US20070245733A1 - Power recovery and energy conversion systems and methods of using same - Google Patents

Abstract

In various illustrative examples, the system may include heat recovery heat exchangers, one or more turbines or expanders, a desuperheater heat exchanger, a condenser heat exchanger, a separator, an accumulator, and a liquid circulating pump, etc. In one example, a bypass desuperheater control valve may be employed. The system comprises a first heat exchanger adapted to receive a heating stream from a heat source after passing through a second heat exchanger and a second portion of a working fluid, wherein, the second portion of working fluid is converted to a hot liquid via heat transfer. An economizer heat exchanger that is adapted to receive a first portion of the working fluid and the hot discharge vapor from at least one turbine may also be provided. The first and second portions of the working fluid are recombined in a first flow mixer after passing through the economizer heat exchanger and first heat exchanger, respectively. A second heat exchanger is provided that receives the working fluid from the first flow mixer and a hot heating stream from a heat source and convert the working fluid to a hot vapor. The hot vapor from the second heat exchanger is supplied to at least one turbine after passing through a separator designed to insure no liquid enters the said at least one turbine or expander. The hot, high pressure vapor is expanded in the turbine to produce mechanical power on a shaft and is discharged as a hot, low pressure vapor.

Description

BACKGROUND OF THE INVENTION
• 1. Field of the Invention
• The present invention generally relates to heat recovery for the purpose of electrical or mechanical power generation. Specifically, the present invention is directed to various systems and methods for the conversion of heat of any quality into mechanical or electrical power.
• 2. Description of the Related Art
• In general, there is a constant drive to increase the operating efficiency of heat and power recovery systems. By increasing the efficiency of such systems, capital costs may be reduced, more power may be generated and there may be a reduction of possible adverse impacts on the environment, e.g., a reduction in the amount of waste heat that must ultimately be absorbed by the environment. In other industrial processes, an excess amount of heat may be generated as a byproduct of the process. In many cases, such waste heat is normally absorbed by the environment through the use of waste heat rejection devices such as cooling towers.
• There are several systems employed in various industries to produce useful work from a heat source. Such systems may including the following:
• Heat Recovery Steam Generators (HRSG)—Typically, waste heat from gas turbines or other, similar, high quality heat sources is recovered using steam at multiple temperatures and pressures. Multiple operating levels are required because the temperature-enthalpy profile is not linear. That is, such prior art systems involve isothermal (constant temperature) boiling as the working fluid, i.e. water, is converted from a liquid to a vapor state. Various embodiments of the present invention eliminate the need for multiple levels and simplify the process while having the capability to recover more heat and to economically recover heat from a much lower quality heat source.
• Rankine Cycle—The classic Rankine cycle is utilized in conjunction with HRSGs to produce power. This process is complex and requires either multiple steam turbines or a multistage steam turbine, feed water heaters, steam drums, pumps, etc. The methods and systems of the present invention are significantly less complex while being more effective than systems employing the Rankine cycle.
• Organic Rankine Cycle—Similar to the classic Rankine cycle, an Organic Rankine cycle utilizes a low temperature working fluid such as isoButane or isoPentane in place of steam in the classic cycle. The system remains complex and is highly inefficient at low operating temperature differences.
• Kalina Cycle—Dr. Kalina's cycle is a next generation enhancement to the Rankine cycle utilizing a binary fluid mixture, typically water and ammonia. Water and ammonia are utilized at different concentrations in various portions of the process to extend the temperature range potential of the cycle and to allow higher efficiencies than are possible in the Rankine cycle. The methods and systems of the present invention simplify the process while having the capability to recover more heat and to recover heat from a low quality heat source.
• The system depicted inFIG. 5 is an example of a prior art system for heat recovery. The system comprises two heatrecovery heat exchangers120 and121, two turbines (expanders)122 and124, and areheater heat exchanger123. The prior art system may or may not have aseparate gas cooler125 andcondenser126. Thesubcritical working fluid102 enter the first heatrecovery heat exchanger120 at approximately the condensing temperature from acondenser126. Theliquid102 is heated via heat transfer with the dischargedhot fluid114 from thereheater heat exchanger123 and is discharged as either a wet ordry vapor103 after boiling either partially or completely in heatrecovery heat exchanger120. The workingfluid103 is further heated in the second heatrecovery heat exchanger121 to adry vapor104 via heat transfer with thehot heat source112 and is supplied to the inlet of thefirst turbine122. In at least some cases, thevapor104 is at a temperature near or slightly above its critical temperature but well below its critical pressure. Thehot vapor104 is expanded inturbine122 and exits as ahot vapor105. Thehot vapor105 is introduced into areheater heat exchanger123 where is heated (reheated) by thehot heating fluid113 discharged from the second heatrecovery heat exchanger121 via heat transfer. The reheated workingfluid106 is then supplied to the inlet of thesecond turbine124 wherein it is expanded and discharged as a hot, typically dry and highly superheated,vapor107. The dischargedvapor107 from thesecond turbine124 may or may not be cooled in agas cooler125 before being condensed in acondenser heat exchanger126.
• In the prior art system ofFIG. 5, thesubcritical working fluid102 enter the first heatrecovery heat exchanger120 at approximately the condensing temperature from acondenser126. Saidliquid102 is heated via heat transfer with the dischargedhot fluid114 from thereheater heat exchanger123 and is discharged as either a wet ordry vapor103 after boiling either partially or completely in heatrecovery heat exchanger120. Said workingfluid103 is further heated in the second heatrecovery heat exchanger121 to adry vapor104 via heat transfer with thehot heat source112 and is supplied to the inlet of thefirst turbine122. In the most preferred embodiment thevapor104 is at a temperature near or slightly above its critical temperature but well below its critical pressure. Thehot vapor104 is expanded inturbine122 and exits as ahot vapor105. Suchhot vapor105 is introduced into areheater heat exchanger123 where is heated (reheated) by thehot heating fluid113 discharged from the second heatrecovery heat exchanger121 via heat transfer. The reheated workingfluid106 is then supplied to the inlet of thesecond turbine124 wherein it is expanded and discharged as a hot, typically dry and highly superheated,vapor107. The dischargedvapor107 from thesecond turbine124 may or may not be cooled in agas cooler125 before being condensed in acondenser heat exchanger126.
• The four largest weaknesses of the prior art system are a) thevapor107 discharged from thesecond turbine124 is significantly superheated and thereby the system ofFIG. 5 fails to recover a portion of the valuable heat, b) the system utilizes a subcritical working fluid which limits the efficiency of the heat recovery in the heatrecovery heat exchangers120 and121 due to the non-linearity of the temperature-enthalpy profile in said exchangers, c) the system generates unnecessary entropy further reducing its output in accordance with the Second Law of Thermodynamics, and d) the complexity of the system having multiple turbines and multiple heat recovery heat exchangers is reflected in an increased cost of the system for a given capacity recovery heat exchanger(s) are usually the largest costs in a system of the type.
• The following patents may be descriptive of various aspects of the prior art: U.S. Pat. No. 5,557,936 to Drnevich; U.S. Pat. No. 5,029,444 to Kalina; U.S. Pat. No. 5,440,882 to Kalina; U.S. Pat. No. 5,095,708 to Kalina; U.S. Pat. No. 5,572,871 to Kalina; Japanese Patent S53-132638A to Nakahara and Fujiwara; U.S. Pat. No. 6,195,997 to Lewis; U.S. Pat. No. 4,577,112 to Smith; U.S. Pat. No. 6,857,268 to Stinger and Mian; each of which are hereby incorporated by reference.
• In general, what is desired are systems and methods for improving the efficiencies of various heat conversion and power generation systems and systems and methods for utilizing waste heat sources to improve operating efficiencies of various power and industrial systems. The present invention is directed to various systems and methods that may solve, or at least reduce, some or all of the aforementioned problems.
SUMMARY OF THE INVENTION
• The following presents a simplified summary of the invention in order to provide a basic understanding of some aspects of the invention. This summary is not an exhaustive overview of the invention. It is not intended to identify key or critical elements of the invention or to delineate the scope of the invention. Its sole purpose is to present some concepts in a simplified form as a prelude to the more detailed description that is discussed later.
• The present invention is generally directed to various systems and methods for producing mechanical power from a heat source. In various illustrative examples, the devices employed in practicing the present invention may include at least two heat recovery heat exchangers, at least one turbine or an expander, a desuperheater heat exchanger, an economizer heat exchanger, a condenser heat exchanger, an accumulator, a separator, and a liquid circulating pump, etc.
• In one illustrative embodiment, the system comprises a first heat exchanger adapted to receive a heating stream from a heat source after passing through a second heat exchanger and a second portion of a working fluid, wherein, when the second portion of the working fluid is passed through the first heat exchanger, the second portion of working fluid is converted to a hot liquid via heat transfer from the heat contained in the heating stream from the heat source after passing through a second heat exchanger. The system is further comprised of an economizer heat exchanger adapted to receive a first portion of the working fluid and the hot discharge vapor from at least one turbine. The first and second portions of the working fluid are recombined in a first flow mixer after passing through the economizer heat exchanger and first heat exchanger, respectively. The system is further comprised of a second heat exchanger adapted to receive the working fluid from the first flow mixer and a hot heating stream from a heat source and convert the working fluid to a hot vapor. The hot vapor from the second heat exchanger is supplied to at least one turbine or expander after passing through a separator designed to insure no liquid enters said at least one turbine. The hot, high pressure vapor is expanded in the turbine to produce mechanical power on a shaft and is discharged as a hot, low pressure vapor. The hot vapor is then routed back to the economizer heat exchanger and then to a second flow mixer (which may function as a desuperheater in some cases) where the hot vapor is mixed with the liquid discharged from the separator. The system further comprises a condenser heat exchanger that is adapted to receive the exhaust vapor from the turbine after passing through the economizer heat exchanger and mixing with the liquid from the separator and a cooling fluid circulated by a cooling fluid pump. The system is further comprised of an accumulator vessel to receive the condensed liquid from the condenser and meter said condensate to a liquid working fluid circulating pump that is adapted to circulate the working fluid to a flow divider. The system is finally comprised of a flow divider that is adapted to split the working fluid into at least two portions, at least one that is supplied to an economizer heat exchanger and at least one that is supplied to a first heat exchanger.
