GAS-TURBINE POWER-PLANT WITH PNEUMATIC MOTOR WITH ISOBARIC INTERNAL COMBUSTION

Abstract

The disclosed concept presents a combination of a gas-turbine power-plant and a pneumatic motor, acting as an isobaric motor-combustor for the gas-turbine power-plant, to the end of achievement of a highly efficient generation of energy/power. In the process of isobaric combustion of fuel within the pneumatic motor, the pneumatic motor, which is supplied with compressed air by an air compressor from the gas-turbine power-plant, simultaneously performs mechanical work of isobaric combustion (in addition to the mechanical work of adiabatic expansion of the gas turbine) and thus increases the overall cycle output and the cycle thermal efficiency. Various combinations between gas-turbine power-plant and pneumatic motor are disclosed: simple, simple-recuperated, intercooled and intercooled-recuperated gas-turbine-cycle configurations, as well as simple and intercooled combined gas-turbine-steam-turbine cycle configurations.

Claims

1. A hybrid energy system with a gas-turbine power-plant in a simple-open cycle and the reciprocating pneumatic motor with internal combustion at constant pressure, consisting of the following interconnected equipment: a) an air compressor (1), sucking atmospheric air and compressing it adiabatically from the atmospheric pressure to a higher pressure, equipped with an air pressure regulator with a screw for pressure setting, connected to a gas turbine (24) and to an additional load (22) by means of a common shaft; b) a pneumatic motor (10) (compressed-air engine) performing the process of an isobaric combustion and the process of a non-adiabatic isobaric expansion of the compressed combusted gas, transferring the compressed-gas power into the reciprocating linear (axial or radial) motion of pistons in two horizontal or vertical single-acting cylinders (11 and 12) and then into rotational motion, thus providing the driving force of the pneumatic motor; whereby, instead of the two said single-acting pneumatic-motor cylinders with reciprocating linear motion, one double-acting cylinder may be used; whereby the said pneumatic motor (10) with reciprocating linear motion contains the following interconnected components: at least two said single-acting cylinders (11 and 12) (or at least one double-acting cylinder) of a circular, quadratic, rectangular or triangular cross-sectional shape; associated pistons (17) with reciprocating linear motion through the said cylinders, each piston being equipped with at least two (2) or preferably three (3) piston rings, like in internal-combustion engines, sealing the said pneumatic motor (10), so that gases cannot escape it, whereby one or two upper/inner piston rings serve primarily for compression sealing (compression rings), whereas the lower/outer ring (oil control ring) serves for controlling the supply of lubrication oil to the said pistons and the said compression rings; associated openings/valves for inlet (14) of the compressed-heated-gas to the said cylinders and outlet (15) of the exhaust gas from the said cylinders, respectively; connecting rods (18) providing physical connection between the said pistons (17) and a crankshaft (19), which facilitates conversion/transformation of the reciprocating linear motion of the pistons into a rotational motion; a camshaft (29) equipped with cams/eccentricities and accurately adjusted with motion of the said crankshaft, facilitating adequate and timely alternate opening/closing of the said opening/closing valves (14 and 15), respectively, by means of its rotation; and a timing belt/timing chain (28) providing an indirect connection and an accurate transmission of motion from the said crankshaft (19) to the said camshaft (29); c) openings (13) for injection of fuel along with electric fuel igniters, comprised within the said pneumatic motor (10), for a complete isobaric combustion of a gaseous (typically natural gas) or a liquid fuel in the stream of compressed air, whereby the desirable locations/ways of injecting the gaseous or liquid fuel and igniting it by an electrical spark (like in a spark-ignition engine) into an operating cylinder of the said pneumatic motor are three-fold: (i) at the top dead center of the operating cylinder, (like in a classic spark-ignition internal-combustion engine); (ii) at the upper side of the operating piston using a flexible fuel pipe (26) inserted thru the piston, so that the fuel ignites during the piston movement towards the bottom dead center and stops igniting when the bottom dead center has been reached; and (iii) combined injection and ignition of fuel by simultaneous use of the said methods (i) and (ii); d) a closed-loop lubrication system of the said pneumatic motor (10), similar as in internal-combustion engines, whereby lubricating oil is sucked out of an oil sump/tank by an oil pump and then forced through an oil filter to the main bearings of the said pneumatic motor (10) and then it passes through feed-holes into drilled passages in the said crankshaft (19) and onto the big-end bearings of the said connecting rods (18), whereas the cylinders walls and piston-pin bearings of the said connecting rods (18) are lubricated by oil drops dispersed by the said rotating crankshaft (19), the excess of the lubricating oil being scraped off by the said lower rings in the said pistons (17), whereas a small fraction of the oil is bled from the main supply passage feeding each bearing of the said camshaft (29), said valves (14 and 15) and valves' springs, while another oil bleed supplies the said timing belt/chain (28) and gears on the said camshaft drive, the excess of lubricating oil being drained back to the said oil sump, where eventually collected heat is being dispersed to the surrounding air; e) preferably, a flywheel (20) of the said pneumatic motor (10) for maintaining the rotational speed of the said crankshaft (19) using its inertial forces (moment of inertia), thus equalizing a potentially fluctuating torque of the pneumatic motor (10) during startup/operation/transients; f) a gearbox (21) for transmission of relatively slow rotational speed of the said crankshaft (19) into a rotational speed needed for a rotor of an electric generator; g) a load (23), typically an electric generator, for electricity generation, connected to the said pneumatic-motor crankshaft (19) by means of the said gearbox (21); h) a well-insulated combusted-gas collecting tank (27) provided at the outlet of the said pneumatic motor (10), for combusted-gas storage and equalizing of a potentially fluctuating combusted-gas flow rate from the said pneumatic motor (10); i) a motorized butterfly valve (25) for adjusting/regulating of the exhaust-gas pressure at the outlet of the said pneumatic motor (10), that is, at the inlet of the said gas turbine (24), at a level close to or slightly lower than the maximum compressed-air pressure; j) the said combustion gas turbine (24) for a full adiabatic expansion process of the combusted gas flowing from the said single-acting cylinders (11 and 12) of the said pneumatic motor (10), driving both the said air compressor (1) and the said additional load (22) by means of a common shaft; k) a cooling-air line for necessary cooling of profiles (stator, rotor) of the said gas turbine (24), by means of branching-off a small fraction of the compressed air from the outlet of the said air compressor (1) and before the inlet to the said pneumatic motor (10); whereby the cooling gas-turbine air may optionally be precooled by means of ambient air or water in an additionally employed cooling-air precooler (9); and l) the said additional load (22), typically an electrical generator, for additional electricity generation, connected to the said common shaft of the said gas turbine (24) and the said air compressor (1).

