—Specific Heat at constant volume in kJ/kg/K ERC—Ericsson Cycle Et, Eta—Thermal efficiency same as η GT—Gas Turbine Ln—natural logarithm m—running stage number MJ—Mega Joules n—number of expansion/compression stages NCn—Novel Closed cycle of n stages NG—natural gas NOn—Novel Open cycle of n stages NOCn—Novel Open/Closed cycle of n stages NRE—non recurring equipment P.sub.o—ambient pressure in kPa P.sub.max—maximum cycle pressure in kPa P.sub.min—minimum cycle pressure in kPa PR—cycle or compression pressure ratio same as π Q.sub.B—rate of heat supplied in kW Q.sub.BT—total rate of heat supplied in kW Q.sub.T—rate of heat supplied to turbine in kW R—gas constant in kJ/kg/K REta—relative thermal efficiency referred to simple Brayton cycle RGi—regenerator number i RW—relative specific power referred to simple Brayton cycle T—turbine T.sub.max—maximum cycle temperature in K T.sub.min—minimum cycle temperature in K To—ambient temperature in K w—air mass flow rate in kg/s W.sub.C—compression power in kW W.sub.CE—compression power in Ericsson cycle in kW W.sub.Cm—compression power of mth stage in kW W.sub.CT—total compression power in kW W.sub.Cy—cycle power in kW W.sub.T—turbine power in kW W.sub.TE—turbine power in Ericsson cycle in kW W.sub.TT—total turbine power in kW γ—ratio of specific heats η—thermal efficiency π—cycle or compression pressure ratio τ—adiabatic compression temperature ratio θ—cycle temperature ratio ω—cycle specific power non-dimensional form or referred to WC.sub.pT.sub.o

APPENDIX B

(86) It can be analytically established that relative specific cycle power ω (in non-dimensional form or referred to wC.sub.pT.sub.o) and the cycle thermal efficiency η for the BRO, ERC and RGO cycles are:
ω.sub.BRO=(θ−τ)(1−1/τ); η.sub.BRO=1−1/τ  (1a)
ω.sub.ERC=[(θ−1)Ln(τ)]; η.sub.ERC1−1/θ  (1b)
ω.sub.RGO=(θ−τ)(1−1/τ); η.sub.RGO=1−τ/θ  (1c)
Assuming ambient temperature T.sub.o=300 K, for typical high/low values of θ=6 (T.sub.max=1800K) and θ=3 (T.sub.max=900K), the behavior of ω and η of the above three cycles is presented on FIG. B-1 and FIG. B-2, for typical variation of cycle pressure ratio π.
Sensitivity of CO.sub.2 Efflux
Consider two typical fossil fuels Coal and Natural Gas. As per [3] the specific CO.sub.2 in (g/MJ) is about 100 for Coal and 60 for Natural Gas, which works out in (kg/kWh) to about (0.36/η) for coal and (0.216/η) for Natural Gas. The specific water generated with Natural gas as fuel will be about (0.162/η) in (kg/kWh) which is not a small quantity to be ignored, where material. FIG. B-3 presents the variation in specific CO.sub.2 and H.sub.2O with thermal efficiency for these fuels. In order that Coal based plants compare with Natural Gas, these must operate with about 67% higher thermal efficiency (30%.fwdarw.50%) to be environmentally competitive.

APPENDIX C

(87) Generic Novel Cycle Performance (NOCn)

(88) It can be analytically established that relative specific cycle power ω (in non-dimensional form or referred to WC.sub.pT.sub.o) and the cycle thermal efficiency η for both the novel closed and novel open generic cycles (NOCn) and n equivalent Ericsson cycles (ERCn) are:
ω.sub.NOCn=½.Math.n.Math.(n+1).Math.[θ(1−1/τ.sup.1/n)−Ln(τ.sup.1/n)]  (2a)
ω.sub.ERCn=½.Math.(n+1).Math.(θ−1).Math.Ln(τ)  (2b)
η.sub.NOCn={1−Ln(τ.sup.1/n)/[θ(1−1/τ.sup.1/n)]}  (2c)
η.sub.ERCn=1−1/θ  (2d) n=1, 2, 3, . . . .
For the same typical high/low values of θ=6 and θ=3, considered earlier, the behavior of ω and η is presented on FIGS. C-1(a), C-1(b), C-2(a), and C-2(b) for the same typical variation of cycle pressure ratio π.