• In another illustrative embodiment, the system comprises a first heat exchanger adapted to receive a heating stream from a heat source after passing through a second heat exchanger and a second portion of a working fluid, wherein, the second portion of the working fluid is passed through the first heat exchanger, the second portion of working fluid is converted to a hot liquid via heat transfer from the heat contained in the heating stream from the heat source after passing through a second heat exchanger. The system is further comprised of an economizer heat exchanger adapted to receive a first portion of the working fluid and the hot discharge vapor from at least one turbine. The system is further comprised of a second flow mixer or desuperheater adapted to receive a third portion of the working fluid via a fluid bypass control valve. The first and second portions of the working fluid are recombined in a first flow mixer after passing through the economizer heat exchanger and first heat exchanger, respectively. The system is further comprised of a second heat exchanger adapted to receive the working fluid from the first flow mixer and a hot heating stream from a heat source and heat the working fluid to a hot vapor via heat transfer. The hot vapor from the second heat exchanger is supplied to at least one turbine or expander after passing through a separator designed to insure no liquid enters said at least one turbine or expander. The hot, high pressure vapor is expanded in the turbine to produce mechanical power on a shaft and is discharged as a hot, low pressure vapor. The hot vapor is then routed back to the economizer heat exchanger and then to a second flow mixer (which may function as a desuperheater) where the hot vapor is mixed with the liquid discharged from the separator and a third portion of the working fluid from the flow divider. The system further comprises a condenser heat exchanger that is adapted to receive the exhaust vapor from the turbine or expander after passing through the economizer heat exchanger and mixing with the liquids from the separator and the flow divider and a cooling fluid circulated by a cooling fluid pump. The system is further comprised of an accumulator vessel to receive the condensed liquid from the condenser and meter said condensate to a liquid working fluid circulating pump that is adapted to circulate the working fluid to a flow divider. The system is finally comprised of a flow divider that is adapted to split the working fluid into at least three portions, at least one that is supplied to an economizer heat exchanger, at least one supplied to a second flow mixer, and at least one that is supplied to a first heat exchanger.
• In yet another illustrative embodiment, the system comprises a first heat exchanger adapted to receive a heating stream from a heat source after passing through a second heat exchanger and a second portion of a working fluid, wherein, when the second portion of the working fluid is passed through the first heat exchanger, the second portion of working fluid is converted to a hot liquid via heat transfer from the heat contained in the heating stream from the heat source after passing through a second heat exchanger. The system is further comprised of an economizer heat exchanger adapted to receive a first portion of the working fluid and the hot discharge vapor from at least one turbine. The first and second portions of the working fluid are recombined in a first flow mixer after passing through the economizer heat exchanger and first heat exchanger, respectively. The system is further comprised of a second heat exchanger adapted to receive the working fluid from the first flow mixer and a hot heating stream from a heat source and convert the working fluid to a hot vapor. The hot vapor from the second heat exchanger is supplied to at least one turbine or expander after passing through a separator designed to insure no liquid enters said at least one turbine or expander. The hot, high pressure vapor is expanded in the turbine or expander to produce mechanical power on a shaft and is discharged as a hot, low pressure vapor. The hot vapor is then routed back to the economizer heat exchanger and then to a second flow mixer where the hot vapor is mixed with the liquid discharged from the separator. The system further comprises a condenser heat exchanger that is adapted to receive the exhaust vapor from the turbine or expander after passing through the economizer heat exchanger and mixing with the liquid from the separator and a gaseous cooling media such as air. The system is further comprised of an accumulator vessel to receive the condensed liquid from the condenser and meter said condensate to a liquid working fluid circulating pump that is adapted to circulate the working fluid to a flow divider. The system is finally comprised of a flow divider that is adapted to split the working fluid into at least two portions, at least one that is supplied to an economizer heat exchanger and at least one that is supplied to a first heat exchanger.
• In a fourth illustrative embodiment, the system comprises a first heat exchanger adapted to receive a heating stream from a heat source after passing through a second heat exchanger and a second portion of a working fluid, wherein, the second portion of the working fluid is passed through the first heat exchanger, the second portion of working fluid is converted to a hot liquid via heat transfer from the heat contained in the heating stream from the heat source after passing through a second heat exchanger. The system is further comprised of an economizer heat exchanger adapted to receive a first portion of the working fluid and the hot discharge vapor from at least one turbine or one expander. The system is further comprised of a second flow mixer adapted to receive a third portion of the working fluid via a fluid bypass control valve. The first and second portions of the working fluid are recombined in a first flow mixer after passing through the economizer heat exchanger and first heat exchanger, respectively. The system is further comprised of a second heat exchanger adapted to receive the working fluid from the first flow mixer and a hot heating stream from a heat source and heat the working fluid to a hot vapor via heat transfer. The hot vapor from the second heat exchanger is supplied to at least one turbine after passing through a separator designed to insure no liquid enters the said at least one turbine or expander. The hot, high pressure vapor is expanded in the turbine or expander to produce mechanical power on a shaft and is discharged as a hot, low pressure vapor. The hot vapor is then routed back to the economizer heat exchanger and then to a second flow mixer where the hot vapor is mixed with the liquid discharged from the separator and a third portion of the working fluid from the flow divider. The system further comprises a condenser heat exchanger that is adapted to receive the exhaust vapor from the turbine or expander after passing through the economizer heat exchanger and mixing with the liquids from the separator and the flow divider and a gaseous cooling media such as air. The system is further comprised of an accumulator vessel to receive the condensed liquid from the condenser and meter said condensate to a liquid working fluid circulating pump that is adapted to circulate the working fluid to a flow divider. The system is finally comprised of a flow divider that is adapted to split the working fluid into at least three portions, at least one that is supplied to an economizer heat exchanger, at least one supplied to a second flow mixer, and at least one that is supplied to a first heat exchanger.
• In all of the illustrative examples, the condenser heat exchanger might be adapted to receive any one or a plurality of cooling fluids such as water from a cooling tower; water from a river or stream; water from a pond, lake, bay, or other freshwater source; seawater from a bay, canal, channel, sea, ocean, or other source; chilled water; fresh air; chilled air; a liquid process stream, e.g. propane; a gaseous process stream, e.g. nitrogen; or other heat sink such as a ground source cooling loop comprised of a plurality of buried pipes.
BRIEF DESCRIPTION OF THE DRAWINGS
• The invention may be understood by reference to the following description taken in conjunction with the accompanying drawings, in which like reference numerals identify like elements, and in which:
• FIG. 1 is a schematic diagram of one illustrative embodiment of the present invention employing a working fluid circulating pump, a flow divider, two heat recovery heat exchangers, an economizer heat exchanger, a first flow mixer, a separator, a turbine or expander, a liquid control valve, a second flow mixer/desuperheater, a liquid cooled condenser heat exchanger, an accumulator, a vent/charge valve, and a cooling liquid circulating pump;
• FIG. 2 is a schematic diagram of one illustrative embodiment of the present invention employing a working fluid circulating pump, a flow divider, two heat recovery heat exchangers, an economizer heat exchanger, a first flow mixer, a separator, a turbine or expander, a liquid control valve, a liquid desuperheater feed bypass flow control valve, a second flow mixer/desuperheater, a liquid cooled condenser heat exchanger, an accumulator, a vent/charge valve, and a cooling liquid circulating pump;
• FIG. 3 is a schematic diagram of one illustrative embodiment of the present invention employing a working fluid circulating pump, a flow divider, two heat recovery heat exchangers, an economizer heat exchanger, a first flow mixer, a separator, a turbine or expander, a liquid control valve, a second flow mixer/desuperheater, a gas cooled condenser heat exchanger, an accumulator, and a vent/charge valve;
• FIG. 4 is a schematic diagram of one illustrative embodiment of the present invention employing a working fluid circulating pump, a flow divider, two heat recovery heat exchangers, an economizer heat exchanger, a first flow mixer, a separator, a turbine or expander, a liquid control valve, a liquid desuperheater feed bypass flow control valve, a second flow mixer/desuperheater, a gas cooled condenser heat exchanger, an accumulator, and a vent/charge valve; and
• FIG. 5 is a schematic diagram of one illustrative embodiment of the prior art employed as an Organic Rankine Cycle with two turbines or expanders and one reheat.
• While the invention is susceptible to various modifications and alternative forms, specific embodiments thereof have been shown by way of example in the drawings and are herein described in detail. It should be understood, however, that the description herein of specific embodiments is not intended to limit the invention to the particular forms disclosed, but on the contrary, the intention is to cover all modifications, equivalents, and alternatives falling within the spirit and scope of the invention as defined by the appended claims.
DETAILED DESCRIPTION OF THE INVENTION
• Illustrative embodiments of the invention are described below. In the interest of clarity, not all features of an actual implementation are described in this specification. It will of course be appreciated that in the development of any such actual embodiment, numerous implementation-specific decisions must be made to achieve the developers' specific goals, such as compliance with system-related and business-related constraints, which will vary from one implementation to another. Moreover, it will be appreciated that such a development effort might be complex and time-consuming, but would nevertheless be a routine undertaking for those of ordinary skill in the art having the benefit of this disclosure.
• The present invention will now be described with reference to the attached drawings which are included to describe and explain illustrative examples of the present invention. The words and phrases used herein should be understood and interpreted to have a meaning consistent with the understanding of those words and phrases by those skilled in the relevant art. No special definition of a term or phrase, i.e., a definition that is different from the ordinary and customary meaning as understood by those skilled in the art, is intended to be implied by consistent usage of the term or phrase herein. To the extent that a term or phrase is intended to have a special meaning, i.e. a meaning other than that understood by skilled artisans, such a special definition will be expressly set forth in the specification in a definitional manner that directly and unequivocally provides the special definition for the term or phrase. Moreover, various streams or conditions may be referred to with terms such as “hot,” “cold,” “cooled, “warm,” etc., or other like terminology. Those skilled in the art will recognize that such terms reflect conditions relative to another process stream, not an absolute measurement of any particular temperature.