2. A hybrid energy system according to claim 1, wherein a compressor intercooler (2) is additionally used to perform an intercooling of the partially adiabatically compressed air between the said air compressor (1), acting as the first stage of the adiabatic air-compression process, and an additional second compressor stage (3), by means of ambient air or water.

3. A hybrid energy system according to claim 1, wherein a recuperative heat exchanger (recuperator) (4) is additionally fitted between the exit of the said gas turbine (24) and the discharge side of the said compressor (1), of such a size as to enable recovering of as much waste heat energy from the exhaust gas as possible and a short-term storage of both the compressed gas and the exhaust gas, that is, a complete internal isobaric heat-exchange between the compressed air and the low-pressure exhausted combustion gas exiting the said gas turbine (24), containing the following interconnected equipment/components: compressed-gas inlet (5) and outlet (6) compartments, respectively, of such a size as to enable a short-term storage of the compressed gas; two parallel tube sheets (7) for support and isolation of compressed-air tubes in the said recuperative heat exchanger (4), perforated with a pattern of holes designed to accept the tubes; several fixed/stationary baffle plates (8) mounted around the outside of the compressed-air tubes (in the recuperator shell space) for the purposes of prolonging the cross-path of exiting exhaust gas through the said recuperator (4).

4. A hybrid energy system according to claim 3, wherein a compressor intercooler (2) is additionally used to perform an intercooling of the partially adiabatically compressed air between the said air compressor (1), acting as the first stage of the adiabatic air-compression process, and an additional second compressor stage (3), by means of ambient air or water.

5. A hybrid energy system according to claim 1, wherein the hybrid system additionally employs a bottoming steam-turbine part in a gas-turbine and steam-turbine combined-cycle power plant, whereas the said steam-turbine part consists of the following interconnected equipment: a) an additional combustion chamber (35), which provides a typical supplementary firing of the exhaust gas-turbine gas, fully expanded in the said gas turbine (24), as needed, to the end of providing a necessary temperature for production of a desired amount of superheated steam in the said bottoming steam-turbine part of the combined-cycle gas-turbine-steam-turbine power-plant; b) a heat recovery boiler (45) for raising of a desired quantity of superheated steam to be expanded in a typical steam turbine of the said bottoming steam-turbine part of the combined-cycle power-plant, containing the following main components: a water heater (economizer) and evaporator, a steam drum (49) for separation of gas and liquid phases (steam and water), a steam superheater and, optionally, a steam reheater; c) the said typical three-cylinder condensing steam turbine enabling a full expansion of the superheated steam raised in the said heat recovery boiler (45) to the lowest cycle pressure in a condenser (44); whereby the said condensing steam turbine consists of: a high-pressure cylinder (41) supplied by superheated steam from the superheater of the said heat recovery boiler (45), an intermediate-pressure cylinder (42) supplied by superheated steam from the reheater of the said heat recovery boiler (45), and a low-pressure cylinder (43); d) the said condenser (44) enabling full liquefaction (condensation) of the steam fully expanded in the said steam-turbine cylinders (41, 42 and 43) to the lowest cycle pressure in the said condenser (44), being equipped with a necessary steam-ejection device for extraction of air and other non-condensable gases from the condensate; e) a condensate pump (46) for pressurizing and circulation of the water condensed in the said condenser; f) an open feedwater tank (47) with deaerator, being fed with an intermediate-pressure steam extracted at the exit of the said intermediate-pressure cylinder (42) of the said condensing steam turbine, intended primarily for degassing of the incoming condensate/feedwater, but also for corresponding condensation of the steam extracted from the intermediate-pressure steam turbine, by mixing with the bulk of the condensate/feedwater in the said feedwater storage tank (47); g) a feedwater pump (48) for pressurizing and circulation of the feedwater from the said feedwater tank (47); and h) an additional electrical generator (34) for additional electricity generation, connected to the said steam-turbine cylinders (41, 42 and 43) via a common rotating shaft.

6. A hybrid energy system according to claim 5, wherein a compressor intercooler (2) is additionally used to perform an intercooling of the partially adiabatically compressed air between the said air compressor (1), acting as the first stage of the adiabatic air-compression process, and an additional second compressor stage (3), by means of ambient air or water.

7. A hybrid energy system according to claims 1, 3 and 5, wherein the said pneumatic motor (10) uses four single-acting cylinders, or optionally two double-acting cylinders, to the end of equalizing of a potentially fluctuating exhaust-gas flow rate and torque from the said pneumatic motor (10) and providing a balanced and smooth operation of the said air compressor (1).

8. A hybrid energy system according to claims 2, 4 and 6, wherein the said pneumatic motor (10) uses four single-acting cylinders, or optionally two double-acting cylinders, to the end of equalizing of a potentially fluctuating exhaust-gas flow rate and torque from the said pneumatic motor (10) and providing a balanced and smooth operation of the said first and second air-compressor stages (1 and 3).

Description

BRIEF DESCRIPTION OF THE DRAWINGS

[0023] FIG. 1 depicts a flow diagram of the hybrid energy system using a simple gas-turbine power-plant configuration with a pneumatic motor with two (2) single-acting cylinders.

[0024] FIG. 2 depicts a version of the flow diagram of the hybrid energy system configuration depicted in FIG. 1 which uses precooling of the gas-turbine cooling air.

[0025] FIG. 3 depicts a flow diagram of the hybrid energy system using a recuperated gas-turbine power-plant configuration with a pneumatic motor with two (2) single-acting cylinders.

[0026] FIG. 4 depicts a version of the flow diagram of the hybrid energy system configuration depicted in FIG. 3 which uses precooling of the gas-turbine cooling air.

[0027] FIG. 5 depicts a flow diagram of the hybrid energy system using an intercooled gas-turbine power-plant configuration with a pneumatic motor with two (2) single-acting cylinders, without or with optional precooling of the gas-turbine cooling air.

