Multiple organic rankine cycle systems and methods

Abstract

Systems and methods are provided for the recovery mechanical power from heat energy sources using a common working fluid comprising, in some embodiments, an organic refrigerant flowing through multiple heat exchangers and expanders. The distribution of heat energy from the source may be portioned, distributed, and communicated to each of the heat exchangers so as to permit utilization of up to all available heat energy. In some embodiments, the system utilizes up to and including all of the available heat energy from the source. The expanders may be operatively coupled to one or more generators that convert the mechanical energy of the expansion process into electrical energy, or the mechanical energy may be communicated to other devices to perform work.

Claims

1. A system for generating power from heat, the system comprising: A. a source of heat energy; B. a working fluid, a working fluid condenser, and one or more working fluid pump(s) in working fluid receiving communication with the condenser; C. a first heat exchanger in working fluid receiving communication with at least one of the one or more working fluid pump(s); D. a second heat exchanger in working fluid receiving communication with at least one of the one or more working fluid pump(s); E. a first heat energy flow control valve in heat energy receiving communication with the source of heat energy and in heat energy sending communication with the first heat exchanger, said flow control valve operative to portion, distribute, and communicate a first portion of heat energy from the source of heat energy to the first heat exchanger; F. a second heat energy flow control valve in heat energy receiving communication with the source of heat energy and in heat energy sending communication with the second heat exchanger, said flow control valve operative to portion, distribute, and communicate a second portion of heat energy from the source of heat energy to the second heat exchanger; G. a first expander in working fluid receiving communication with the first heat exchanger and in working fluid sending communication with the condenser; H. a second expander in working fluid receiving communication with the second heat exchanger and in working fluid sending communication with the condenser; wherein: i. at least one of the one or more working fluid pump(s) is configured to provide sufficient motive force to establish and maintain a flow of a first portion of working fluid from the condenser through the first heat exchanger, and then through the first expander, and then back to the condenser; ii. at least one of the one or more working fluid pump(s) is configured to provide sufficient motive force to establish and maintain a flow of a second portion of working fluid from the condenser through the second heat exchanger, then through the second expander, and then back to the condenser; and iii. the system is configured to allow the first portion and second portion of working fluid to be heated during passage through the first heat exchanger and the second heat exchanger, respectively, and to expand during passage through the first expander and the second expander, respectively, thereby generating mechanical output power at the first expander and the second expander, respectively.

2. The system of claim 1 wherein the first expander and the second expander are mechanically independent.

3. The system of claim 1 wherein the mechanical output power generated by the first expander is separately generated from the mechanical output power generated by the second expander.

4. The system of claim 1 further configured to communicate the mechanical output power generated by the first expander, the mechanical output power generated by the second expander, or the mechanical output power generated by the first expander and the second expander to at least one of any of an electric power generator, a prime mover, a pump, a combustion engine, a fan, a turbine, or a compressor.

5. The system of claim 1 wherein the first portion of heat energy and the second portion of heat energy comprise in combination up to and including all of the heat energy available from the source of heat energy.

6. The system of claim 1 wherein the source of heat energy is jacket cooling fluid from an internal combustion engine.

7. The system of claim 6 wherein the first portion of heat energy and the second portion of heat energy comprise in combination up to and including all of the heat energy available from the source of heat energy.

8. The system of claiml further comprising a working fluid receiver disposed between the condenser and the one or more working fluid pumps, and wherein the system is additionally configured to establish and maintain a flow of the first portion and the second portion of working fluid from the condenser to the first heat exchanger and from the condenser to the second heat exchanger, respectively, via the working fluid receiver.

9. The system of claim 1 further comprising a working fluid separator disposed between the second expander and the condenser, and wherein the system is additionally configured to establish and maintain a flow of working fluid from the second expander to the condenser via the working fluid separator.

10. The system of claim 1 further configured such that the first expander is in working fluid sending communication with the second expander and that at least one of the one or more working fluid pumps is configured to provide sufficient motive force to establish and maintain a flow of working fluid from the first expander to the condenser via the second expander.

11. The system of claim 10 further comprising a working fluid separator disposed between the first expander and the second expander, and that the system is further configured to establish and maintain a flow of working fluid from the first expander to the second expander via the working fluid separator.

12. A system for generating power from heat, the system comprising: A. a first portion and a second portion of working fluid; B. a first heat exchanger and a second heat exchanger; C. a first expander and a second expander; D. a source of heat energy (i) in controllable heat energy sending communication with said first heat exchanger and in heat transfer sending communication with said first portion of working fluid passing through the first heat exchanger, and (ii) in controllable heat energy sending communication with said second heat exchanger and in heat transfer sending communication with said second portion of working fluid passing through the second heat exchanger; and E. first and second heat energy flow control valves disposed between the source of heat energy and the first and second heat exchangers, respectively, said flow control valves operative to provide the requisite amount of heat energy from the source of heat energy to each of the first and second heat exchangers; wherein the system is configured to: i. communicate the first portion of working fluid from the first heat exchanger to the first expander and allow said first portion of working fluid to expand in the first expander, thereby generating mechanical output power; ii. communicate the second portion of working fluid from the second heat exchanger to the second expander and allow said second portion of working fluid to expand in the second expander, thereby generating mechanical output power; and iii. communicate the mechanical output power generated by the first expander, the second expander, or the first expander and the second expander to at least one of any of an electric power generator, a prime mover, a pump, a combustion engine, a fan, a turbine, or a compressor.