(89) The changes in specific power and thermal efficiency, as referenced to sum of n equivalent Ericsson cycles for same π and θ, with n compression/expansion stages and typical pressure ratios π=10, 20, 30, 40 and θ=3 and 6, are presented on FIG. C-3 and FIG. C-4.

(90) Relative Air to Fuel Ratios

(91) Methane will need more oxygen to burn compared to coal. Although, higher hydrocarbon components will need still more oxygen than methane, their mass fractions are very small. For simplicity the fuel is considered as methane only. The variation of Relative Air to Fuel Ratio for two and three stage expansion in the Combustion Chambers (CCnm) for the Novel Generic Open Cycle schemes are presented for the high value of θ=6 in FIG. C-5. It is clear that the Air to Fuel mixture will get richer in the successive burners due to mixing of the turbine exhaust of the previous stage. The Relative Air to Fuel Ratio for the last burner must be in the feasible range of stable and complete combustion. The issue will be less material for smaller values of θ and immaterial for closed cycles with external heating.

(92) Discussion of Results of Limited Simulation

(93) Refer to FIGS. C-1(a), C-1(b), C-2(a), and C-2(b). The basic objective, set forth for the Novel Gas Turbine Cycles, of achieving simultaneously high specific power and closer to Carnot thermal efficiency without a bottoming Rankine cycle, has been realised in substantial measure. Specific power monotonically increases with pressure ratio and thermal efficiency decreases but the rate of increase in specific power is much higher in contrast to fall in thermal efficiency. As the number of compression/expansion stages increase from 1 to 2 to 3, the difference between the specific power of the generic cycle and the equivalent sum of Ericsson cycles reduces. The thermal efficiency trends are also similar.

(94) Refer to FIG. C-3. Relative specific power, of the novel cycles referenced to equivalent sum of Ericsson cycles, approaches to 87% at high value of θ=6 at lowest considered pressure ratio of π=10. This reduces to about 82% for the highest pressure ratio considered of π=40. For the low value of θ=3 these numbers are 85% and 76% respectively.

(95) Refer to FIG. C-4. Thermal efficiency increases with the number of compression/expansion stages n and approaches Carnot limit. For n=3, the % difference being about 2-3% for θ=6 and 6-9% for θ=3. The point to be noted is that the thermal efficiency of the novel cycles is closest to Ericsson cycle efficiency at lowest considered pressure ratio of π=10 with highest possible specific power for all cycle temperature ratios θ. The choice of lowest pressure ratio will be governed by acceptable specific power.

(96) Refer to FIG. C-5. The Relative Air to Fuel Ratio even for high value of θ=6 may become restrictive only for very high values of compressor pressure ratio and that too for the last burner. Thus, this is not a restriction, when it comes to optimizing the novel generic cycle configurations.

(97) Principles of Component Commonality

(98) As established, the Novel Closed/Open Cycles substantially enhance the performance in terms of both efficiency and specific power, albeit with increased layout complexity and number of rotating and stationary components. This is also the case with the Combined Cycle schemes and the sum of equivalent Ericsson cycles scheme considered above for comparison. The major cost component of Gas Turbine hardware is the number of different Types of rotor and stator blades of axial flow compressors and turbines, especially for the cooled turbine. Using more number of the same components will reduce the overall cost as the onetime NRE cost of ‘masters’, ‘dies’, jigs and fixtures, machine-tools and balancing rigs is shared. This feature is illustrated for the major components.