[0028] FIG. 6 depicts a flow diagram of the hybrid energy system using an intercooled-recuperated gas-turbine power-plant configuration with a pneumatic motor with two (2) single-acting cylinders, without or with optional precooling of the gas-turbine cooling air.

[0029] FIG. 7 depicts a flow diagram of the hybrid energy system using a simple combined gas-turbine/steam-turbine power-plant configuration with a pneumatic motor with two (2) single-acting cylinders, without or with optional precooling of the gas-turbine cooling air.

[0030] FIG. 8 depicts a flow diagram of the hybrid energy system using an intercooled combined gas-turbine/steam-turbine power-plant configuration with a pneumatic motor with two (2) single-acting cylinders, without or with optional precooling of the gas-turbine cooling air.

[0031] FIGS. 9 and 10 depict flow diagrams of the hybrid-energy-system configurations depicted in FIGS. 1, 2, 3 and 4, respectively, using a simple and a recuperated gas-turbine power-plant cycle with a pneumatic motor with four (4) single-acting cylinders, without or with optional precooling of the gas-turbine cooling air.

[0032] FIGS. 11 and 12 depict flow diagrams of the hybrid-energy-system configurations depicted in FIGS. 5 and 6, respectively, using an intercooled and an intercooled-recuperated gas-turbine power-plant configuration with a pneumatic motor with four (4) single-acting cylinders, without or with optional precooling of the gas-turbine cooling air.

[0033] FIGS. 13 and 14 depict graphs of the net plant thermal efficiency vs. maximum heat-input temperature and of the net plant thermal efficiency vs. net plant specific work, respectively, comparing the two hybrid-energy-system configurations using simple gas-turbine power-plant and simple combined gas-turbine/steam-turbine power-plant, depicted in FIGS. 1 & 2 (or FIG. 9) and FIG. 7, respectively.

[0034] FIGS. 15 and 16 depict graphs of the net plant thermal efficiency vs. maximum heat-input temperature and of the net plant thermal efficiency vs. net plant specific work, respectively, comparing the two hybrid-energy-system configurations using intercooled gas-turbine power-plant and intercooled combined gas-turbine/steam-turbine power-plant, depicted in FIG. 5 (or FIG. 11) and FIG. 8, respectively.

[0035] FIGS. 17 and 18 depict graphs of the net plant thermal efficiency vs. maximum heat-input temperature and of the net plant thermal efficiency vs. net plant specific work, respectively, for the hybrid-energy-system configurations using recuperated gas-turbine power-plant, depicted in FIGS. 3 & 4 (or FIG. 10), respectively.

[0036] FIGS. 19 and 20 depict graphs of the net plant thermal efficiency vs. maximum heat-input temperature and of the net plant thermal efficiency vs. net plant specific work, respectively, for the hybrid-energy-system configuration using intercooled-recuperated gas-turbine power-plant, depicted in FIG. 6 (or FIG. 12).

[0037] FIGS. 21 and 22 depict graphs of the net plant thermal efficiency vs. maximum heat-input temperature and of the net plant thermal efficiency vs. net plant specific work, respectively, comparing the two hybrid-energy-system configurations using recuperated gas-turbine power-plant and simple combined gas-turbine/steam-turbine power-plant, depicted in FIGS. 3 & 4 (or FIG. 10) and FIG. 7, respectively.

DETAILED DESCRIPTION OF THE INVENTION CONFIGURATIONS

[0038] In general, the direction of flow of the various working media on all flow diagrams is marked with arrows: solid line denotes the gaseous working fluid flow or the cooling-water flow (where applicable), dashed line denotes an optional/alternative gaseous working fluid flow and cooling-fluid flow, dash-dot line denotes center lines, while dash-double-dot line denotes fuel supply. All flow diagrams shown in different figures that correspond substantially to one another are arranged so that corresponding reference numbers and explanations are valid for common components depicted in such circuit diagrams. Therefore, explanations of such common components are not repeated in the description of similar figures.