13. The system of claim 12 wherein the sum of the heat energy communicated to the first and second heat exchangers is up to and including all of the available heat energy available from the source of heat energy.

14. The system of claim 12 wherein the sum of the heat energy communicated to the first and second heat exchangers is up to and including all of the available heat energy available from the source of heat energy.

15. The system of claim 12 further configured to communicate the first portion of working fluid from the first expander to the second expander, communicate the second portion of working fluid from the second heat exchanger to the second expander, combine said first portion of working fluid with said second portion of working fluid at the second expander, and expand the combined first and second portions of working fluid in the second expander.

16. A system for generating power from heat, the system comprising: A. more than one working fluid heat exchanger and more than one expander, said heat exchangers and expanders being equal in number; B. a source of heat energy in heat transfer communication with each of the more than one working fluid heat exchangers; C. more than one heat energy flow control valve, at least one of said more than one valves disposed between the source of heat energy and each of the more than one working fluid heat exchangers, said flow control valves being operative to provide the requisite amount of heat energy from the source of heat energy to each of the more than one portions of heated working fluid via the one of the more than one working fluid heat exchangers exclusively associated with each portion of working fluid; D. a working fluid comprising more than one portion of said working fluid, the number of said portions being equal to the number of the more than one working fluid heat exchangers and the number of the more than one expanders, where each portion of working fluid is exclusively associated with one of the more than one heat exchanger; wherein the system is configured to: i. communicate a controllable, predetermined amount of heat energy from the source of heat energy to each said portion of working fluid via one of the more than one working fluid heat exchangers to create more than one portion of heated working fluid; ii. expand each of said more than one portions of heated working fluid in each of one of the more than one expanders, thereby generating mechanical output power; and iii. communicate the mechanical output power generated by at least one of the more than one expanders to at least one of any of an electric power generator, a prime mover, a pump, a combustion engine, a fan, a turbine, or a compressor.

17. The system of claim 16 further configured such that up to and including all of the available heat energy available from the source of heat energy is communicated in combination to the more than one portions of the working fluid via the more than one working fluid heat exchangers.

18. The system of claim 16 further configured such that at least one portion of the working fluid is communicated from at least one of the more than one expanders to at least one other expander of the more than one expanders and combined with another portion of the working fluid prior to expansion in said other expander.

19. The system of claim 18 further configured such that up to and including all of the heat energy available from the source of heat energy is communicated in combination to the more than one portions of the working fluid via the more than one working fluid heat exchangers.

20. The system of claim 16 wherein up to and including all of the heat energy available from the source of heat energy is communicated in combination to the more than one portions of the working fluid via the more than one working fluid heat exchangers.

Description

BRIEF DESCRIPTION OF THE DRAWINGS

(1) Without limiting the invention to the features and embodiments depicted, certain aspects this disclosure, including the preferred embodiment, are described in association with the appended figures in which:

(2) FIG. 1 is a block diagram of a prior art ORC system used to convert waste heat energy into electric power;

(3) FIG. 2 is a block diagram of an embodiment of a 2P multi-pressure ORC system with two expanders;

(4) FIG. 3 is a flow chart describing the method in one embodiment of determining the operating parameters for a 2P ORC system;

(5) FIG. 4 depicts the temperature versus heat energy of the source and a hypothetical working fluid during the heat energy transfer process from the source to the ORC working fluid in the low pressure cycle of a 2P multi-pressure ORC system;

(6) FIG. 5 is a block diagram of a prior art ORC system used to convert waste heat energy into electric power including heat extraction ports that can be used to provide heat for other applications; and

(7) FIG. 6 is a block diagram of the energy flow in a prior art system including a prime mover, an ORC system used to convert waste heat energy into electric power, and heat extraction ports for other non-system applications.

DETAILED DESCRIPTION OF THE PREFERRED AND OTHER EMBODIMENTS

(8) FIG. 2 depicts a multi-pressure ORC system 200 that utilizes two expanders 224, 242 operating at different pressures. This configuration is an embodiment of a dual-pressure or 2P ORC system.

(9) By way of example and not limitation, this embodiment as described is suitable for use with a J316 ICE engine, as specified and manufactured by the Jenbacher Gas Engine division of General Electric Energy, as the prime mover. Those skilled in the art will recognize that different configurations suitable for other applications are clearly envisioned by this invention, such as the use of prime movers including but not limited to ICEs with power outputs ranging from 250 kW to 8,000 kW. In this embodiment, the J316 serves a single prime mover for the 2P ORC system and supplies heat energy from both exhaust gas flow and jacket cooling water.