• The present invention is generally related to pending allowed U.S. patent application Ser. No. 10/616,074, now U.S. Pat. No. ______. That pending application is hereby incorporated by reference in its entirety.
• One illustrative embodiment of the present invention will now be described with reference toFIG. 1. As shown therein, a high pressure,liquid working fluid2 enters aflow divider26 and is split into twoportions3,10. Afirst portion3 of the working fluid enters aneconomizer heat exchanger27 adapted to receive ahot vapor discharge8 from a turbine orexpander31 and thefirst portion3 of the working fluid is heated via heat transfer with thehot vapor8 and exits as ahot liquid4. For purposes of the present application, the term “turbine” will be understood to include both turbines and expanders or any device wherein useful work is generated by expanding a high pressure gas within the device. Asecond portion10 of the working fluid enters afirst heat exchanger37 that is adapted to receive ahot heating stream20 from a heat source (via line19) after passing through asecond heat exchanger29. Thesecond portion10 of the working fluid is heated via heat transfer with thehot heating stream20 in thefirst heat exchanger37. Thehot heating stream20 discharges from thefirst heat exchanger37 as acool vapor21 that is near or below its dew point. Thesecond portion10 of the working fluid exits the first heat exchanger as ahot liquid11. Thehot liquid4 and thehot liquid11 are mixed in afirst flow mixer28 and discharged as a combined hotliquid stream5. The combined hotliquid stream5 is introduced into asecond heat exchanger29 that is adapted to receive aheating stream19 and exits as asuperheated vapor6 due to heat transfer with a hot fluid, either a gas, a liquid, or a two-phase mixture of gas and liquid entering at19 and exiting at20. Thevapor6 may be a subcritical or supercritical vapor.
• Theheat exchangers27,29, and37 may be any type of heat exchanger capable of transferring heat from one fluid stream to another fluid stream. For example, theheat exchangers27,29, and37 may be shell-and-tube heat exchangers, a plate-fin-tube coil type of exchangers, bare tube or finned tube bundles, welded plate heat exchangers, etc. Thus, the present invention should not be considered as limited to any particular type of heat exchanger unless such limitations are expressly set forth in the appended claims.
• The source of thehot heating stream19 for thesecond heat exchanger29 may either be a waste heat source (from any of a variety of sources) or heat may intentionally be supplied to the system, e.g. by a gas burner, a fuel oil burner, or the like. In one illustrative embodiment, the source of thehot heating stream19 for thesecond heat exchanger29 is a waste heat source such as the exhaust from an internal combustion engine (e.g. a reciprocating diesel engine), a combustion gas turbine, a compressor, or an industrial or manufacturing process. However, any heat source of sufficient quantity and temperature may be utilized if it can be obtained economically. In some cases, the first andsecond heat exchangers37,29 may be referred to either as “waste heat recovery heat exchangers,” indicating that the source of theheating stream19 is from what would otherwise be a waste heat source, although the present invention is not limited to such situations, or “heat recovery heat exchangers” indicating that the source of theheating stream19 is from what would be any heat source.
• In one embodiment, thevapor6 then enters aseparator30 that is designed to protect theturbine31 from any liquid that might be entrained in thevapor6 and to separate the normally dry, highlysuperheated vapor6 into adry vapor7 and aliquid component12. Theliquid component12 is routed away from theseparator30 via aliquid control valve38 to prevent accumulation of the liquid in theseparator30. Thevapor7 then enters the turbine (expander)31. Thevapor7 is expanded in the turbine (expander)31 and the design of theturbine31 converts kinetic and potential energy of thedry vapor7 into mechanical energy in the form of torque on anoutput shaft32. Any type of commercially available turbine suited for use in the systems described herein may be employed, e.g. an expander, a turbo-expander, a power turbine, etc. The shaft horsepower available on theshaft32 of theturbine31 can be used to produce power by driving one or more generators, compressors, pumps, or other mechanical devices, either directly or indirectly. Several illustrative embodiments of how such useful power may be used are described further in the application. Additionally, as will be recognized by those skilled in the art after a complete reading of the present application, a plurality ofturbines31 or heatrecovery heat exchangers29 or37 may be employed with the system depicted inFIG. 1.
• The low pressure,high temperature discharge8 from theturbine31 is routed to aneconomizer heat exchanger27 that is adapted to receive thefirst portion3 of the liquid working fluid. Theeconomizer heat exchanger27 cools thehot vapor8 via heat transfer with thefirst portion3 of the liquid working fluid and discharges the hot vapor as acool vapor9 at or near its dew point. Thecool vapor9 is routed to a second flow mixer ordesuperheater33 that is adapted to receive the cooledvapor9 and a hot incidental fluid13 from theliquid control valve38. The hotincidental fluid13, intermittently discharged during startup, shutdown, or upset conditions may be either a liquid or a vapor containing both a liquid and a gas and would not normally be a gas exclusively. After the combination of the cooledvapor9 and theincidental fluid13 in the second fluid mixer ordesuperheater33 the combinedstream14 is routed to acondenser heat exchanger34 that is adapted to receive a coolingfluid23. Thecondenser34 condenses the slightly superheated to partially wet,low pressure vapor14 and condenses it to the liquid state using water, seawater, or other liquid or boilingfluids23 which might be circulated by a low pressureliquid circulating pump39 which provides the necessary motive force to circulate the cooling fluid frompoint22 topoint24. Thecondenser34 may be utilized to condense the hot working fluid from avapor14 to a liquid15 at a temperature ranging from approximately 50-250° F.
• Thecondensed liquid15 is introduced into anaccumulator drum35. Thedrum35 may serve several purposes, such as, for example: (a) the design of thedrum35 ensures that thepump25 has sufficient head to avoid cavitation; (b) the design of thedrum35 ensures that the supply ofliquid18, 1 to thepump25 is steady; (c) the design of thedrum35 ensures that thepump25 will not be run dry; (d) the design of thedrum35 provides an opportunity to evacuate any non-condensable vapors from the system through avent valve36 vialines16,17; (e) the design of thedrum35 allows for the introduction of process liquid into the system; and (f) the design of thedrum35 allows for the introduction of makeup quantities of the process liquid in the event that a small amount of operating fluid is lost. Thehigh pressure discharge2 of thepump25 is fed to thefirst flow divider26. Thepump25 may be any type of commercially available pump sufficient to meet the pumping requirements of the systems disclosed herein. In various embodiments, thepump25 may be sized such that the discharge pressure of the working fluid ranges from approximately 300 psia to 1500 psia. In the most preferred embodiment, the selection of the discharge pressure of thepump25 is dependent on the critical pressure of the workingfluid2 and should be approximately 5 psia to 500 psia greater than the critical pressure of the workingfluid2 although pressures lower than the critical pressure may be utilized with a reduction in the efficiency of the system.
• In the illustrative embodiment depicted inFIG. 1, the working fluid enters the first heatrecovery heat exchanger37 and theeconomizer heat exchanger27 as a cool, high pressure liquid and, after being recombined, leaves as ahot liquid5. The workingfluid5 then enters the second heatrecovery heat exchanger29 and leaves as asuperheated vapor6. The high pressure,superheated vapor6 is then expanded through aturbine31 to produce mechanical power after passing through aseparator30 and split into adry vapor7 and a liquid12. Thevapor8 exiting theturbine31 is at low pressure and in the superheated state and thevapor8 is passed through theeconomizer heat exchanger27 and thesecond fluid mixer33. In some applications, thesecond fluid mixer33 may function as a desuperheater. After thesecond fluid mixer33, the vapor is then introduced into thecondenser heat exchanger34 which may be water cooled, air cooled, evaporatively cooled, or used as a heat source for district heating, domestic hot water, or similar heating load. The condensedlow pressure liquid15 is fed to the suction of apump25 viadrum35 and is pumped to the high pressure required for the first heatrecovery heat exchanger37 and theeconomizer heat exchanger27.
• The present invention may employ a single component working fluid that may be comprised of, for example, ammonia (NH3), bromine (Br2), carbon tetrachloride (CCl4), ethyl alcohol or ethanol (CH3CH2OH, C2H6O), furan (C4H4O), hexafluorobenzene or perfluorobenzene (C6F6), hydrazine (N2H4), methyl alcohol or methanol (CH3OH), monochlorobenzene or chlorobenzene or chlorobenzol or benzine chloride (C6H5Cl), n-pentane or normal pentane (nC5), i-hexane or isohexane (iC5), pyridene or azabenzene (C5H5N), refrigerant 11 or freon 11 or CFC-11 or R-11 or trichlorofluoromethane (CCl3F), refrigerant 12 or freon 12 or R-12 or dichlorodifluoromethane (CCl2F2), refrigerant 21 or freon 21 or CFC-21 or R-21 (CHCl2F), refrigerant 30 or freon 30 or CFC-30 or R-30 or dichloromethane or methylene chloride or methylene dichloride (CH2Cl2), refrigerant 115 or freon 115 or CFC-115 or R-115 or chloropentafluoroethane or monochloropentafluoroethane, refrigerant 123 or freon 123 or HCFC-123 or R-123 or 2,2 dichloro-1,1,1-trifluoroethane, refrigerant 123a or freon 123a or HCFC-123a or R-123a or 1,2-dichloro-1,1,2-trifluoroethane, refrigerant 123b1 or freon 123b1 or HCFC-123b1 or R-123b1 or halothane or 2-bromo-2-chloro-1,1,1-trifluoroethane, refrigerant 134A or freon 134A or HFC-134A or R-134A or 1,1,1,2-tetrafluoroethane, refrigerant 150A or freon 150A or CFC-150A or R-150A or dichloroethane or ethylene dichloride (CH3CHCl2), thiophene (C4H4S), toluene or methylbenzene or phenylmethane or toluol (C7H8), water (H2O), etc. In some applications, the working fluid may be comprised of multiple components. For example, one or more of the compounds identified above may be combined or with a hydrocarbon fluid, e.g. isobutene, etc. Further, several simple hydrocarbons compounds may be combined such as isopentane, toluene, and hexane to create a working fluid. In the context of the present application, reference may be made to the use of methyl alcohol or methanol as the working fluid and to provide certain illustrative examples. However, after a complete reading of the present application, those skilled in the art will recognize that the present invention is not limited to any particular type of working fluid or refrigerant. Thus, the present invention should not be considered as limited to any particular working fluid unless such limitations are clearly set forth in the appended claims.
• In the present invention, as the workingfluid5 passes through the second heatrecovery heat exchanger29, it changes from a liquid state to a vapor state in a non-isothermal process using an approximately linear temperature-enthalpy profile, i.e., the slope of the temperature-enthalpy curve does not change significantly even though the working fluid changes state from a subcooled liquid to a superheated vapor. The slope of the temperature-enthalpy graph may vary depending upon the application. Moreover, the temperature-enthalpy profile may not be linear over the entire range of the curve.
• The temperature-enthalpy profile of the working fluid of the present invention is fundamentally different from other systems. For example, a temperature-enthalpy profile for a typical Rankine cycle undergoes one or more essentially isothermal (constant temperature) boiling processes as the working fluid changes from a liquid state to a vapor state. Other systems, such as a Kalina cycle, may exhibit a more non-isothermal conversion of the working fluid from a liquid state to a vapor state, but such systems employ binary component working fluids, such as ammonia and water.