[0039] A flow diagram of the first preferred hybrid-energy-system configuration with a gas-turbine power-plant in the simple open-loop cycle and a reciprocating pneumatic motor with isobaric internal combustion with two single-acting cylinders is depicted in FIG. 1 and it consists of the following interconnected equipment/processes: [0040] an air compressor 1 of axial, centrifugal, screw or reciprocating type, intaking atmospheric air and compressing it adiabatically from atmospheric pressure to a higher working pressure (adiabatic process “1”-“2”), equipped with an air pressure regulator (not depicted) with a screw for pressure setting, which is coupled with a gas turbine 24 and an additional load 22 by means of a common shaft; [0041] a pneumatic motor 10 (compressed-air engine) performing a non-adiabatic isobaric expansion and isobaric combustion process “2”-“3” (for air as an ideal gas) of the compressed combustion gas, transferring the compressed-gas power into the reciprocating linear motion of pistons 17 in two horizontal or vertical single-acting cylinders 11 and 12, and then into rotational motion, thus providing a driving force of the pneumatic motor with internal combustion; although two single-acting pneumatic-motor cylinders 11 and 12 are shown with reciprocating linear motion, one double-acting cylinder may be used; whereby the said pneumatic motor 10 with reciprocating linear motion of pistons contains the following interconnected equipment/components: at least two said single-acting cylinders 11 and 12 (or at least one double-acting cylinder) of a circular, quadratic, rectangular or triangular cross-sectional shape; associated pistons 17 with reciprocating linear motion through the said cylinders 11 and 12, each piston being equipped with at least two (2) or preferably three (3) piston rings, like in internal-combustion engines, sealing the said pneumatic motor 10 so that gases cannot escape it, whereby one or two upper/inner piston rings serve primarily for compression sealing (compression rings), whereas the lower/outer ring (oil control ring) serves for controlling the supply of lubrication oil to the said pistons 17 and the said compression rings; associated openings/valves for inlet (14) of the compressed-heated-gas to the said cylinders and outlet (15) of the exhaust gas from the said cylinders, respectively; connecting rods 18 providing physical connection between the said pistons 17 and a crankshaft 19, which facilitates conversion/transformation of the reciprocating linear motion of the pistons into a rotational motion; a camshaft 29 equipped with cams/eccentricities and accurately adjusted with motion of the said crankshaft 19, facilitating adequate and timely alternate opening and closing of the said opening/closing valves 14/15, respectively, by means of its rotation; and a timing belt/timing chain 28 providing an indirect connection and an accurate transmission of motion from the said crankshaft 19 to the said camshaft 29; [0042] openings for injection of fuel in complete with electric fuel igniters (13), comprised within the said pneumatic motor 10, for a complete isobaric combustion (isobaric process “2”-“3”) of a gaseous (typically natural gas) or a liquid fuel in the stream of compressed air, whereby the desirable locations/ways of injecting the gaseous or liquid fuel and igniting it by an electrical spark (like in a spark-ignition engine) into an operating cylinder of the said pneumatic motor 10 are three-fold: (i) at the top dead center of the operating cylinder, (like in a classic spark-ignition internal-combustion engine); (ii) at the upper side of the operating piston using a flexible fuel pipe 26 inserted thru the piston, so that the fuel ignites during the piston movement towards the bottom dead center and stops igniting when the bottom dead center has been reached; and (iii) combined injection and ignition of fuel by simultaneous use of the said methods (i) and (ii); [0043] a closed-loop lubrication system of the said pneumatic motor 10 similar as in internal-combustion engines, whereby lubricating oil is sucked out of an oil sump/tank by an oil pump and then forced through an oil filter to the main bearings of the said pneumatic motor 10 and then it passes through feed-holes into drilled passages in the said crankshaft 19 and onto the big-end bearings of the said connecting rods 18, whereas the cylinders walls and piston-pin bearings of the said connecting rods 18 are lubricated by oil drops dispersed by the rotating crankshaft 19, the excess of the lubricating oil being scraped off by the said lower rings in the said pistons 17, whereas a small fraction of the oil is bled from the main supply passage feeding each bearing of the said camshaft 29, said valves (14 and 15) and valves' springs, while another oil bleed supplies the said timing belt/chain 28 and gears on the said camshaft (29) drive, the excess of lubricating oil being drained back to the said oil sump, where eventually collected heat is being dispersed to the surrounding air; [0044] preferably, a flywheel 20 of the said pneumatic motor 10 for maintaining rotational speed of the said crankshaft 19 using its inertial forces (moment of inertia), thus equalizing a potentially fluctuating torque of the pneumatic motor during startup/operation/transients; [0045] a gearbox 21 for transmission of relatively slow rotational speed of the said crankshaft 19 into a rotational speed needed for a rotor of an electric generator; [0046] a load, typically an electric generator 23 for electricity generation, connected to the said pneumatic-motor crankshaft 19 by means of the said gearbox 21; [0047] a well-insulated combusted-gas collecting tank 27 provided at the outlet of the said pneumatic motor 10, for combusted-gas storage and equalizing of a potentially fluctuating combusted-gas flow rate from the said pneumatic motor 10; [0048] a motorized butterfly valve 25 for adjusting/regulating of the exhaust-gas pressure at the outlet of the said pneumatic motor 10, that is, at the inlet of a combustion-gas turbine 24, at a level close to or slightly lower than the maximum compressed-air pressure; [0049] the said combustion gas turbine 24 for a full adiabatic expansion process “3”-“4” of the combusted gas from the said single-acting cylinders 11 and 12 of the said pneumatic motor 10, driving both the said air compressor 1 and the mentioned additional load 22 by means of a common shaft; [0050] a cooling-air line for necessary cooling of profiles (stator, rotor) of the said gas turbine 24, by means of branching-off a small fraction of the compressed air from the outlet of the said air compressor 1 and before the inlet to the said pneumatic motor 10; and [0051] the said additional load, typically an electrical generator 22, for additional electricity generation, connected to the said common shaft of the said combustion-turbine 24 and the said air compressor 1.

[0052] FIG. 2 depicts a flow diagram of a version of the hybrid-energy-system configuration with a gas-turbine power-plant in the simple cycle and a reciprocating pneumatic motor with two single-acting cylinders depicted in FIG. 1, additionally employing an ambient-water-precooling of cooling air of the said gas turbine 24 profiles (isobaric cooling process “2”-“2′”) be means of an additional cooling-air precooler 9, to the end of reducing of the cooling-air fraction for the same or a similar gas-turbine cooling effect.

[0053] FIG. 3 depicts a flow diagram of a recuperated version of the hybrid-energy-system configuration with a gas-turbine power-plant in the simple cycle and a reciprocating pneumatic motor with two single-acting cylinders depicted in FIG. 1, employing a highly-effective recuperative heat exchanger/recuperator 4 of such a size as to enable recovering of as much waste heat energy from the exhaust gas as possible and a short-term storage of both the compressed gas and the exhaust gas, that is, a complete internal isobaric heat-exchange between the compressed air (isobaric process “2”-“3”) and the exhausted combustion gas exiting the said gas turbine 24 (isobaric process “5”-“6”), containing the following interconnected equipment/components: compressed-gas inlet and outlet compartments 5 and 6, respectively, of such a size as to enable a short-term storage of the compressed gas; two parallel tube sheets 7 for support and isolation of compressed-air tubes in the said recuperative heat exchanger 4, perforated with a pattern of holes designed to accept the tubes; several fixed/stationary baffle plates 8 mounted around the outside of the compressed-air tubes (in the recuperator shell space) for the purposes of prolonging the cross-path of exiting exhaust gas through the said recuperator 4 and thus resulting in a better gas-to-air heat transfer and a higher recuperator effectiveness.

[0054] FIG. 4 depicts a flow diagram of a version of the hybrid-energy-system configuration with a gas-turbine power-plant in the recuperated cycle and a reciprocating pneumatic motor with two single-acting cylinders depicted in FIG. 3, additionally employing an ambient-water-precooling of cooling air of the said gas turbine 24 profiles (isobaric cooling process “2”-“2′”) be means of an additional cooling-air precooler 9, to the end of reducing of the cooling-air fraction for the same or a similar gas-turbine cooling effect.

[0055] FIG. 5 depicts a flow diagram of an intercooled version of the hybrid-energy-system configuration with a gas-turbine power-plant in the simple cycle and a reciprocating pneumatic motor with two single-acting cylinders, with optional water-precooling of cooling air of the said gas turbine 24 profiles (isobaric cooling process “4”-“4′”), depicted in FIGS. 1 and 2, in a gas-turbine cycle with intercooling, additionally introducing a typical intercooling of the partially compressed air between the said compressor 1 acting as the first stage of an adiabatic compression process “1”-“2”, and a second compressor stage 3 (adiabatic compression process “3”-“4”) in an intercooler 2 (isobaric cooling process “2”-“3”), by means of ambient air or water, to the end of reducing the mechanical work required for driving of the said air compressor stages 1 and 3 and thus achieving a greater cycle specific work.