(10) Heat energy contained in the exhaust gas flow of the prime mover is supplied at 201 to a thermal oil heat transfer subsystem 203 operatively coupled to first high pressure cycle evaporator 205 via a recirculating flow of oil through conduits 204 and 206. Thermal oil heat transfer subsystem 203 may include an exhaust gas heat exchanger such as those manufactured and sold by E.J. Bowman Ltd. of Birmingham, UK. The oil flow through this intermediate heat transfer system is facilitated by a pump 207. Following extraction of up to all of the useful heat energy from the exhaust gas flow, at least to the degree of a desired working fluid temperature increase through the first high pressure cycle evaporator 205, the reduced temperature exhaust gas exits the thermal oil heater subsystem at 202. The first high pressure cycle evaporator 205 may be a brazed plate heat exchanger such as those supplied by GEA Heat Exchangers GmbH of Bochum. Germany.

(11) In this particular embodiment, the temperature of the exhaust gas at 201 is approximately 950° F. and approximately 350° F. at 202. Extracting additional heat energy from the exhaust gas flow would further reduce the temperature at 202, resulting in the condensation and precipitation of certain corrosive agents from the exhaust gas flow that would damage and adversely affect the performance of the system. So-called “bad actor” corrosive agents include residual and largely non-combustible elements and compounds present in the fuel supplied to the prime mover ICE, particularly those found in biogas produced by decomposition of unknown biological and/or other materials. Sulfur is one particularly notorious bad actor, as it may combine to form hydrogen sulfide gas (H.sub.2S) or sulfuric acid (H.sub.2SO.sub.4). Both are extremely corrosive and toxic and, if allowed to precipitate within the exhaust gas heat exchanger portion of thermal oil heat transfer subsystem 203, would significantly degrade the performance and reduce the operating life of that subsystem. For optimum system performance, it is desirable that these bad actors remain in the vapor state until expelled from the system's exhaust stack.

(12) In one embodiment, the working fluid may be heated by any different form of intermediate heat transfer system. In one embodiment, the working fluid may be heated directly by the exhaust gas without the use of an intermediate heat transfer system such as thermal oil heat transfer subsystem 203. For example, the working fluid may be directed through conduits and manifolds directly exposed to the high temperature exhaust gasses, thereby heating the working fluid directly without the use of intermediate media such as oil.

(13) In one embodiment, the temperature of working fluid as heated by high pressure cycle evaporator 205 does not exceed the saturation temperature of the working fluid vapor. One common type of working fluid, (Genetron R-245fa), has a saturation temperature of approximately 280° F. at a pressure of 390 psia. High pressure cycle evaporator 205, such as the GBS series of brazed plate heat exchangers manufactured and sold by GEA Heat Exchangers GmbH of Bochum, Germany, can be used in this embodiment to heat this particular working fluid to 280° F. at a pressure of 390 psia. As the amount of heat energy transferred to the working fluid increases to a point, the enthalpy of the working fluid will increase and the proportion of vaporized working fluid to liquid working fluid will increase, but the temperature will not exceed 280° F. at a pressure of 390 psia. If the system pressure is increased without adding any additional heat energy, the working fluid temperature will increase but the fluid maintains a constant enthalpy. Similarly, if the system pressure is decreased adiabatically, the working fluid temperature will decrease but the fluid will maintain a constant enthalpy. Were a superheater to be employed to transfer sufficient additional heat energy to the working fluid, the enthalpy of the heated working fluid would continue to increase until the working fluid in this example would eventually be completely vaporized and its temperature would then begin to exceed 280° F. at the pressure of 390 psia. This process of increasing the enthalpy of the working fluid to a point such that the temperature of the heated working fluid exceeds its temperature of vaporization at the operative pressure is referred to as superheating. However, the 2 P ORC system of this embodiment utilizes a wet working fluid throughout and does not require or utilize a superheater or superheated working fluid. Superheating typically requires recuperation to prevent loss of heat energy in the post-expansion working fluid and the elimination of superheated working fluid and the recuperation process represents an improvement over the prior art. The proportion of liquid state working fluid to vapor state working fluid at any point in the system may vary from completely liquid to completely vaporized depending upon the enthalpy and pressure of the working fluid at that point.

(14) Heat energy contained in the jacket cooling water from the prime mover is supplied at inlet 208 to a jacket water distribution subsystem 210, which consists of a series flow control valves such as the D08 series of proportional control valves available from Continental Hydraulics of Savage, Minn. Under the control of microprocessor-based control subsystem 219 such as the DirectLogic series of programmable logic controllers (PLCs) available from Automation Direct of Cumming, Ga., the control valves in the jacket water distribution system outlet 211 provide the requisite amount of heated jacket water to the high pressure cycle preheater 212 at inlet 213 and to the low pressure cycle preheater and evaporator 215 at inlet 214. These preheaters and evaporators may also be those such as the GBS series of brazed plate heat exchangers manufactured and sold by GEA Heat Exchangers GmbH of Bochum. Germany.

(15) In one embodiment, the low pressure cycle preheater and evaporator 215 described above is a single unit. In one embodiment, the low pressure cycle preheater and evaporator 215 comprises two separate units of similar origin and functionality. In one embodiment, one or more separate preheaters and/or evaporators may be used. All of the heated jacket water received at inlet 208 is provided to either inlet 213 or inlet 214. After passing through the high pressure cycle preheater 212 and the low pressure cycle preheater and evaporator 215, the reduced-temperature jacket water is returned via outlets 216 and 217, respectively, to inlet 218 of jacket water distribution subsystem 210 where it is returned to the prime mover via outlet 209 for recirculation. In this embodiment, the temperature of the jacket water at outlet 211 is approximately 195° F. Subsequent to the transfer of heat within the high pressure cycle evaporator 205 and low pressure cycle preheater and evaporator 215, the temperature of the jacket water at inlet 218 is approximately 160° F. The temperature of the jacket water returned to the prime mover at outlet 209 is maintained within the manufacturer's specified range for proper operation of the prime mover. For the Jenbacher 316 ICE, this range is nominally 50° C. (122° F.) to 90° C. (194° F.).