• The non-isothermal process used in practicing aspects of the present invention is very beneficial in that it provides a greater heat capacity that may be recaptured when the vapor is cooled back to a liquid. That is, due to the higher temperatures involved in such a non-isothermal process, the working fluid, in the superheated vapor state, contains much more useable heat energy that may be recaptured and used for a variety of purposes. Further, the nearly linear temperature-enthalpy profile allows the exiting temperature of the (waste) heat source to approach more closely to the workingfluid temperature2,10 entering the first heatrecovery heat exchanger37.
• By way of example, with reference toFIG. 1, in one illustrative embodiment where the working fluid is methyl alcohol or methanol, the temperature of the working fluid atpoint2 may be between approximately 50-250° F. at approximately 1120 psia to 1220 psia at the discharge of thepump25. The working fluid atpoint15 may be at a pressure of approximately 1 psia to 92 psia at the discharge of the condenser34 (seeFIG. 1) for a system pressure ratio of between approximately between twelve to one (12:1) and one thousand two hundred and twenty to one (1220:1). In one particularly illustrative embodiment, the pressure ratio would be as large as practical. The temperature of themethanol working fluid6 at the exit of theheat exchanger29 may be approximately 500-1000° F. or more. The temperature of themethanol working fluid8 at the exit of theturbine31 may be between approximately 90° F. (at a pressure of approximately 3 psia) and 670° F. (at a pressure of approximately 92 psia). The temperature of themethanol working fluid8 at the exit of theturbine31 may be superheated by between approximately 10° F. (at a pressure of approximately 8 psia when thevapor7 entering theturbine31 is at 650° F.) and approximately 415° F. (at a pressure of 92 psia when thevapor7 entering theturbine31 is at 1000° F.). The amount of superheat at8 is functionally related to the pressure ratio of the system, the efficiency of theturbine31, the thermodynamic properties of the working fluid, the degree of superheat at7 entering theturbine31, the flow ratio of thestreams3,10 exiting theflow divider26, and the hot heatingstream discharge temperature21. In one particularly illustrative embodiment of the present invention, the temperature of the working fluid atpoint8 exiting theturbine31 will be selected, along with other parameters, to produce acondenser34 inlet temperature as close as possible to the dew point of the workingfluid14 at the conditions entering thecondenser34. The present embodiment will allow large amounts of superheat at7 and at8 and still remain more efficient than previous, related art.
• In another illustrative embodiment where the working fluid is bromine, the temperature of the working fluid atpoint2 may be between approximately 50-250° F. at approximately 1540 psia at the discharge of thepump25. The working fluid atpoint15 may be at a pressure of approximately 11 psia at the discharge of thecondenser34 for a system pressure ratio of approximately one hundred and forty to one (140:1). The temperature of thebromine working fluid6 at the exit of theheat exchanger29 may be approximately 650-1000° F. The temperature of thebromine working fluid8 at the exit of theturbine31 may be approximately 130° F. at a pressure of approximately 13 psia.
• In another illustrative embodiment where the working fluid is carbon tetrachloride, the temperature of the working fluid atpoint2 may be between approximately 50-250° F. at approximately 690 psia at the discharge of thepump25. The working fluid atpoint15 may be at a pressure of approximately 6 psia at the discharge of thecondenser34 for a system pressure ratio of approximately one hundred thirty to one (130:1). The temperature of the carbontetrachloride working fluid6 at the exit of theheat exchanger29 may be approximately 550-770° F. The temperature of the carbontetrachloride working fluid8 at the exit of theturbine31 may be approximately 155-400° F. at a pressure of approximately 8 psia.
• In another illustrative embodiment where the working fluid is ethyl alcohol or ethanol, the temperature of the working fluid atpoint2 may be between approximately 50-250° F. at approximately 1000 psia at the discharge of thepump25. The working fluid atpoint15 may be at a pressure of approximately 4 psia at the discharge of thecondenser34 for a system pressure ratio of approximately two hundred and fifty to one (250:1). The temperature of the ethyl alcohol orethanol working fluid6 at the exit of theheat exchanger29 may be approximately 500-800° F. The temperature of the ethyl alcohol orethanol working fluid8 at the exit of theturbine31 may be approximately 135-400° F. at a pressure of approximately 6 psia.
• In another illustrative embodiment where the working fluid is R-150A, the temperature of the working fluid atpoint2 may be between approximately 50-250° F. at approximately 770 psia at the discharge of thepump25. The working fluid atpoint15 may be at a pressure of approximately 11 psia at the discharge of thecondenser34 for a system pressure ratio of approximately seventy to one (70:1). The temperature of the R-150A working fluid6 at the exit of theheat exchanger29 may be approximately 500-705° F. The temperature of the R-150A working fluid8 at the exit of theturbine31 may be approximately 155-400° F. at a pressure of approximately 13 psia.
• In another illustrative embodiment where the working fluid is thiophene, the temperature of the working fluid atpoint2 may be between approximately 50-250° F. at approximately 900 psia at the discharge of thepump25. The working fluid atpoint15 may be at a pressure of approximately 4.5 psia at the discharge of thecondenser34 for a system pressure ratio of approximately two hundred to one (200:1). The temperature of thethiophene working fluid6 at the exit of theheat exchanger29 may be approximately 600-730° F. The temperature of thethiophene working fluid8 at the exit of theturbine31 may be approximately 220-400° F. at a pressure of approximately 6.5 psia.
• In another illustrative embodiment where the working fluid is a mixture of hydrocarbon compounds, the temperature of the working fluid atpoint2 may be between approximately 50-250° F. at approximately 576 psia at the discharge of thepump25. The working fluid atpoint15 may be at a pressure of approximately 36 psia at the discharge of thecondenser34 for a system pressure ratio of approximately sixteen to one (16:1). The temperature of the mixture of hydrocarboncompounds working fluid6 at the exit of theheat exchanger29 may be approximately 520-655° F. The temperature of the mixture of hydrocarboncompounds working fluid8 at the exit of theturbine31 may be approximately 375-550° F. at a pressure of approximately 38 psia. For this illustrative example, the mixture of hydrocarbons on a molar basis is approximately 10% propane, 10% isobutane, 10% isopentane, 20% hexane, 20% heptane, 10% octane, 10% nonane, and 10% decane. This mixture is one of an infinite number of possible mixtures that might be selected to suit specific needs of a particular embodiment and is in no way representative of the only or best solution.
• The methods and systems described herein may be most effective for pressure ratios greater than three to one (3:1) and the pressure ratio is determined by the physical characteristics of the working fluid being utilized. The specific embodiments of this invention significantly improve the efficiency of the specific embodiments of this invention over the previous inventions of the prior art and of this specific art to allow usage at almost any pressure ratio. The specific selection of the low cycle pressure is determined by the condensing pressure of the working fluid and will be, typically, the saturation pressure of the working fluid at between approximately 0-250° F., depending on the cooling medium or condenser heat exchanger type and the ambient temperature or ultimate heat sink temperature. The specific selection of the high cycle pressure is determined by the thermodynamic properties of the working fluid plus a margin, as a minimum, and by cycle efficiency, pump power consumption, and maximum component design pressures as a maximum.
• In another illustrative embodiment of the present invention a system substantially similar toFIG. 1 will now be described with reference toFIG. 2. As shown therein, a high pressure,liquid working fluid2 enters aflow divider26 and is split into threeportions3,10,40. Afirst portion3 of the working fluid enters theeconomizer heat exchanger27 that is adapted to receive thehot vapor discharge8 from theturbine31 wherein the workingfluid3 is heated via heat transfer with thehot vapor8 and exits as ahot liquid4. Asecond portion10 of the working fluid enters thefirst heat exchanger37 that is adapted to receive thehot heating stream20 from the heat source (via line19) after passing through asecond heat exchanger29, wherein the workingfluid10 is heated to ahot liquid11 via heat transfer with thehot heating stream20, that ultimately discharges from thefirst heat exchanger37 as acool vapor21 near or below its dew point. Athird portion40 of the working fluid is routed to a second fluid mixer33 (which may function as a desuperheater in some cases) that is adapted to receive a portion of the workingfluid42, acool vapor9 from theeconomizer heat exchanger27, and the incidental liquid13 from theseparator30. Thehot liquid4 and thehot liquid11 are mixed in afirst flow mixer28 and discharged as a combined hotliquid stream5. The combined hotliquid stream5 is introduced into thesecond heat exchanger29 that is adapted to receive theheating stream19 and exits as asuperheated vapor6 due to heat transfer with a hot fluid, either a gas, a liquid, or a two-phase mixture of gas and liquid entering at19 and exiting at20. Thevapor6 may be a subcritical or supercritical vapor.
• Theheat exchangers27,29, and37 may be any type of heat exchanger capable of transferring heat from one fluid stream to another fluid stream. For example, theheat exchangers27,29, and37 may be shell-and-tube heat exchangers, a plate-fin-tube coil type of exchangers, bare tube or finned tube bundles, welded plate heat exchangers, etc. Thus, the present invention should not be considered as limited to any particular type of heat exchanger unless such limitations are expressly set forth in the appended claims.
• The source of thehot heating stream19 for thesecond heat exchanger29 may either be a waste heat source (from any of a variety of sources) or heat may intentionally be supplied to the system, e.g. by a gas burner, a fuel oil burner, or the like. In one illustrative embodiment, the source of thehot heating stream19 for thesecond heat exchanger29 is a waste heat source such as the exhaust from an internal combustion engine (e.g. a reciprocating diesel engine), a combustion gas turbine, a compressor, or an industrial or manufacturing process. However, any heat source of sufficient quantity and temperature may be utilized if it can be obtained economically. In some cases, the first andsecond heat exchangers37,29 may be referred to either as “waste heat recovery heat exchangers,” indicating that the source of theheating stream19 is from what would otherwise be a waste heat source, although the present invention is not limited to such situations, or “heat recovery heat exchangers” indicating that the source of theheating stream19 is from what would be any heat source.
• In one embodiment, thevapor6 then enters theseparator30 that is designed to protect theturbine31 from any liquid that might be in thevapor6 and to separate the normally dry, highlysuperheated vapor6 into adry vapor7 and aliquid component12. Theliquid component12 is routed away from theseparator30 via aliquid control valve38 to prevent accumulation of the liquid in theseparator30. Thevapor7 then enters the turbine (expander)31. Thevapor7 is expanded in the turbine (expander)31 and the design of theturbine31 converts kinetic and potential energy of thedry vapor7 into mechanical energy in the form of torque on anoutput shaft32. Any type of commercially available turbine suited for use in the systems described herein may be employed, e.g. an expander, a turbo-expander, a power turbine, etc. The shaft horsepower available on theshaft32 of theturbine31 can be used to produce power by driving one or more generators, compressors, pumps, or other mechanical devices, either directly or indirectly. Several illustrative embodiments of how such useful power may be used are described further in the application. Additionally, as will be recognized by those skilled in the art after a complete reading of the present application, a plurality ofturbines31 or heatrecovery heat exchangers29 or37 may be employed with the system depicted inFIG. 2.