[0056] FIG. 6 depicts a flow diagram of an intercooled-recuperated version of the hybrid-energy-system configuration with a gas-turbine power-plant in the recuperated cycle and a reciprocating pneumatic motor with two single-acting cylinders, with optional water-precooling of cooling air of the said gas turbine 24 profiles (isobaric cooling process “4”-“4′”), depicted in FIGS. 3 and 4, in a gas-turbine cycle with recuperation and intercooling, additionally employing a typical intercooling of the partially compressed air between the said compressor 1 acting as the first stage of an adiabatic compression process “1”-“2”, and a second compressor stage 3 (adiabatic compression process “3”-“4”) in an intercooler 2 (isobaric cooling process “2”-“3”), by means of ambient air or water, to the end of reducing the mechanical work required for driving of the said air compressor stages 1 and 3 and thus achieving a greater cycle specific work.

[0057] FIG. 7 depicts a flow diagram of a version of the hybrid-energy-system configuration with a gas-turbine power-plant in the simple cycle and a reciprocating pneumatic motor with two single-acting cylinders, with optional water-precooling of cooling air of the said gas turbine 24 profiles (isobaric cooling process “2”-“2′”), depicted in FIG. 1, in a combined-cycle of a gas turbine and a steam turbine, optionally using a typical supplementary firing of the exhaust combusted gas, fully expanded in the said gas turbine 24, in an additional combustion chamber 35 (isobaric heating process “4”-“5”), to the end of providing necessary temperature for production of a desired amount of steam in the bottoming steam-cycle power-plant of the combined gas-turbine-steam-turbine power-plant, whereas the said steam-turbine part of the said combined gas-turbine-steam-turbine power-plant consists of the following interconnected equipment/processes: [0058] a heat recovery boiler 45 for raising of a desired quantity of superheated steam to be expanded in a typical steam turbine of the said bottoming steam-cycle power-plant, containing: a water heater (economizer) and evaporator, a steam drum 49 for separation of gas and liquid phases (steam and water), a steam superheater and optionally a steam reheater; [0059] the said typical three-cylinder condensing steam turbine enabling a full expansion of the superheated steam raised in the said heat recovery boiler 45 to the lowest cycle pressure in a condenser 44; the said condensing steam turbine consisting of: a high-pressure cylinder 41 supplied by superheated steam from the said superheater of the said heat recovery boiler 45, an intermediate-pressure cylinder 42 supplied by superheated steam from the said reheater of the said heat recovery boiler 45, and a low-pressure cylinder 43; [0060] the said condenser 44 enabling full liquefaction (condensation) of the steam fully expanded in the said steam-turbine cylinders (41, 42 and 43) to the lowest cycle pressure in the said condenser 44, being equipped with a necessary steam-ejection device for extraction of air and other non-condensable gases from the condensate; [0061] a condensate pump 46 for pressurizing and circulation of the water condensed in the said condenser 44; [0062] an open feedwater tank with deaerator 47, being fed with an intermediate-pressure steam extracted at the exit of the said intermediate-pressure condensing-steam-turbine cylinder 42, intended primarily for degassing of the incoming condensate/feedwater, as well as a corresponding condensation of the steam extracted from the intermediate-pressure steam-turbine cylinder by means of its mixing with the bulk of the condensate/feedwater in the said deaerator storage tank 47; [0063] a feedwater pump 48 for pressurizing and circulation of the feedwater from the said feedwater tank 47; and [0064] an additional electrical generator 34 for additional electricity generation, connected to the said steam-turbine cylinders (41, 42 and 43) via a common rotating shaft.

[0065] FIG. 8 depicts a flow diagram of an intercooled version of the hybrid-energy-system configuration with a gas-turbine power-plant in the simple-combined gas-turbine/steam-turbine cycle and a reciprocating pneumatic motor with two single-acting cylinders, with optional water-precooling of cooling air of the said gas turbine 24 profiles (isobaric cooling process “4”-“4′”), depicted in FIG. 7, in a gas-turbine cycle with intercooling, additionally employing a typical intercooling of the partially compressed air between the said compressor 1 acting as the first stage of an adiabatic compression process “1”-“2”, and a second compressor stage 3 (adiabatic compression process “3”-“4”) in an intercooler 2 (isobaric cooling process “2”-“3”), by means of ambient air or water, to the end of reducing the mechanical work required for driving of the said air compressor stages 1 and 3 and thus achieving a greater cycle specific work.

[0066] FIG. 9 depicts a flow diagram of the hybrid-energy-system configurations with a gas-turbine power-plant in the simple cycle depicted in FIGS. 1 and 2, with optional water-precooling of cooling air of the said gas turbine 24 profiles (isobaric cooling process “2”-“2′”), which uses a pneumatic motor with four single-acting cylinders, or optionally with two double-acting cylinders, to the end of equalizing of a potentially fluctuating combusted-gas flow rate and torque from the pneumatic motor and providing a more balanced and a smoother operation of the said air compressor 1.

[0067] FIG. 10 depicts a flow diagram of the hybrid-energy-system configurations with a gas-turbine power-plant in the recuperated cycle depicted in FIGS. 3 and 4, with optional water-precooling of cooling air of the said gas turbine 24 profiles (isobaric cooling process “2”-“2′”), which uses a pneumatic motor with four single-acting cylinders, or optionally with two double-acting cylinders, to the end of equalizing of a potentially fluctuating combusted-gas flow rate and torque from the pneumatic motor and providing a more balanced and a smoother operation of the said air compressor 1.

[0068] FIG. 11 depicts a flow diagram of the hybrid-energy-system configuration with a gas-turbine power-plant in the intercooled cycle depicted in FIG. 5, with optional water-precooling of cooling air of the said gas turbine 24 profiles (isobaric cooling process “4”-“4′”), which uses a pneumatic motor with four single-acting cylinders, or optionally with two double-acting cylinders, to the end of equalizing of a potentially fluctuating combusted-gas flow rate and torque from the pneumatic motor and providing a more balanced and a smoother operation of the said air compressors 1 and 3.