(16) In one embodiment, high pressure cycle preheater 212 heats the working fluid to the saturation temperature of the working fluid at the operating pressure. In one embodiment, high pressure cycle preheater 212 heats the working fluid to a temperature less than the saturation temperature of the working fluid. For example, high pressure cycle preheater 212 may heat the working fluid to a temperature of 280° F. at a pressure of 390 psia or any other temperature between the working fluid temperature at inlet 221 (nominally 90° F.) and 280° F. However, the high pressure cycle preheater 212 can only heat the working fluid to a maximum temperature that, owing to limitations of the heat transfer apparatus and laws of thermodynamics, approaches but may never exceed the maximum temperature of the input flow of heated jacket water at inlet 213, which in the preferred embodiment is approximately 195° F. A further discussion of the difference between the temperature of input heat energy and the maximum temperature of the heated working fluid output (known as the “pinch”) is provided below. Heating the working fluid to a greater temperature will necessitate a higher grade of waste heat energy input to jacket water distribution subsystem 210.

(17) Control subsystem 219 is also operatively coupled to a plurality of sensors, control valves, and other control and monitoring devices throughout the 2 P ORC system. To maintain clarity of the Figures, these operative couplings are not depicted in FIG. 2 but are well known to those of ordinary skill in the art. The correct allocation of jacket water heat energy is essential for optimization of 2 P ORC operation, and the method for determining and accomplishing this distribution as implemented by control subsystem 219 is described more fully below.

(18) In one embodiment, 2 P ORC system 200 utilizes a single closed loop of working fluid typically comprising a mixture of lubrication oil and organic refrigerant suitable for heating and expansion within the range of temperatures provided by the prime mover. By way of example and not limitation, the refrigerant may be R-245fa, commercially known as Genetron® and manufactured by Honeywell. The performance of the working fluid described in association with FIG. 4 is similar but not identical to R-245fa. However, any organic refrigerant including but not limited to R123, R134A, R22, and the like as well as any other suitable hydrocarbons or other fluids may be employed in other embodiments. In some embodiments, a small percentage of lubrication oil by volume is mixed with the refrigerant for lubrication purposes. Any miscible oil suitable for the intended purpose may be used, including but not limited to Emkarate RL 100E refrigerant lubricant, product number 4317-66 manufactured by Nu-Calgon.

(19) The working fluid is pressurized by centrifugal fluid pumps and variable frequency drive (“VFD”) motors 220 and 239 collectively referred to as VFD pumps, operatively monitored and controlled by control subsystem 219. In one embodiment, a single VFD pump may be utilized with suitable valves and controls to serve both ORC cycles. Within the high pressure ORC cycle, VFD pump 220 pressurizes the working fluid to a nominal pressure of 400 psia to cause the working fluid to flow directly through high pressure cycle preheater 212 where it receives heat energy from a portion of the heated jacket water, and then directly to high pressure cycle evaporator 205 where it receives additional heat energy from the exhaust gas flow. The combined heat energy transferred to the working fluid as it passes through these two evaporators causes the working fluid to change state from a heated liquid to a saturated heated vapor. In some embodiments, the heated working fluid may be partially in a liquid state and partially in a vaporized state. The heated and vaporized working fluid is applied to the input of the high pressure cycle expander 224 at an approximate pressure of 390±100 psia and a temperature of 280±25° F. Following expansion, the working fluid flows directly from the expander outlet via 226 at an approximate pressure of 90±30 psia and an approximate temperature of 185±20° F. to a pressurized tank serving as a high pressure cycle separator 227 where any liquid phase portion of the working fluid in equilibrium with the vapor phase portion of the working fluid within the separator may be removed at the bottom. The remaining working fluid in its vapor phase leaves the separator at or near the top and is retained for use in the low pressure ORC cycle, described below, while the liquid working fluid is conveyed directly via 229 to a pressurized tank serving as a low pressure cycle separator 230. In another embodiment, low pressure cycle separator 230 is optional and may be omitted. In such embodiment, low pressure cycle expander outlet 244 may be directly coupled to inlet 231 of condenser subsystem 232 such as the fin fan air cooled condensers available from Guntner U.S. LLC of Schaumburg, Ill., and outlet 229 may be directly coupled via a throttle valve to inlet 231 of condenser subsystem 232.