• The low pressure,high temperature discharge8 from theturbine31 is routed to theeconomizer heat exchanger27 that is adapted to receive thefirst portion3 of the liquid working fluid. Theeconomizer heat exchanger27 cools thehot vapor8 via heat transfer with thefirst portion3 of the liquid working fluid and discharges the hot vapor as acool vapor9 at or near its dew point. Thecool vapor9 is routed to a second fluid mixer ordesuperheater33 that is adapted to receive the cooledvapor9, a hot incidental fluid13 from theliquid control valve38, and a portion of the cool, liquid workingfluid42 after the liquid flows through a liquidbypass control valve41 and aline40. The hotincidental fluid13, intermittently discharged during startup, shutdown, or upset conditions may be either a liquid or a vapor containing both a liquid and a gas and would not normally be a gas exclusively. After the combination of the cooledvapor9, theincidental fluid13, and the workingfluid42 in the second fluid mixer ordesuperheater33 the combinedstream14 is routed to acondenser heat exchanger34 that is adapted to receive a coolingfluid23. Thecondenser34 condenses the slightly superheated to partially wet,low pressure vapor14 to the liquid state using water, seawater, or other liquid or boilingfluids23 which might be circulated by a low pressureliquid circulating pump39 which provides the necessary motive force to circulate the cooling fluid frompoint22 topoint24. Thecondenser34 may be utilized to condense the hot working fluid from avapor14 to a liquid15 at a temperature ranging from approximately 0-250° F.
• Thecondensed liquid15 is introduced into anaccumulator drum35. Thedrum35 may serve several purposes, such as, for example: (a) the design of thedrum35 ensures that thepump25 has sufficient head to avoid cavitation; (b) the design of thedrum35 ensures that the supply ofliquid18, 1 to thepump25 is steady; (c) the design of thedrum35 ensures that thepump25 will not be run dry; (d) the design of thedrum35 provides an opportunity to evacuate any non-condensable vapors from the system through avent valve36 vialines16,17; (e) the design of thedrum35 allows for the introduction of process liquid into the system; and (f) the design of thedrum35 allows for the introduction of makeup quantities of process liquid in the event that a small amount of operating fluid is lost. Thehigh pressure discharge2 of thepump25 is fed to thefirst flow divider26. Thepump25 may be any type of commercially available pump sufficient to meet the pumping requirements of the systems disclosed herein. In various embodiments, thepump25 may be sized such that the discharge pressure of the working fluid ranges from approximately 300 psia to 1500 psia. In one particularly illustrative embodiment, the selection of the discharge pressure of thepump25 is dependent on the critical pressure of the workingfluid2 and should be approximately 5 psia to 500 psia greater than the critical pressure of the workingfluid2 although pressures lower than the critical pressure may be utilized with a reduction in the efficiency of the system.
• In the illustrative embodiment depicted inFIG. 2, the working fluid enters the first heatrecovery heat exchanger37 and theeconomizer heat exchanger27 as a cool, high pressure liquid and leaves as ahot liquid5. The workingfluid5 then enters the second heatrecovery heat exchanger29 and leaves as asuperheated vapor6. The high pressure,superheated vapor6 is then expanded through aturbine31 to produce mechanical power after passing through aseparator30 and split into adry vapor7 and a liquid12. Thevapor8 exiting theturbine31 is at low pressure and in the superheated state, and it is passed through theeconomizer heat exchanger27 and thesecond fluid mixer33. After thesecond fluid mixer33, vapor is then introduced into thecondenser heat exchanger34 which may be water cooled, air cooled, evaporatively cooled, or used as a heat source for district heating, domestic hot water, or similar heating load. The condensedlow pressure liquid15 is fed to the suction of apump25 via adrum35 and is pumped to the high pressure required for the first heatrecovery heat exchanger37, theeconomizer heat exchanger27 and theliquid bypass valve41.
• As described above, the present invention may employ a single component working fluid that may be comprised of any of the previously mentioned or similar fluids. After a complete reading of the present application, those skilled in the art will recognize that the present invention is not limited to any particular type of working fluid or refrigerant. Thus, the present invention should not be considered as limited to any particular working fluid unless such limitations are clearly set forth in the appended claims.
• In another illustrative embodiment of the present invention a system substantially similar toFIG. 1 will now be described with reference toFIG. 3. As shown therein, a high pressure,liquid working fluid2 enters theflow divider26 and is split into twoportions3,10. Afirst portion3 of the working fluid enters theeconomizer heat exchanger27 that is adapted to receive ahot vapor discharge8 from theturbine31, wherein the workingfluid3 is heated via heat transfer with thehot vapor8 and exits as ahot liquid4. Asecond portion10 of the working fluid enters thefirst heat exchanger37 that is adapted to receive thehot heating stream20 from a heat source after passing through asecond heat exchanger29, wherein the workingfluid10 is heated to ahot liquid11 via heat transfer with thehot heating stream20, that ultimately discharges as acool vapor21 near or below its dew point. Thehot liquid4 and thehot liquid11 are mixed in thefirst flow mixer28 and discharged as a combined hotliquid stream5. The combined hotliquid stream5 is introduced into thesecond heat exchanger29 that is adapted to receive theheating stream19 and exits as asuperheated vapor6 due to heat transfer with a hot fluid, either a gas, a liquid, or a two-phase mixture of gas and liquid entering at19 and exiting at20. Thevapor6 may be a subcritical or supercritical vapor.
• Theheat exchangers27,29, and37 may be any type of heat exchanger capable of transferring heat from one fluid stream to another fluid stream. For example, theheat exchangers27,29, and37 may be shell-and-tube heat exchangers, a plate-fin-tube coil type of exchangers, bare tube or finned tube bundles, welded plate heat exchangers, etc. Thus, the present invention should not be considered as limited to any particular type of heat exchanger unless such limitations are expressly set forth in the appended claims.
• The source of thehot heating stream19 for thesecond heat exchanger29 may either be a waste heat source (from any of a variety of sources) or heat may intentionally be supplied to the system, e.g. by a gas burner, a fuel oil burner, or the like. In one illustrative embodiment, the source of thehot heating stream19 for thesecond heat exchanger29 is a waste heat source such as the exhaust from an internal combustion engine (e.g. a reciprocating diesel engine), a combustion gas turbine, a compressor, or an industrial or manufacturing process. However, any heat source of sufficient quantity and temperature may be utilized if it can be obtained economically. In some cases, the first andsecond heat exchangers37,29 may be referred to either as “waste heat recovery heat exchangers,” indicating that the source of theheating stream19 is from what would otherwise be a waste heat source, although the present invention is not limited to such situations, or “heat recovery heat exchangers” indicating that the source of theheating stream19 is from what would be any heat source.
• In one embodiment, thevapor6 then enters theseparator30 that is designed to protect theturbine31 from any liquid that might be in thevapor6 and to separate the normally dry, highlysuperheated vapor6 into adry vapor7 and aliquid component12. Theliquid component12 is routed away from theseparator30 via theliquid control valve38 to prevent accumulation of the liquid in theseparator30. Thevapor7 then enters the turbine (expander)31. Thevapor7 is expanded in the turbine (expander)31 and the design of theturbine31 converts kinetic and potential energy of thedry vapor7 into mechanical energy in the form of torque on anoutput shaft32. Any type of commercially available turbine suited for use in the systems described herein may be employed, e.g. an expander, a turbo-expander, a power turbine, etc. The shaft horsepower available on theshaft32 of theturbine31 can be used to produce power by driving one or more generators, compressors, pumps, or other mechanical devices, either directly or indirectly. Several illustrative embodiments of how such useful power may be used are described further in the application. Additionally, as will be recognized by those skilled in the art after a complete reading of the present application, a plurality ofturbines31 or heatrecovery heat exchangers29 or37 may be employed with the system depicted inFIG. 3.
• The low pressure,high temperature discharge8 from theturbine31 is routed to aneconomizer heat exchanger27 adapted to receive afirst portion3 of the liquid working fluid. Theeconomizer heat exchanger27 cools thehot vapor8 via heat transfer with thefirst portion3 of the liquid working fluid and discharges the hot vapor as acool vapor9 at or near its dew point. Thecool vapor9 is routed to a second fluid mixer ordesuperheater33 that is adapted to receive the cooledvapor9 and a hot incidental fluid13 from theliquid control valve38. The hotincidental fluid13, intermittently discharged during startup, shutdown, or upset conditions may be either a liquid or a vapor containing both a liquid and a gas and would not normally be a gas exclusively. After the combination of the cooledvapor9 and theincidental fluid13 in the second fluid mixer ordesuperheater33 the combinedstream14 is routed to acondenser heat exchanger43 adapted to be gas cooled. Thecondenser43 condenses the slightly superheated to partially wet,low pressure vapor14 to the liquid state using air, nitrogen, hydrogen, or other gas. Thecondenser43 may be utilized to condense the hot working fluid from avapor14 to a liquid15 at a temperature ranging from approximately 0-250° F.
• Thecondensed liquid15 is introduced into anaccumulator drum35. Thedrum35 may serve several purposes, such as, for example: (a) the design of thedrum35 ensures that thepump25 has sufficient head to avoid cavitation; (b) the design of thedrum35 ensures that the supply ofliquid18, 1 to thepump25 is steady; (c) the design of thedrum35 ensures that thepump25 will not be run dry; (d) the design of thedrum35 provides an opportunity to evacuate any non-condensable vapors from the system through avent valve36 vialines16,17; (e) the design of thedrum35 allows for the introduction of process liquid into the system; and (f) the design of thedrum35 allows for the introduction of makeup quantities of liquid in the event that a small amount of operating fluid is lost. Thehigh pressure discharge2 of thepump25 is fed to thefirst flow divider26. Thepump25 may be any type of commercially available pump sufficient to meet the pumping requirements of the systems disclosed herein. In various embodiments, thepump25 may be sized such that the discharge pressure of the working fluid ranges from approximately 300 psia to 1500 psia. In the most preferred embodiment, the selection of the discharge pressure of thepump25 is dependent on the critical pressure of the workingfluid2 and should be approximately 5 psia to 500 psia greater than the critical pressure of the workingfluid2 although pressures lower than the critical pressure may be utilized with a reduction in the efficiency of the system.
• In the illustrative embodiment depicted inFIG. 3, the working fluid (3,10) enters the first heatrecovery heat exchanger37 and theeconomizer heat exchanger27 as a cool, high pressure liquid and leaves (after being combined) as ahot liquid5. The workingfluid5 then enters the second heatrecovery heat exchanger29 and leaves as asuperheated vapor6. The high pressure,superheated vapor6 is then expanded through aturbine31 to produce mechanical power after passing through aseparator30 and split into adry vapor7 and a liquid12. Thevapor8 exiting theturbine31 is at low pressure and in the superheated state and is passed through theeconomizer heat exchanger27 and thesecond fluid mixer33. Thereafter, this vapor is then introduced into thecondenser heat exchanger43 that is adapted to be gas cooled. The condensedlow pressure liquid15 is fed to the suction of apump25 via adrum35 and is pumped to the high pressure required for the first heatrecovery heat exchanger37 and theeconomizer heat exchanger27.
• After a complete reading of the present application, those skilled in the art will recognize that the present invention is not limited to any particular type of working fluid or refrigerant. Thus, the present invention should not be considered as limited to any particular working fluid unless such limitations are clearly set forth in the appended claims.