[0069] FIG. 12 depicts a flow diagram of the hybrid-energy-system configuration with a gas-turbine power-plant in the intercooled-recuperated cycle, depicted in FIG. 6, with optional water-precooling of cooling air of the said gas turbine 24 profiles (isobaric cooling process “4”-“4′”), which uses a pneumatic motor with four single-acting cylinders, or optionally with two double-acting cylinders, to the end of equalizing of a potentially fluctuating combusted-gas flow rate and torque from the pneumatic motor and providing a more balanced and a smoother operation of the said air compressors 1 and 3.

[0070] Applied Mathematical Model

[0071] A brief thermodynamic and fluid mechanic analysis of applied simple mathematical model of this concept is presented hereinafter.

[0072] General assumption No. 1 presumes equality/conservation of volumetric and mass flow rates of compressible gaseous fluid (air) in the pneumatic motor with internal combustion.

[0073] General assumption No. 2 presumes existence of isobaric (constant-pressure) flow of compressed heated compressible gas (compressed hot air) throughout the pneumatic-motor cylinder, from a smaller to a larger volume. This general assumption thus presumes that the compressed compressible fluid performs an isobaric work during the process of isobaric expansion between a smaller and a larger volume in the pneumatic-motor cylinder.

[0074] General assumption No. 3 presumes that the compressed compressible fluid flows in an isenthalpic (isothermal for ideal gases) manner during the processes of flowing in and out of the pneumatic-motor cylinder.

[0075] There are several basic thermodynamic and fluid mechanic equations applied in this mathematical model. The first and the second equation give the well-known relationships of the temperature- and pressure-ratios of the compressible-fluid compression and expansion processes, respectively, based on isentropic efficiencies of the compressor and the gas turbine:

[00001]$\begin{array}{cc}\frac{T_{\mathrm{Com},\mathrm{out}}}{T_{\mathrm{Com},in}}=1+\left[{\left(\frac{p_{\mathrm{Com},\mathrm{out}}}{p_{\mathrm{Com},in}}\right)}^{\frac{\left({\kappa }_C-1\right)}{{\kappa }_C}}-1\right].Math.\frac{1}{{\eta }_{iC}}=1+\left[\mathrm{CP}{R}^{\frac{R_{air}}{C_{p,C}}}-1\right].Math.\frac{1}{{\eta }_{iC}}& \left[1\right]\\ \frac{T_{\mathrm{GT},\mathrm{out}}}{T_{\mathrm{GT},in}}=1-\left[1-{\left(\frac{p_{\mathrm{GT},in}}{p_{\mathrm{GT},\mathrm{out}}}\right)}^{\frac{\left(1-{\kappa }_{\mathrm{EX}}\right)}{{\kappa }_{\mathrm{EX}}}}\right].Math.{\eta }_{iEX}=1-\left[1-{\mathrm{EPR}}^{\frac{\left(-R_{air}\right)}{C_{p,\mathrm{EX}}}}\right].Math.{\eta }_{iEX}& \left[2\right]\end{array}$

where: T.sub.Com,in (15° C. or 288 K by default) and T.sub.Com,out (in K) are air static temperatures before and after compression, respectively; p.sub.Com,in (in bar or kPa) and p.sub.Com,out (in bar or kPa) are air static pressures before and after compression, respectively; CPR is the compressor pressure ratio (p.sub.Com,out/p.sub.Com,in); R.sub.air is the air gas constant (0.287 kJ/kg*K by default); k.sub.c is an average ratio of the specific heats (C.sub.p/C.sub.v) of air during compression (1.40 by default); C.sub.p,c (in kJ/[kg*K]) is mean constant-pressure specific heat of air during compression, assumed to have a value of ˜1.005 kJ/(kg*K) corresponding to the value of 1.40 of an average ratio of specific heats (C.sub.p/C.sub.v) of air during compression process; T.sub.GT,in (in K) and T.sub.GT,out (in K) are exhaust-gas static temperatures before and after expansion, respectively; p.sub.GT,in (in bar or kPa) and p.sub.GT,out (in bar or kPa) are exhaust-gas static pressures before and after expansion, respectively; EPR is the expansion pressure ratio (p.sub.GT,in/p.sub.GT,out), k.sub.EX is an average ratio of the specific heats (C.sub.p/C.sub.v) of exhaust gas during expansion (typically 1.33); C.sub.p,EX (in kJ/[kg*K]) is mean constant-pressure specific heat for expansion and combustion processes in the gas turbine and the combustion chamber of the power-plant, assumed to have a value of ˜1.157 kJ/(kg*K) corresponding to the value of 1.33 of an average ratio of specific heats (C.sub.p/C.sub.v) of air during expansion and combustion processes; η.sub.iC is the compressor isentropic efficiency, assumed to have a constant value of 85%, regardless of the compressor pressure ratio and losses; and η.sub.iEX is the gas-turbine isentropic efficiency, assumed to have a constant value of 90%, regardless of the expansion pressure ratio and losses.

[0076] The expansion pressure ration (EPR) is defined by the following expression:

[00002]$\begin{array}{cc}\mathrm{EPR}=\frac{p_{\mathrm{GT},in}}{p_{\mathrm{GT},\mathrm{out}}}=\mathrm{CPR}.Math.\frac{{\left(0.95\right)}^3.Math.{\left(0.95\right)}^2.Math.{\left(0.975\right)}^{\left(N_{\mathrm{Com}}-1\right)}}{\left(1.11\right)}& \left[3\right]\end{array}$

where: factor (0.95).sup.3 involves inevitable parasitic pressure losses, assumed to be 5% through the combustion chamber, 5% through the compressed-air path side of the recuperative heat exchanger (if there is any in the GT configuration) and also 5% through pneumatic-motor components (pipes, inlet opening/valve); factor (0.95).sup.2 similarly involves inevitable parasitic pressure losses, assumed to be 5% through the exhaust-gas path side of the recuperative heat exchanger (if there is any in the GT configuration) and also 5% through the pneumatic-motor components (outlet opening/valve, pipes); factor (0.975).sup.(Ncom-1) similarly involves inevitable parasitic pressure losses, assumed to be 2.5% through any intercooler (if there is any in the GT configuration) of the partially-compressed-air, whereas N.sub.com is the number of air-compressor stages; and factor (1.11) takes into account additional pressure losses due to flowing of the exhaust GT gas to atmosphere at a velocity greater than zero.