(99) Rotating Components

(100) Compressors: For example, refer to FIG. 7 or FIG. 10 with n=3. All compressors operate with same mass flow and inlet conditions. Compressor of the first loop is a three stage compressor, say, C1, C2 and C3 in series, each with a pressure ratio of π.sup.1/3. The second loop compressor comprises C1 and C2 in series, each with a pressure ratio of π.sup.1/3. The third loop compressor is just C1 with a pressure ratio of π.sup.1/3. Thus we need the more of same set of rotor and stator blades, as in the baseline cycle with the total compressor hardware being (3C1+2C2+C3).

(101) Turbines: The commonality in turbines is not evident from the scheme shown on FIG. 7 or FIG. 10. It may be noted that all turbines operate with the same pressure ratio π.sub.T=π.sup.1/3 and same entry temperature of θT.sub.o. It may be noted that each successive turbine stage has to handle the cumulative mass flow and the pressure level will reduce by π.sub.T. Following the principles of thermodynamics of isentropic compressible flow, it can be shown that the turbine flow area required for the turbines T1, T2 and T3 will be:
A.sub.T1∝{[W(RθT.sub.o).sup.1/2]/[πP.sub.of(π.sub.T,γ)]}∝W/π  (3a)
A.sub.T2∝2(W/π)π.sub.T  (3b)
A.sub.T3∝3(W/π)π.sub.T.sup.2  (3c)
It is, thus, possible to use multiple units of T1 for the turbines T2 and T3. For example, for a case with π=27 and π.sub.T=3 we will need six T1 units for T2 and twenty seven T1 units for T3. For low pressure ratio of π=8 and π.sub.T=2 we will need four T1 units for T2 and twelve T1 units for T3. Thus, we need the same set of rotor and stator blades, as in the baseline cycle but substantially increase the quantity. The total turbine hardware will be thus (34T1 or 14T1) respectively for π=27 and 8. Although these numbers seem to be too large and impracticable, the overall plant complexity and costs need to be carefully worked out before pronouncing final decision. Obviously, mechanical shaft connectivity, between compressors, turbines and load, certainly needs some innovative solutions. Such concepts are considered, for example, in Blended-Wing-Body configuration with distributed propulsion.
Stationary Components
Regenerators: Refer to FIG. 7. All regenerators have identical hot-side flow conditions. On cold-side the mass flow rates and temperatures are also same but the pressures are different as πP.sub.o, π.sup.2/3P.sub.o and π.sup.1/3P.sub.o respectively. Following similar analysis for isentropic compressible flow (as in case of turbines) we can deduce:
A.sub.RG1∝W/π  (4a)
A.sub.RG2∝W/π.sup.2/3  (4b)
A.sub.RG3∝W/π.sup.1/3  (4c)
First regenerator RG1 will be of baseline size and the other two regenerators RG2 and RG3 can be made of multiple units of RG1. We need innovative packaging solutions for practicality.
Burners/Heaters: Refer to FIG. 7. All Burners/Heaters operate between same temperature limits but differ in mass flow capacities and pressure levels. The first burner B1 operates with mass flow W at pressure πP.sub.o. It may be noted that each successive burner/heater stage has to handle the cumulative mass flow and the pressure level will reduce by π.sup.1/3. Following the principles of thermodynamics of isentropic compressible flow it can be shown that the burner/heater flow area required for the units B1, B2 and B3 will be:
A.sub.B1∝W/π  (5a)
A.sub.B2∝2W/π.sup.2/3  (5b)
A.sub.B3∝3W/π.sup.1/3  (5c)
Thus, the number of burner/heater units required will be identical to the respective turbine units employed, as in Eq. 3(a) to 3(c).