(20) In some embodiments, condenser subsystem may be a water cooled condenser where cold water input is supplied at inlet 233 and subsequently outlet at 234. In some embodiments, condenser subsystem 232 may be an air-cooled condenser. In some embodiments, condenser subsystem 232 may be utilized to provide heat energy for a desirable secondary purpose, including but not limited to the heating of buildings, domestic or industrial hot water, heating bacterial cultures used for anaerobic digestion of biodegradable waste materials, and the like. In one embodiment, condenser subsystem 232 may be cooled by any suitable alternative means, including but not limited to those utilizing natural environmental resources to dissipate the residual heat energy in the working fluid. The condensed working fluid, now in its liquid state at an approximate temperature of 84° F., is conveyed via outlet 235 directly to working fluid receiver 237 and conveyed via 238 directly to low pressure cycle VFD pump 239. Low pressure cycle VFD pump 239 provides the motive force (nominally 95 psia in this embodiment) necessary to pressurize the low pressure ORC cycle and also provides a portion of the motive force necessary to pressurize the high pressure ORC cycle, the balance of which is provided by high pressure cycle VFD pump 220. In one embodiment, a single VFD pump may provide sufficient motive force for both cycles.

(21) Low pressure cycle VFD pump 239 provides liquid state working fluid via 240 directly to the input of low pressure cycle preheater and evaporator 215, which transfers heat energy from a portion of the jacket water to the working fluid to heat and effect a change of state of the working fluid from liquid to partially or fully vaporized state. The fully or partially vaporized working fluid, at approximate pressure of 90 psia and approximate temperature of 160° F., is then directly conveyed to high pressure cycle separator 227 where it is combined with the partially or fully vaporized working fluid previously expanded in the high pressure cycle expander 224. The partially or fully vaporized working fluid from both sources is applied directly to the inlet 228 of low pressure cycle expander 242 at an approximate pressure of 90±15 psia and approximate temperature of 160°±10° F. Within the expander, the partially or fully vaporized working fluid is expanded, removed at outlet 244 at an approximate pressure of 27 psia and approximate temperature of 113° F., directly conveyed to low pressure cycle separator 230, condenser subsystem 232, and then to VFD pump 239 for repressurization as previously described.

(22) High pressure and low pressure cycle expanders 224 and 242 may be any devices capable of translating a decrease in pressure into mechanical energy, including but not limited to screw-type expanders, other positive displacement machines such as scroll expanders or turbines, and the like. In multi-pressure systems including the 2 P ORC system, the expanders may be of similar or different types. In some embodiments, the expanders will be identical machines of the twin screw configuration as taught by Stosic in U.S. Pat. No. 6,296,461. These expanders can be of identical characteristics or may be different.

(23) Such units are available, for example, in the XRV series from Howden Compressors of Glasgow, Scotland. Such expanders utilized in association with the specific temperatures discussed in association with Figures 204 herein are twin screw expanders and operable with wet (i.e., non-superheated) working fluid from the input through to the output of these expanders. They can thus be operated at much lower temperatures than expanders that require superheated working fluid. They can also be utilized with lower temperature heat sources than those that will superheat typical working fluids such as disclosed herein if the ORC system seeks to utilize up to all of the available heat energy from such a source.

(24) High pressure cycle expander 224 is operatively coupled to electric generator 225, such as the Magnaplus series available from Marathon Electric of Wausau, Wis., so that the mechanical energy produced by expansion of the working fluid may be converted into electric power. Similarly, low pressure cycle expander 242 is operatively coupled to electric generator 243 of similar make and origin. Either or both generators may be coupled to the local power grid for the purpose of delivering electrical energy to the grid.

(25) In some embodiments, either or both of these generators may be used to provide power for local use, particularly when commercial electric power is not available at the location of the prime mover and 2 P ORC system. This power may be used for the parasitic loads of the ORC and prime mover, including the numerous pumps and condenser systems often used to support system operation.

(26) The generators may be of the synchronous or asynchronous type, depending upon the particular requirements of the system. In one embodiment, the generators are asynchronous induction machines with their stators operatively coupled to the commercial power grid so that the mechanical energy imparted by the expander to the rotor of the induction machine causes alternating current electric power to be generated and delivered to the commercial power grid.

(27) In one embodiment, the mechanical power from the expander shafts may be coupled to one or more other device or system, including but not limited to the prime mover, a pump, fans, and other power utilizing structure or systems in lieu of being coupled to an electric generator.

(28) From the foregoing, it can be seen that the decrease in pressure of the single working fluid in the 2 P ORC system that results from its expansion occurs partially in the high pressure cycle expander 224 and partially in the low pressure cycle expander 242. This distribution and proportion of pressure reduction between the two expanders is one substantial benefit of this invention. As with all physical components, certain operating limitations are imposed on the expanders due to the constraints of fabrication materials, size, and geometry. The prior art does not allow the capture and use of all available heat energy from the prime mover, as is taught in the detailed embodiment described herein, or the heat energy from other prime movers in different applications, for conversion using a single expander and single working fluid or multiple expanders and a shared single working fluid. Attempting to do so would result in the dissipation of wasted heat energy in the ORC system condenser subsystem. By dividing the expansion of highly pressurized working fluid between two expanders, arranged in what can be essentially a series configuration with a precise allocation of the available input heat energy between the two interconnected ORC cycles with a single shared working fluid, better, and in some embodiments the most efficient, operation and output of recovered energy is realized. Additionally, this may also be characterized as an induction configuration with two sources of fully or partially vaporized working fluid supplied to the low pressure cycle expander 242.