[0077] The third applied equation is a thermodynamics expression for the specific work/output of the pneumatic motor, w.sub.PM, which is proportional to the difference of volumes of a pneumatic-motor cylinder during the suction of the pressurized air/gas at a constant pressure, as follows:

[00003]$\begin{array}{cc}{w}_{PM}=p.Math.\frac{\Delta V}{m_{air}}.Math.{\eta }_{PM}=p.Math.\frac{\left(V_{m\mathrm{ax}}-V_{min}\right)}{m_{air}}.Math.{\eta }_{PM}=R_{air}.Math.\left(T_{m\mathrm{ax}}-T_{\mathrm{Com},\mathrm{out}}\left[\mathrm{or}{T}_{rec}\right]\right).Math.{\eta }_{PM}\left[\frac{\mathrm{kJ}}{\mathrm{kg}}\right]& \left[4\right]\end{array}$

where: m.sub.air (in kg) is total mass of the compressed heated working gas contained in a pneumatic-motor cylinder; R.sub.aft is the gas constant for air (0.287 kJ/(kg*K) by default); p.sub.Com,out (kPa) is a maximum static pressure through the pneumatic motor; V.sub.mm (m.sup.3) is a minimum volume of a pneumatic-motor cylinder; V.sub.max (m.sup.3) is the maximum volume of a pneumatic-motor cylinder; T.sub.max (K) is the maximum air/gas static temperature at the outlet of the combustion chamber (the pneumatic motor); T.sub.rec (in K) is the static air temperature at the recuperative-heat exchanger-outlet (if there is any in the GT configuration), that is, at the combustor inlet; and η.sub.PM is mechanical efficiency of the pneumatic motor (taking into account mechanical losses in the gearbox and the electric motor), assumed to amount to 98%.

[0078] The fourth applied equation presents an energy equation for the combustion process in the combustion chamber. Specific heat input, q.sub.in (in kJ/kg), of the pneumo-engine can be expressed from the corresponding energy equation:

[00004]$\begin{array}{cc}{q}_{in}=C_{p,\mathrm{EX}}.Math.\left(T_{m\mathrm{ax}}-T_{rec}\right)=C_{p,\mathrm{EX}}\left\{T_{m\mathrm{ax}}-.Math.T_{\mathrm{Com},\mathrm{out}}+{\eta }_{rec}.Math.\left(T_{\mathrm{GT},\mathrm{out}}-T_{\mathrm{Com},\mathrm{out}}\right).Math.\right\}==C_{p,\mathrm{EX}}\left[T_{m\mathrm{ax}}-T_{\mathrm{Com},\mathrm{out}}.Math.\left(1-{\eta }_{rec}\right)-T_{\mathrm{GT},\mathrm{out}}.Math.{\eta }_{rec}\right]\left[\frac{\mathrm{kJ}}{\mathrm{kg}}\right]& \left[5\right]\end{array}$

where: T.sub.max (in K) is the maximum air/exhaust-gas static temperature at the combustor outlet; and η.sub.rec (is effectiveness of the recuperative heat exchanger (if there is any in the GT configuration).

[0079] Total cycle thermal efficiency (η.sub.th,PM-GT) of the hybrid energy system is given in the form of the ratio net-power-output/total-heat-input, taking into account inefficiencies in the compression and expansion processes, as follows:

[00005]$\begin{array}{cc}{\eta }_{\mathrm{th},\mathrm{PM}\text{-}\mathrm{GT}}=\frac{\begin{array}{c}{w}_{\mathrm{PM}}.Math.\left(1-\frac{{\mathrm{cool}}_{}}{{\mathrm{air}}_{}}\right)+\\ C_{p,\mathrm{EX}}.Math.\left(T_{\mathrm{GT},in}-T_{\mathrm{GT},\mathrm{out}}\right)-C_{p,C}.Math.\left(T_{\mathrm{Com},\mathrm{out}}-T_{\mathrm{Com},in}\right)\end{array}}{q_{in}.Math.\left(1-\frac{{\mathrm{cool}}_{}}{{\mathrm{air}}_{}}\right)}& \left[6\right]\end{array}$

where: m.sub.air (in kg or kg/s) is total mass/mass flow rate of the compressed working gas at the compressor discharge; m.sub.cool (in kg or kg/s) is a fraction of the total mass/mass flow rate of the compressed air that is branched off before the said pneumatic motor-combustor to provide necessary cooling of profiles (stator, rotor) of the gas turbine; and η.sub.th,PM-GT is the total cycle thermal efficiency of the hybrid energy system with pneumatic motor.

[0080] Using the described mathematical model, the following graphs (FIGS. 13 through 22) were obtained, shown in the drawings section of this patent application. The first graph in FIG. 13 compares changes of the net plant thermal efficiency vs. maximum heat-input temperature for the two configurations of hybrid energy system gas turbine—pneumatic motor using simple gas-turbine power-plant and simple combined gas-turbine/steam-turbine power-plant, depicted in FIGS. 1 and 2 (or FIG. 9) and FIG. 7, respectively. From the graph it may be noticed that the cycle thermal efficiency logically increases with the maximum heat-input temperature, reaching the following peak values at the maximum considered heat-input temperature of 1,700 K (1,427° C.): 0.641 (64.1%) at a constant arbitrarily chosen optimum compressor pressure ratio (CPR) of 35:1 for the configuration depicted at FIGS. 1 and 2 (or FIG. 9), and 0.772 (77.2%) (!) at a constant arbitrarily chosen optimum compressor pressure ratio (CPR) of 15:1 and at an arbitrarily chosen supplementary-firing temperature of 900 K for the configuration depicted at FIG. 7.

[0081] The second graph in FIG. 14 compares changes of the net plant thermal efficiency vs. net plant specific work/output for the two configurations of hybrid energy system gas turbine—pneumatic motor using simple gas-turbine power-plant and simple combined gas-turbine/steam-turbine power-plant, depicted in FIGS. 1 and 2 (or FIG. 9) and FIG. 7, respectively. From the graph it may be noticed that the net cycle specific work also increases with the maximum heat-input temperature, reaching the following peak values at the maximum considered heat-input temperature of 1,700 K (1,427° C.): 438.6 kJ/kg at a constant arbitrarily chosen optimum compressor pressure ratio (CPR) of 35:1 for the configuration depicted at FIGS. 1 and 2 (or FIG. 9), and 715.6 kJ/kg at a constant arbitrarily chosen optimum compressor pressure ratio (CPR) of 15:1 and at an arbitrarily chosen supplementary-firing temperature of 900 K (627° C.) for the configuration depicted at FIG. 7.