APPENDIX D

(102) Derivation of Equations

(103) The detailed derivation of the equations, given herein, is appended here for the sake of completeness, especially for the novel cycle schemes. Dry ambient conditions, common to all cycles, are pressure P.sub.o in (kPa) and temperature T.sub.o in (K). Air Mass Flow is w.sub.in (kg/s) and properties gas constant R in (kJ/kg/K), specific heat at constant pressure Cp in (kJ/kg/K) and ratio of specific heats γ=C.sub.p/Cv are assumed constant. The two more variables, cycle temperature ratio θ and cycle pressure ratio π, define all the cycles. θ is the cycle temperature ratio (T.sub.max/T.sub.min) and π is the cycle pressure ratio (P.sub.max/P.sub.min). The cycle pressure ratio π is also the overall compression pressure ratio and hence the isentropic compression temperature ratio τ is equal to π.sup.γ−1/γ, where γ is ratio of specific heats C.sub.p/C.sub.v. As seen below, for analytical derivations, ‘τ’ is more convenient variable than ‘π’, even for ‘Isothermal’ processes, for which when τ.sub.T=const.=1, as [R.Math.Ln(π)=C.sub.p.Math.Ln(τ)]. The novel cycles need one more parameter n, which is the number of compression/expansion stages (n=1, 2, 3, etc). The expansion stage pressure ratio is thus π.sup.1/n and the compression stage pressure ratios will be π.sup.m/n, m reducing from n to 1. ‘SI’ system of units is used so that (w C.sub.pT.sub.o) is in kW.

(104) Brayton Open Cycle (BRO)

(105) Refer to FIG. 2. Adiabatic Compressor Power in (kW) is given by
W.sub.C=w C.sub.pT.sub.o(τ−1)  (a1)
Adiabatic TurbinePower in (kW) is given by
W.sub.T=w C.sub.pT.sub.oθ(1−1/τ)  (a2)
Cycle Power is given by
W.sub.Cy=W.sub.T−W.sub.C=wC.sub.pT.sub.o[θ(1−1/τ)−(τ−1)]  (a3)
Specific Power ω is defined as W.sub.Cy/(C.sub.pT.sub.o) or
ω.sub.BRO=(θ−τ)(1−1/τ)  (1a)
Heat Input rate in (kW) will be given by
Q.sub.B=wC.sub.pT.sub.o(θ−τ)  (a4)
and thus Cycle Thermal Efficiency will be given by
η.sub.BRO=W.sub.Cy/Q.sub.B=(1−1/τ)  (1a)
Ericsson Cycle (ERC)
Refer to FIG. 3. Isothermal Compressor Power in (kW) is given in terms of adiabatic compression temperature ratio τ=π.sup.γ−1/γ by
W.sub.C=w R T.sub.o Ln(π)=w C.sub.pT.sub.o Ln(τ)  (a5)
Rate of Heat rejected during compression Q.sub.C=W.sub.C
Similarly Isothermal TurbinePower in (kW) is given by
W.sub.T=w C.sub.pT.sub.oθ Ln(τ)  (a6)
Rate of Heat supply during expansion will be Q.sub.T=W.sub.T
Cycle Power is
W.sub.Cy=W.sub.T−W.sub.C=wC.sub.pT.sub.o Ln(τ)(θ−1)  (a7)
Specific Power ω=W.sub.Cy/(w C.sub.pT.sub.o) or
ω.sub.ERC=Ln(τ)(θ−1)  (1b)
Cycle Thermal Efficiency will be η=W.sub.Cy/Q.sub.T or
η.sub.ERC=(1−1/θ)  (1b)
Regenerative Open Cycle RGO)
Refer to FIG. 4. Adiabatic Compressor Power in (kW) is given by
W.sub.C=w C.sub.pT.sub.o(τ−1)  (a8)
Adiabatic TurbinePower in (kW) is given by
W.sub.T=w C.sub.pT.sub.oθ(1−1/τ)  (a9)
Cycle Power is
W.sub.Cy=W.sub.T−W.sub.C=wC.sub.pT.sub.o[θ(1−1/τ)−(τ−1)].  (a10)
Specific Power ω=W.sub.Cy/(w C.sub.pT.sub.o) or
ω.sub.RGO=(θ−τ)(1−1/τ)  (1c)
Noting that the air is heated in the regenerator up to turbine exit temperature, the Heat Input rate in (kW) will be
Q.sub.B=wC.sub.pT.sub.o(θ−θ/τ)  ((a11)
Thus the Cycle Thermal Efficiency will be η=W.sub.Cy/Q.sub.B
η.sub.RGO=(θ−τ)θ=(1−τ/θ)  (1c)
Generic Novel Open/Closed n stage Cycles According to Various Embodiments
Refer to FIGS. 5-10. All compression stages handle same air/gas mass flow and operate isothermally. Successive compression stages operate with reduced pressure ratios of π.sup.m/n with m reducing from n to 1. Thus the stage compression power can be expressed as
W.sub.Cm=w R T.sub.o Ln(τ.sup.m/n)  (a12)
And the total compression power can be expressed as
W.sub.CT=w R T.sub.o Ln(π)Σ(m)/n  (a13)
Further noting that Σ(m)=½ n(n+1), τ=π.sup.γ−1/γ and τ=(τ.sup.1/n).sup.n,
We can write the expression for total compression power as
W.sub.CT=½n(n+1)wC.sub.pT.sub.o Ln(τ.sup.1/n)  (a14)
All turbines operate with same entry temperature T.sub.oθ, same pressure ratio of (π.sup.1/n) or temperature ratio of (τ.sup.1/n). Mass flow of each successive stage is cumulative. Thus the total turbine power can be expressed as