• In another illustrative embodiment of the present invention a system substantially similar toFIG. 2 will now be described with reference toFIG. 4. As shown therein, a high pressure,liquid working fluid2 enters aflow divider26 and is split into threeportions3,10,40. Afirst portion3 of the working fluid enters theeconomizer heat exchanger27 that is adapted to receive ahot vapor discharge8 from aturbine31, wherein the workingfluid3 is heated via heat transfer with thehot vapor8 and exits as ahot liquid4. Asecond portion10 of the working fluid enters thefirst heat exchanger37 that is adapted to receive thehot heating stream20 from theheat source19 after passing through asecond heat exchanger29, wherein the workingfluid10 is heated to a hot liquid via heat transfer with thehot heating stream20, that ultimately discharges from thefirst heat exchanger37 as acool vapor21 near or below its dew point. Athird portion40 of the working fluid is routed to a secondfluid mixer33 that is adapted to receive aportion40 of the workingfluid42, acool vapor9 from theeconomizer heat exchanger27, and an incidental liquid13 from theseparator30. Thehot liquid4 and thehot liquid11 are mixed in afirst flow mixer28 and discharged as a combined hotliquid stream5. The combined hotliquid stream5 is introduced into thesecond heat exchanger29 that is adapted to receive theheating stream19 and exits as asuperheated vapor6 due to heat transfer with a hot fluid, either a gas, a liquid, or a two-phase mixture of gas and liquid entering at19 and exiting at20. Thevapor6 may be a subcritical or supercritical vapor.
• Theheat exchangers27,29, and37 may be any type of heat exchanger capable of transferring heat from one fluid stream to another fluid stream. For example, theheat exchangers27,29, and37 may be shell-and-tube heat exchangers, a plate-fin-tube coil type of exchangers, bare tube or finned tube bundles, welded plate heat exchangers, etc. Thus, the present invention should not be considered as limited to any particular type of heat exchanger unless such limitations are expressly set forth in the appended claims.
• The source of thehot heating stream19 for thesecond heat exchanger29 may either be a waste heat source (from any of a variety of sources) or heat may intentionally be supplied to the system, e.g. by a gas burner, a fuel oil burner, or the like. In one illustrative embodiment, the source of thehot heating stream19 for thesecond heat exchanger29 is a waste heat source such as the exhaust from an internal combustion engine (e.g. a reciprocating diesel engine), a combustion gas turbine, a compressor, or an industrial or manufacturing process. However, any heat source of sufficient quantity and temperature may be utilized if it can be obtained economically. In some cases, the first andsecond heat exchangers37,29 may be referred to either as “waste heat recovery heat exchangers,” indicating that the source of theheating stream19 is from what would otherwise be a waste heat source, although the present invention is not limited to such situations, or “heat recovery heat exchangers” indicating that the source of theheating stream19 is from what would be any heat source.
• In one embodiment, thevapor6 then enters theseparator30 that is designed to protect theturbine31 from any liquid that might be in thevapor6 and to separate the normally dry, highlysuperheated vapor6 into adry vapor7 and aliquid component12. Theliquid component12 is routed away from the separator via aliquid control valve38 to prevent accumulation of the liquid in theseparator30. Thevapor7 then enters the turbine (expander)31. Thevapor7 is expanded in the turbine (expander)31 and the design of theturbine31 converts kinetic and potential energy of thedry vapor7 into mechanical energy in the form of torque on anoutput shaft32. Any type of commercially available turbine suited for use in the systems described herein may be employed, e.g. an expander, a turbo-expander, a power turbine, etc. The shaft horsepower available on theshaft32 of theturbine31 can be used to produce power by driving one or more generators, compressors, pumps, or other mechanical devices, either directly or indirectly. Several illustrative embodiments of how such useful power may be used are described further in the application. Additionally, as will be recognized by those skilled in the art after a complete reading of the present application, a plurality ofturbines31 or heatrecovery heat exchangers29 or37 may be employed with the system depicted inFIG. 4.
• The low pressure,high temperature discharge8 from theturbine31 is routed to aneconomizer heat exchanger27 that is adapted to receive thefirst portion3 of the liquid working fluid. Theeconomizer heat exchanger27 cools thehot vapor8 via heat transfer with thefirst portion3 of the liquid working fluid and discharges the hot vapor as acool vapor9 at or near its dew point. Thecool vapor9 is routed to a secondfluid mixer33 that is adapted to receive the cooledvapor9, a hot incidental fluid13 from theliquid control valve38, and a portion of the cool, liquid workingfluid42 after the liquid flows through a liquidbypass control valve41 and aline40. The hotincidental fluid13, intermittently discharged during startup, shutdown, or upset conditions may be either a liquid or a vapor containing both a liquid and a gas and would not normally be a gas exclusively. After the combination of the cooledvapor9, theincidental fluid13, and the workingfluid42 in the second fluid mixer ordesuperheater33 the combinedstream14 is routed to acondenser heat exchanger43 that is adapted to be gas cooled. Thecondenser43 condenses the slightly superheated to partially wet,low pressure vapor14 and condenses it to the liquid state using air, nitrogen, hydrogen, or other gas. Thecondenser43 may be utilized to condense the hot working fluid from avapor14 to a liquid15 at a temperature ranging from approximately 0-250° F.
• Thecondensed liquid15 is introduced into anaccumulator drum35. Thedrum35 may serve several purposes, such as, for example: (a) the design of thedrum35 ensures that thepump25 has sufficient head to avoid cavitation; (b) the design of thedrum35 ensures that the supply ofliquid18, 1 to thepump25 is steady; (c) the design of thedrum35 ensures that thepump25 will not be run dry; (d) the design of thedrum35 provides an opportunity to evacuate any non-condensable vapors from the system through avent valve36 vialines16,17; (e) the design of thedrum35 allows for the introduction of process liquid into the system; and (f) the design of thedrum35 allows for the introduction of makeup quantities of process liquid in the event that a small amount of operating fluid is lost. Thehigh pressure discharge2 of thepump25 is fed to thefirst flow divider26. Thepump25 may be any type of commercially available pump sufficient to meet the pumping requirements of the systems disclosed herein. In various embodiments, thepump25 may be sized such that the discharge pressure of the working fluid ranges from approximately 300 psia to 1500 psia. In one particularly illustrative embodiment, the selection of the discharge pressure of thepump25 is dependent on the critical pressure of the workingfluid2 and should be approximately 5 psia to 500 psia greater than the critical pressure of the workingfluid2 although pressures lower than the critical pressure may be utilized with a reduction in the efficiency of the system.
• In the illustrative embodiment depicted inFIG. 4, the working fluid enters the first heatrecovery heat exchanger37 and theeconomizer heat exchanger27 as a cool, high pressure liquid and leaves as ahot liquid5. The workingfluid5 then enters the second heatrecovery heat exchanger29 and leaves as asuperheated vapor6. The high pressure,superheated vapor6 is then expanded through aturbine31 to produce mechanical power after passing through aseparator30 and split into adry vapor7 and a liquid12. Thevapor8 exiting theturbine31 is at low pressure and in the superheated state and it is passed through theeconomizer heat exchanger27 and thesecond fluid mixer33. After thesecond fluid mixer33, this vapor is then introduced into thecondenser heat exchanger43. The condensedlow pressure liquid15 is fed to the suction of apump25 via adrum35 and is pumped to the high pressure required for the first heatrecovery heat exchanger37, theeconomizer heat exchanger27 and theliquid bypass valve41.
• As described above, the present invention may employ a single component working fluid that may be comprised of any of the previously mentioned or similar fluids. After a complete reading of the present application, those skilled in the art will recognize that the present invention is not limited to any particular type of working fluid or refrigerant. Thus, the present invention should not be considered as limited to any particular working fluid unless such limitations are clearly set forth in the appended claims.
• In one specific embodiment of the present invention, the mechanical power available at the output shaft of the turbine may be utilized directly or through a gearbox to provide mechanical work to drive an electrical power generator to produce electrical power either as a constant voltage and constant frequency AC source or as a DC source which might be rectified to produce AC power at a constant voltage and constant frequency.
• In another specific embodiment, the mechanical power available at the output shaft of the turbine may be utilized directly or through a gearbox to provide mechanical work to drive any combination of mechanical devices such as a compressor, a pump, a wheel, a propeller, a conveyer, a fan, a gear, or any other mechanical device(s) requiring or accepting mechanical power input. Moreover, the present invention is not restricted to stationary devices, as it may be utilized in or on an automobile, a ship, an aircraft, a spacecraft, a train, or other non-stationary vessel.
• A specific byproduct of the method of the present invention is an effective and dramatic reduction in the emissions of both pollutants and greenhouse gases. This method may not require any fuel nor does it generate any pollutants or greenhouse gases or any other gases as byproducts. Any process to which this method may be applied, such as a gas turbine or a diesel engine, will generate significantly more power with no increase in fuel consumption or pollution. The effect of this method is a net reduction in the specific pollution generation rate on a mass per power produced basis.
• The present invention is generally directed to various systems and methods for producing mechanical power from a heat source. In various illustrative examples, the devices employed in practicing the present invention may include heat recovery heat exchangers, turbines or expanders, an economizer heat exchanger, a desuperheater heat exchanger, a condenser heat exchanger, an accumulator, a separator, and a liquid circulating pump, etc. In one illustrative embodiment, the system comprises heat exchangers adapted to receive a fluid from a heat source and a working fluid, wherein, when the working fluid is passed through the first heat exchanger, the working fluid is converted to a vapor via heat transfer from the heat contained in the fluid from the heat source, at least one turbine adapted to receive the vapor, and an economizer heat exchanger adapted to receive exhaust vapor from the turbine and a portion of the working fluid, wherein a temperature of the working fluid is adapted to be increased via heat transfer with the exhaust vapor from the turbine prior to the introduction of the working fluid into the second heat exchangers. The system further comprises a condenser heat exchanger that is adapted to receive the exhaust vapor from the turbine after the exhaust vapor has passed through the economizer heat exchanger and a cooling fluid, wherein a temperature of the exhaust vapor is reduced via heat transfer with the cooling fluid, and a pump that is adapted to circulate the working fluid to the first and second heat exchanger and the economizer heat exchanger.
• The particular embodiments disclosed above are illustrative only, as the invention may be modified and practiced in different but equivalent manners apparent to those skilled in the art having the benefit of the teachings herein. For example, the process steps set forth above may be performed in a different order. Furthermore, no limitations are intended to the details of construction or design herein shown, other than as described in the claims below. It is therefore evident that the particular embodiments disclosed above may be altered or modified and all such variations are considered within the scope and spirit of the invention. Accordingly, the protection sought herein is as set forth in the claims below.