[0082] The third graph in FIG. 15 compares changes of the net plant thermal efficiency vs. maximum heat-input temperature for the two configurations of hybrid energy system gas turbine—pneumatic motor using intercooled gas-turbine power-plant and intercooled combined gas-turbine/steam-turbine power-plant, depicted in FIG. 5 (or FIG. 11) and FIG. 8, respectively. From the graph it may be noticed that the cycle thermal efficiency logically increases with the maximum heat-input temperature, reachin, the following peak values at the maximum considered heat-input temperature of 1,700 K (1,427° C.): 0.635 (63.5%) at a constant arbitrarily chosen optimum compressor pressure ratio (CPR) of 60:1 for the configuration depicted at FIG. 5 (or FIG. 11), and 0.722 (72.2%) at a constant arbitrarily chosen optimum compressor pressure ratio (CPR) of 17:1 and at an arbitrarily chosen supplementary-firing temperature of 900 K (627° C.) for the configuration depicted at FIG. 8.

[0083] The fourth graph in FIG. 16 compares changes of the net plant thermal efficiency vs. net plant specific work/output for the two configurations of hybrid energy system gas turbine—pneumatic motor using intercooled gas-turbine power-plant and intercooled combined gas-turbine/steam-turbine power-plant, depicted in FIG. 5 (or FIG. 11) and FIG. 8, respectively. From the graph it may be noticed that the net cycle specific work also increases with the maximum heat-input temperature, reaching the following peak values at the maximum considered heat-input temperature of 1,700 K (1,427° C.): 674.2 kJ/kg at a constant arbitrarily chosen optimum compressor pressure ratio (CPR) of 60:1 for the configuration depicted at FIG. 5 (or FIG. 11), and 871.3 kJ/kg at a constant arbitrarily chosen optimum compressor pressure ratio (CPR) of 17:1 and at an arbitrarily chosen supplementary-firing temperature of 900 K (627° C.) for the configuration depicted at FIG. 8.

[0084] The fifth graph and the sixth graph in FIGS. 17 and 18 depict changes of the net plant thermal efficiency vs. maximum heat-input temperature and the net plant thermal efficiency vs. net plant specific work/output, respectively, for the configuration of hybrid energy system gas turbine—pneumatic motor using recuperated gas-turbine power-plant with water-precooling of the gas-turbine cooling air, depicted in FIG. 4 (or FIG. 10). From the graph it may be noticed that both the cycle thermal efficiency and the net cycle specific work logically increase with the maximum heat-input temperature and with the recuperator effectiveness, reaching the following peak values at the maximum considered heat-input temperature of 1,700 K (1,427° C.), at a constant arbitrarily chosen optimum compressor pressure ratio (CPR) of 9:1 and at three different arbitrarily assumed values of the recuperator effectiveness (90%, 95% and 98%): 0.67 (67%) and 481.5 kJ/kg, 0.686 (68.6%) and 475.4 kJ/kg, and 0.696 (69.6%) and 471.7 kJ/kg, respectively.

[0085] Similarly, the seventh graph and the eighth graph in FIGS. 19 and 20 depict changes of the net plant thermal efficiency vs. maximum heat-input temperature and the net plant thermal efficiency vs. net plant specific work/output, respectively, for the configuration of hybrid energy system gas turbine—pneumatic motor using intercooled-recuperated gas-turbine power-plant with water-precooling of the gas-turbine cooling air, depicted in FIG. 6 (or FIG. 12). From the graph it may be noticed that both the cycle thermal efficiency and the net cycle specific work logically increase with the maximum heat-input temperature and with the recuperator effectiveness, reaching the following peak values at the maximum considered heat-input temperature of 1,700 K (1,427° C.), at a constant arbitrarily chosen optimum compressor pressure ratio (CPR) of 15:1 and at three different arbitrarily assumed values of the recuperator effectiveness (90%, 95% and 98%): 0.72 (72%) and 595.9 kJ/kg, 0.735 (73.5%) and 589.6 kJ/kg, and 0.745 (74.5%) and 585.8 kJ/kg, respectively.

[0086] Finally, the ninth graph and the tenth graph in FIGS. 21 and 22 compare changes of the net plant thermal efficiency vs. maximum heat-input temperature and the net plant thermal efficiency vs. net plant specific work/output, respectively, for the configurations of hybrid energy system gas turbine—pneumatic motor using recuperated gas-turbine power-plant with water-precooling of the gas-turbine cooling air and simple combined gas-turbine/steam-turbine power-plant, depicted in FIG. 4 (or FIG. 10) and FIG. 7, respectively. From the graph it may be noticed that both the cycle thermal efficiency and the net cycle specific work logically increase with the maximum heat-input temperature and the recuperator effectiveness, reaching the following peak values at the maximum considered heat-input temperature of 1,700 K (1,427° C.): 0.696 (69.6%) and 471.7 kJ/kg, respectively, at a constant arbitrarily chosen optimum compressor pressure ratio (CPR) of 9:1 and at an arbitrarily assumed value of the recuperator effectiveness of 98% for the configuration depicted at FIG. 4 (or FIG. 10), and 0.772 (77.2%) and 715.6 kJ/kg, respectively, at a constant arbitrarily chosen optimum compressor pressure ratio (CPR) of 15:1 and at an arbitrarily chosen supplementary-firing temperature of 900 K (627° C.) for the configuration depicted at FIG. 7.

[0087] All numbers expressing process or cycle parameters, cycle thermal efficiencies, specific cycle outputs, and so forth, used in this specification and claims are to be understood as being modified in all instances by the term “about” or “approximately”. The matter set forth in the foregoing description and accompanying drawings is offered by way of illustration only and not as a limitation. Since further modifications, applications or adaptations of the invention may become apparent to those skilled in the art, aim of the appended patent claims is to cover all such changes and modifications as fall within the true spirit and scope of the invention.