(106) $\begin{array}{cc}\begin{array}{c}{W}_{\mathrm{TT}}={\mathrm{wC}}_p{T}_o\theta \left(1-1/{\tau }^{1/n}\right).Math.\left(m\right)\\ =1/2n\left(n+1\right){\mathrm{wC}}_p{T}_o\theta \left(1+1/{\tau }^{1/n}\right)\end{array}& \left(\mathrm{a15}\right)\end{array}$
Cycle Power will be

(107) $\begin{array}{cc}\begin{array}{c}{W}_{\mathrm{Cy}}=W_{\mathrm{TT}}-W_{\mathrm{CT}}\\ =1/2n\left(n+1\right){\mathrm{wC}}_p{T}_o\left[\theta \left(1-1/{\tau }^{1/n}\right)-\mathrm{Ln}\left({\tau }^{1/n}\right)\right]\end{array}& \left(\mathrm{a16}\right)\end{array}$
Specific Power ω=W.sub.Cy/(w C.sub.pT.sub.o) or
ω.sub.NOCn=½n(n+1)[θ(1−1/τ.sup.1/n)−Ln(τ.sup.1/n)]  (2a)
Burners/Heaters operate between same temperature differences of (T.sub.oθ−T.sub.oθ/τ.sup.1/n) but their air and fuel mass flow rates are cumulative. Total rate of heat added for both types will be similarly equal to
Q.sub.BT=½n(n+1)wC.sub.pT.sub.oθ(1−1/τ.sup.1/n)  (a17)
The cycle thermal efficiency will be η=(W.sub.Cy/Q.sub.BT) or
η.sub.NOCn={1−Ln(τ.sup.1/n)/[θ(1−1/τ.sup.1/n)]}  (2c)
Sum of n Equivalent Ericsson Cycles
n stage Novel Generic Cycle is considered equivalent to n Ericsson cycles together. Each operating with same cycle temperature ratio θ but pressure ratio of π.sup.m/n, m reducing from n to 1.
Since Σ(m/n)=½(n+1) the total compression, expansion and cycle power of n cycles will be:
W.sub.CE=½(n+1)wC.sub.pT.sub.o Ln(τ)  (a18)
W.sub.TE=½(n+1)wC.sub.pT.sub.o Ln(τ)  (a19)
W.sub.CY=½(n+1)wC.sub.pT.sub.o(θ−1)Ln(τ)  (a20)
ω.sub.ERCn=½.Math.(n+1).Math.(θ−1).Math.Ln(τ)  (2b)
η.sub.ERCn−1−1/θ  (2d) n=1, 2, 3, . . . .

(108) The described embodiments of the present invention are intended to be illustrative rather than restrictive, and are not intended to represent every embodiment of the present invention. Various modifications and variations can be made without departing from the spirit or scope of the invention as set forth in the following claims both literally and in equivalents recognized in law.