Claims (36)

1. A system, comprising:

a flow divider adapted to receive a working fluid and divide said working fluid into at least first and second portions;

a first heat exchanger adapted to receive a hot heating stream from a heat source after said hot heating stream passes through a second heat exchanger and said first portion of a working fluid, wherein, when the first portion of said working fluid is passed through the first heat exchanger, the first portion of said working fluid is converted to a first hot liquid stream via heat transfer with said hot heating stream from said heat source;

an economizer heat exchanger adapted to receive a hot vapor discharged from at least one turbine and said second portion of said working fluid, wherein, when the second portion of said working fluid is passed through the economizer heat exchanger, the second portion of said working fluid is converted to a second hot liquid stream via heat transfer with said hot vapor discharged from said at least one turbine;

a first flow mixer adapted to receive at least said first hot liquid stream and said second hot liquid stream and discharge said at least first hot liquid stream and said second hot liquid stream as a combined hot liquid working fluid;

a second heat exchanger adapted to receive a hot heating stream from a heat source and said combined hot liquid working fluid, wherein, when the combined hot liquid working fluid is passed through the second heat exchanger, the combined hot liquid working fluid is converted to a hot vapor via heat transfer with said hot heating stream from said heat source;

at least one separator adapted to receive said hot vapor from said second heat exchanger and separate said hot vapor into its liquid and gaseous phases for the purpose of preventing liquid from entering said at least one turbine;

at least one liquid control valve to relieve said liquid from said at least one separator, wherein said at least one turbine is adapted to receive said hot vapor from said separator and produce rotational, mechanical power at a shaft that is adapted to transmit said power to at least one device adapted to receive said power;

a second flow mixer that is adapted to receive and combine said vapor discharged from said at least one turbine after said vapor passes through said economizer heat exchanger and a fluid from the discharge of said liquid control valve into a single stream to be condensed;

a condenser heat exchanger that is adapted to receive said single stream to be condensed and a cooling fluid, wherein the temperature of said single stream to be condensed is reduced via heat transfer with said cooling fluid;

a liquid accumulator that is adapted to receive said cooled working fluid, provide storage for said cooled working fluid, and provide a surge volume for said system; and,

at least one pump that is adapted to circulate said cooled working fluid to said first heat exchanger and said economizer heat exchanger via said flow divider.

2. The system ofclaim 1 wherein said working fluid enters said second heat exchanger as a supercritical liquid and via heat transfer with the fluid from the heat source changes state from a supercritical liquid to a supercritical vapor.

3. The system ofclaim 1 wherein said cooling fluid for said condenser heat exchanger comprises at least one of a liquid and a gas.