(29) ORC waste heat recovery systems can be inherently inefficient due to a number of factors. Notably, the physical characteristics of the chosen working fluid can limit the range of temperatures within which the ORC system can effectively convert heat energy via the expansion of pressurized working fluid vapor. Effective heat energy transfer through the heat exchange subsystems, including the thermal oil heat transfer subsystem 203, high pressure cycle evaporator 205, and low pressure cycle preheater and evaporator 215 may each approach 80% only under ideal conditions and may actually yield lower performance than 80%. When cascaded, these sub-unity efficiencies are multiplied and yield an even lower total effective transfer (80% of 80% is 64%). Further, the use of recuperation processes within an ORC system constitute an attempt to recover a portion of excess heat energy that has previously be applied to the system but is not useful for conversion to electrical or mechanical energy and is therefore potentially wasted. As with any thermal process, recuperation is not fully efficient so heat energy is inevitably lost. As a result, in these types of prior art systems much of the available waste heat energy produced by the prime mover is not actually being recovered and transferred to the working fluid. Further, there are significant heat losses within the system due in large measure to the considerable residual heat energy that remains in the post-expansion working fluid and which must be dissipated by the condenser system prior to repressurization by the VFD pump(s). The combined effect of these various losses applied to a prior art ORC system depicted in FIG. 1 that utilize a single twin screw expander, evaporator, and condenser as generally described above along with the same working fluid (R-245fa) can achieve a nominal efficiency of approximately 7% in sustained operation when supplied with the waste heat energy available from a suitable prime mover, such as the Jenbacher J316 in one embodiment taught herein.

(30) Embodiments of 2 P ORC specified in FIGS. 2-4 and associated text above can improve, and in some embodiments dramatically improve, upon this performance. When supplied with the waste heat energy available from a Jenbacher J316 as the specified prime mover to the particular system identified above, approximately 921 kW of recoverable waste heat energy from exhaust gas above 356° F. and jacket cooling water heat is available for recovery and use by the 2 P ORC system. Approximately 458 kW is available from the exhaust gas flow and the remaining 463 kW is present in the jacket water. When all of the available 458 kW of waste heat energy from the exhaust gas flow is provided to the high pressure cycle evaporator 205 via thermal oil heat transfer subsystem 203, 216 kW of available waste heat energy from the jacket cooling water is applied to high pressure cycle preheater 212, and the remaining 247 kW of available waste heat energy from the jacket cooling water is applied low pressure cycle preheater and evaporator 215, the 2 P ORC system can produce at least approximately 45 kW of electric power from high pressure cycle generator 225 and another 58 kW of electric power will be produced by low pressure cycle generator 243. The combined 103 kW of electric power generated by the 2 P ORC system constitutes an overall conversion efficiency of 11.2% of the waste heat energy of 920 kW available from the prime mover. Accordingly, the 2 P ORC system provides an increase of 58% compared to the nominal 7% conversion efficiency of the present art system. This represents a very significant improvement by industry standards.

(31) Additionally, the prior art multiple ORC+superheating systems inherently allocate available heat energy in a fashion that cannot be converted and therefore, in some embodiments, is recovered by the recuperation process to salvage some efficiency. Since, however, the superheating/recuperation process itself imposes substantial energy loss to drive the process, the 2 P ORC system specified in association with FIGS. 2-4 is substantially more efficient than these types of processes because all or in any event more available heat is allocated to generating power from the specified closed wet working fluid multiple ORC system.

(32) Another significant advantage of the specified 2 P ORC system is its ability to fully utilize up to all of the recoverable waste heat energy available in the jacket water of a suitably-matched prime mover. In prior art systems known to the applicants, only a portion of the heat energy in the jacket water can be utilized and the remainder is cooled through the use of conventional radiators that require additional electric power to operate the cooling fans. In the specified embodiment of this specification, however, the 2 P ORC system is combined with waste heat generated by, for example, a widely-used prime mover (such as the Jenbacher J316 internal combustion engine) so that up to all of the available heat energy in the jacket water flow may be fed to the 2 P ORC system for waste heat energy conversion into electric power. This can obviate the need for a traditional radiator system to support the prime mover that would consume rather than generate electric power. In addition, a substantial portion of the waste heat in the exhaust gas flow can be captured and converted by the specified 2 P ORC system and others disclosed herein. Embodiments of these systems also can reduce and, in some embodiments, minimize thermal pollution of the environment.

(33) The distribution of waste heat energy from each source to each of the two ORC cycles in the 2 P ORC system is an operating condition that can be calculated and maintained in order to achieve desired, and in some embodiments, optimal performance. The method of determining the distribution of heat energy between the high and low pressure cycles also overcomes the limitations of the prior art which require heat recuperation from the working fluid to minimize losses and therefore constitutes a significant improvement over the prior art. The method may also be utilized to determine and maintain any desired lesser degree of utilization of available waste heat available from the prime mover at the most efficient point of system operation. In addition the following description, the method of determining the 2 P ORC system control and set points is provided as a flow chart in FIG. 3.

(34) The first steps in the iterative method of determining the control and set points for 2 P ORC system operation require the computation of the available heat energies in the exhaust gas flow and the jacket cooling water (301, 302). For the exhaust gas, the temperature differential ΔT(ex) between the exhaust gas flow T(ex_1) at the input 201 and T(ex_2) at the output 202 to the thermal oil heat transfer subsystem 203 may be measured if such apparatus is available for measurement under operating conditions. If said apparatus is not available, the available heat energy from the exhaust gas flow may be determined from the manufacturer's specification data for the prime mover. If neither is available, the values may be estimated based on best available information, recognizing that errors may be introduced by inaccurate estimations and that further refinement and parameter adjustment will likely be required to compensate for difference between estimated and actual values later realized in practice.

(35) For the jacket water, the same temperature differential between T(jw_1) at the input 208 and T(jw_2) at the output 209 of the jacket water distribution subsystem 210 may be measured, calculated, or estimated using best available resources (303).

(36) The mass flow rates M(ex) of the exhaust gas flow and M(jw) of the jacket water flow of the prime mover may be measured, calculated, or estimated based on best available information (304).

(37) The heat energy Q(ex) contained in the exhaust gas is defined as

(38) $Q\left(\mathrm{ex}\right)=M\left(\mathrm{ex}\right){\int }_{T\left(\mathrm{ex}\_2\right)}^{T\left(\mathrm{ex}\_1\right)}\mathrm{Cp}\mathrm{dT}$
where Cp is the specific heat of the exhaust gas mixture, which is generally calculated based on the composition of the exhaust gas and dT is the variable of integration. Assuming that the temperature differential is sufficiently low so that Cp may be considered to be constant at its mean value, Q(ex) may be calculated (305) via
Q(ex)=M*Cp*ΔT(ex)
where ΔT(ex)=T(ex_1)−T(ex_2). The minimum final temperature of the exhaust gas, T(ex_2), is normally set by the engine manufacturer at some safe level above the acid dew point temperature of the gas depending on the fuel used. As previously described, cooling the exhaust gas below the acid dew point will likely cause damage, including corrosion to the engine exhaust system and waste heat recovery heat exchanger.

(39) The temperature of the heated working fluid may approach that of the waste heat source but never be able to reach it due to the limitations imposed by the Second Law of Thermodynamics and the physical limitations of heat exchangers used to transfer the heat from the source to the working fluid. As a principal consequence, the final temperature of the working fluid being heated can never reach the highest temperature of the source being cooled.

(40) FIG. 4 is a general depiction of the heat energy versus temperature of the source heat and working fluid during a heat transfer process at a pressure similar to that which may occur in the low pressure ORC cycle. The data depicted in this figure is illustrative of the performance of some embodiments but is not meant to be an accurate numerical representation of any particular embodiment. However, the properties of the example working fluid closely resemble those of R-245fa Genetron refrigerant which exhibits a saturation temperature of 70° C. at a nominal pressure of 90 psia as may exist at inlet 228 to low pressure cycle expander 242. Line segment 401 represents the source heat and segment 402 represents the working fluid. Point 404 depicts the state of the jacket water at inlet 214 and point 403 represents the state of the jacket water at outlet 217 of low pressure cycle preheater and evaporator 215. In this example, the jacket water experiences a decrease in temperature of approximately 35° C. (from 100° C. to 65° C.). In a similar manner, point 405 represents the state of the working fluid at inlet 240 and point 406 represents the state of the working fluid at outlet 241 of low pressure cycle preheater and evaporator 215. Along this path, it can be seen that the temperature of the working fluid increases from 30° C. to 70° C., which in this example is the temperature at which the working fluid begins to vaporize at the liquid saturation temperature. Although the temperature does not increase beyond this vaporization temperature in this example, the heat energy content of the working fluid continues to increase as it receives additional heat energy from the jacket water and the working fluid is increasingly vaporized.

(41) During this heat transfer process, the paths representing the working fluid heating and jacket water cooling processes do not intersect, lest there be no additional heat transfer between the source and working fluid, in accordance with the Second Law of Thermodynamics. That is, the temperature of the working fluid can never equal that of the waste heat energy input and will always be lower by a certain amount. The temperature at the closest distance between these two paths, point 407, is normally referred to as the “pinch point”. It is the minimum temperature difference between the source and working fluid at any point in the heat exchanger. In the design of ORC power plant evaporators, condensers, heat exchangers, and the like, the pinch point is used to determine the pressure, temperature and mass flow of the working fluid leaving the heat exchanger.

(42) In some embodiments, the pinch may be selected to be as low as 3° C. and as high as 10° C. However, the pinch is usually selected by ORC design engineers to be approximately 5° to 10° C. depending on the absolute temperature of the source. The pinch value depicted in the example of FIG. 4 is approximately 5° C. Selection of a larger pinch value reduces system efficiency while selection of a pinch value that is too small increases surface requirements of the heat exchanger and corresponding cost. Since the temperature of the waste heat energy flow decreases as it passes through the evaporator, in the preferred embodiment the working fluid output is in closest contact with the waste heat energy input and the working fluid input in closest contact with the waste heat energy output (counterflow).

(43) In one embodiment, the heat contained in the prime mover's exhaust gas is applied to high pressure cycle heat exchanger 205 either directly or via thermal oil heat transfer subsystem 203, and the design conditions of the high pressure ORC cycle are generally set by the temperature and pressure specifications and limitations of the expander. Those limits are imposed by the heat exchanger's pinch point. In particular, the temperature and pressure of the working fluid heated by the exhaust gas flow may not exceed the rated values for the expander's inlet.