4. The system ofclaim 1 wherein said cooling fluid for said condenser heat exchanger is a partially or fully vaporized liquid as it passes through said condenser heat exchanger.

5. The system ofclaim 1 wherein said condenser heat exchanger is adapted to condense the exhaust vapor from said at least one turbine or expander to a liquid at a temperature between approximately 0-250° F.

6. The system ofclaim 1, wherein said working fluid is R-123 or one of its derivatives and said fluid from said heat source has a temperature of between approximately 450-1500° F., the maximum temperature of the working fluid is between approximately 363-700° F., and wherein said pump is adapted to operate at a discharge pressure greater than approximately 550 psia.

7. The system ofclaim 1, wherein said working fluid is R-134A or one of its derivatives and said fluid from said heat source has a temperature of between approximately 275-1500° F., the maximum temperature of the working fluid is between approximately 214-650° F., and wherein said pump is adapted to operate at a discharge pressure greater than approximately 600 psia.

8. The system ofclaim 1, wherein said working fluid is methyl alcohol (methanol) or one of its derivatives and said fluid from said heat source has a temperature of between approximately 500-2500° F., the maximum temperature of the working fluid is between approximately 463-963° F., and wherein said pump is adapted to operate at a discharge pressure greater than approximately 1070 psia.

9. The system ofclaim 1, wherein said working fluid is bromine and said fluid from said heat source has a temperature of between approximately 500-2500° F., the maximum temperature of the working fluid is between approximately 592-1092° F., and wherein said pump is adapted to operate at a discharge pressure greater than approximately 1500 psia.

10. The system ofclaim 1, wherein said working fluid is carbon tetrachloride and said fluid from said heat source has a temperature of between approximately 600-2500° F., the maximum temperature of the working fluid is between approximately 542-1042° F., and wherein said pump is adapted to operate at a discharge pressure greater than approximately 1000 psia.

11. The system ofclaim 1, wherein said working fluid is ethyl alcohol or one of its derivatives and said fluid from said heat source has a temperature of between approximately 500-2500° F., the maximum temperature of the working fluid is between approximately 470-970° F., and wherein said pump is adapted to operate at a discharge pressure greater than approximately 920 psia.

12. The system ofclaim 1, wherein said working fluid is R-150A and said fluid from said heat source has a temperature of between approximately 500-2500° F., the maximum temperature of the working fluid is between approximately 482-982° F., and wherein said pump is adapted to operate at a discharge pressure greater than approximately 730 psia.

13. The system ofclaim 1, wherein said working fluid is thiophene and said fluid from said heat source has a temperature of between approximately 600-2500° F., the maximum temperature of the working fluid is between approximately 583-1083° F., and wherein said pump is adapted to operate at a discharge pressure greater than approximately 730 psia.

14. The system ofclaim 1, wherein said working fluid is a mixture of hydrocarbons containing ten or fewer carbon atoms per molecule, said fluid from said heat source has a temperature of between approximately 400-2500° F., the maximum temperature of the working fluid is between approximately 400-1000° F., and wherein said pump is adapted to operate at a discharge pressure greater than approximately 300 psia.

15. The system ofclaim 1, wherein said at least one turbine drives at least one electrical generator to produce electrical power.

16. The system ofclaim 1, wherein said at least one turbine drives at least one compressor.

17. The system ofclaim 1, wherein said at least one turbine drives said at least one pump.

18. The system ofclaim 1, wherein said at least one turbine drives at least one electrical generator to produce electrical power and drives said at least one pump.

19. A system, comprising:

a flow divider adapted to receive a working fluid and divide said working fluid into at least three portions;

a first heat exchanger adapted to receive a hot heating stream from a heat source after said hot heating stream passes through a second heat exchanger and a first portion of a working fluid, wherein, when the first portion of said working fluid is passed through the first heat exchanger, the first portion of said working fluid is converted to a first hot liquid stream via heat transfer with said heating stream from said heat source;

an economizer heat exchanger adapted to receive a hot vapor discharged from at least one turbine and a second portion of said working fluid, wherein, when the second portion of said working fluid is passed through the economizer heat exchanger, the second portion of said working fluid is converted to a second hot liquid stream via heat transfer from the heat contained in said hot vapor discharged from at least one turbine;

a bypass desuperheater liquid control valve to regulate the flow of a third portion of said working fluid to a second flow mixer or desuperheater;

a first flow mixer adapted to receive said first hot liquid stream and said second hot liquid stream and discharge said first and second hot liquid streams as a combined hot liquid working fluid stream;

said second heat exchanger being adapted to receive said hot heating stream from said heat source and said combined hot liquid working fluid, wherein, when the combined hot liquid working fluid is passed through the second heat exchanger, the combined hot liquid working fluid is converted to a vapor via heat transfer with said hot heating stream from said heat source;

at least one separator adapted to receive said hot vapor from said second heat exchanger and separate said hot vapor into its liquid and gaseous phases for the purpose of preventing liquid from entering said at least one turbine;

at least one liquid control valve to relieve said liquid from said at least one separator, wherein said at least one turbine is adapted to receive said hot vapor and produce rotational, mechanical power at a shaft that is adapted to transmit said power to at least one device adapted to receive said power;

a second flow mixer that is adapted to receive and combine said vapor discharged from said at least one turbine after said vapor passes through said economizer heat exchanger, a fluid from the discharge of said liquid control valve, and said third portion of the working fluid from said bypass desuperheater liquid control valve into a single stream to be condensed;

a condenser heat exchanger that is adapted to receive said single stream to be condensed and a cooling fluid, wherein the temperature of said single stream to be condensed is reduced via heat transfer with said cooling fluid;

a liquid accumulator that is adapted to receive said cooled working fluid, provide storage for said cooled working fluid, and provide a surge volume for said system; and,

at least one pump that is adapted to circulate said cooled working fluid to said first heat exchanger, said economizer heat exchanger, and said second fluid mixer or desuperheater via said bypass desuperheater liquid control valve via said flow divider.

20. The system ofclaim 19, wherein said working fluid enters said second heat exchanger as a supercritical liquid and via heat transfer with the fluid from the heat source changes state from a supercritical liquid to a supercritical vapor.

21. The system ofclaim 19, wherein said cooling fluid for said condenser heat exchanger comprises at least one of a liquid and a gas.

22. The system ofclaim 19, wherein said cooling fluid for said condenser heat exchanger is a partially or fully vaporized liquid as it passes through said condenser heat exchanger.

23. The system ofclaim 19, wherein said condenser heat exchanger is adapted to condense the exhaust vapor from said at least one turbine or expander to a liquid at a temperature between approximately 0-250° F.

24. The system ofclaim 19, wherein said working fluid is R-123 or one of its derivatives and said fluid from said heat source has a temperature of between approximately 450-1500° F., the maximum temperature of the working fluid is between approximately 363-700° F., and wherein said pump is adapted to operate at a discharge pressure greater than approximately 550 psia.

25. The system ofclaim 19, wherein said working fluid is R-134A or one of its derivatives and said fluid from said heat source has a temperature of between approximately 275-1500° F., the maximum temperature of the working fluid is between approximately 214-650° F., and wherein said pump is adapted to operate at a discharge pressure greater than approximately 600 psia.

26. The system ofclaim 19, wherein said working fluid is methyl alcohol (methanol) or one of its derivatives and said fluid from said heat source has a temperature of between approximately 500-2500° F., the maximum temperature of the working fluid is between approximately 463-963° F., and wherein said pump is adapted to operate at a discharge pressure greater than approximately 1070 psia.

27. The system ofclaim 19, wherein said working fluid is bromine and said fluid from said heat source has a temperature of between approximately 500-2500° F., the maximum temperature of the working fluid is between approximately 592-1092° F., and wherein said pump is adapted to operate at a discharge pressure greater than approximately 1500 psia.

28. The system ofclaim 19, wherein said working fluid is carbon tetrachloride and said fluid from said heat source has a temperature of between approximately 600-2500° F., the maximum temperature of the working fluid is between approximately 542-1042° F., and wherein said pump is adapted to operate at a discharge pressure greater than approximately 1000 psia.

29. The system ofclaim 19, wherein said working fluid is ethyl alcohol or one of its derivatives and said fluid from said heat source has a temperature of between approximately 500-2500° F., the maximum temperature of the working fluid is between approximately 470-970° F., and wherein said pump is adapted to operate at a discharge pressure greater than approximately 920 psia.

30. The system ofclaim 19, wherein said working fluid is R-150A and said fluid from said heat source has a temperature of between approximately 500-2500° F., the maximum temperature of the working fluid is between approximately 482-982° F., and wherein said pump is adapted to operate at a discharge pressure greater than approximately 730 psia.

31. The system ofclaim 19, wherein said working fluid is thiophene and said fluid from said heat source has a temperature of between approximately 600-2500° F., the maximum temperature of the working fluid is between approximately 583-1083° F., and wherein said pump is adapted to operate at a discharge pressure greater than approximately 730 psia.

32. The system ofclaim 19, wherein said working fluid is a mixture of hydrocarbons containing ten or fewer carbon atoms per molecule, said fluid from said heat source has a temperature of between approximately 400-2500° F., the maximum temperature of the working fluid is between approximately 400-1000° F., and wherein said pump is adapted to operate at a discharge pressure greater than approximately 300 psia.

33. The system ofclaim 19, wherein said at least one turbine drives at least one electrical generator to produce electrical power.

34. The system ofclaim 19, wherein said at least one turbine drives at least one compressor.

35. The system ofclaim 19, wherein said at least one turbine drives said at least one pump.

36. The system ofclaim 19, wherein said at least one turbine drives at least one electrical generator to produce electrical power and drives said at least one pump.

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