(44) Having determined the heat energy of the exhaust gas and assuming that all of this heat is transferred to the working fluid, the mass flow rate of the working fluid M(wf) may be computed (306) via
M(wf)=Q(ex)/ΔH(wf_hpe)
where ΔH(wf_hpe) represents the difference in the enthalpy, or total energy, of the working fluid between the high pressure cycle evaporator 205 outlet 223 and inlet 222 which corresponds to a temperature approximately 5° C. below the maximum temperature of the low temperature source. In other words, the working fluid mass flow rate can be determined by the amount of exhaust heat used and by the minimum and maximum enthalpy of the working fluid heated either directly or indirectly (via thermal oil loop) by the exhaust gas.

(45) The total heat energy available from all jacket cooling water is typically provided by the engine manufacturer and also may be calculated (307) via
Q(jw_tot)=M(jw)*Cp*ΔT(jw)
where ΔT(jw) represents the difference in the temperature of the jacket cooling water between the inlet 208 and the outlet 209 of the jacket water distribution subsystem 210.

(46) As previously described, waste heat energy from the jacket cooling water may be provided to the high pressure ORC cycle via the high pressure cycle preheater 212 that receives a portion of the jacket cooling water from jacket water distribution subsystem 210, depending on the maximum temperature of the jacket water. The amount of jacket water heat energy required for the high pressure cycle may be calculated (308) via
Q(jw_hp)=M(wf)*ΔH(wf_hpp)

(47) where ΔH(wf_hpp) represents the difference in the enthalpy of the working fluid between the outlet 222 and the inlet 221 to high pressure cycle preheater 212.

(48) The quantity of jacket water provided to the high pressure cycle by jacket water distribution subsystem 210 and control subsystem 219 is determined by the temperature difference of the jacket water circuit as specified by the manufacturer of the prime mover. That mass flow rate may be calculated at the outlet 222 of high pressure cycle preheater 212 (309):
M(jw_hp)=(Q(jw_hp)/(ΔT(jw)*Cp)

(49) VFD pump 220 controls the pressure at the input to high pressure cycle expander 224, and via control subsystem 219, the mass flow rate of the working fluid in the high pressure cycle is set to achieve the desired temperature and pressure at the inlet of high pressure cycle expander 224.

(50) The total waste heat energy contained in the jacket water available for the low pressure cycle is the difference between the total jacket water heat available and that already applied to the high pressure cycle preheater 212 as calculated above:
Q(jw_lp)=Q(jw_tot)−Q(jw_hp)

(51) The temperature and pressure at low pressure cycle expander inlet 228 for optimal system performance may now be determined iteratively via the following method: 1) Assume that the temperature of the vaporized working fluid T(wf_v) is equal to the minimum temperature of the jacket water T(jw_pinch) in the low pressure cycle. This is equivalent to setting the initial value of the pinch in the cycle to zero (310). 2) Calculate the mass flow rate of the working fluid in the low pressure cycle (311) via
M(wf_lp)=Q(jw_lp)/ΔH(wf_lpe)  where ΔH(wf_lpe) represents the difference in enthalpy of the working fluid leaving the low pressure cycle preheater and evaporator 215 at 241 (where its enthalpy is maximum) and at the entry to the low pressure cycle preheater and evaporator 215 at 240. 3) Using the working fluid property tables, determine the enthalpies (312): a) H(wf_cond) of the working fluid in the low pressure cycle at the outlet 235 of condenser subsystem 232, b) H(wf_v) at the point of initial vaporization (saturated liquid), and c) H(wf_hps) at high pressure cycle separator 227 inlet flow 241. 4) Calculate heat addition at the pinch point Qp (313):
Qp=[(H(wf_v)−H(wf_cond))/(H(wf_hps)−H(wf_cond))]*Q(jw_lp) 5) Because
Qp=M(jw_lp)*Cp*(T(jw_pinch)−T(jw_o))  we may calculate (314)
T(jw_pinch)=(Qp/(M(jw_lp)*Cp))+T(jw_o)  where T(jw_pinch) is the temperature of the jacket water at the pinch point and T(jw_o) is the temperature of the jacket water at the outlet 217 of low pressure cycle preheater and evaporator 215. 6) Compare (315) T(jw_pinch) to T(wf_v). If the difference is less than 5° C. (316) (the desired pinch value), reduce T(wf_v) by 2° C. (317) and repeat the iteration. If the difference between T(jw_pinch) and T(wf_v) is greater than 5° C. (318), increase T(wf_v) by 2° C. (319) and reiterate. 7) Continue the iteration until the pinch (T(jw_pinch)−T(wf_v)) is 5° C. plus or minus 1° C.

(52) Finally, once the parameters of the low pressure cycle have been determined in this manner, the pressure at the high pressure cycle expander outlet 226 may be set to the pressure of the low pressure cycle expander inlet 228 (320). In one embodiment, one or more control valves or other means of controlling the pressure may be incorporated in the system.

(53) With respect to the depiction of heated extraction ports in the prior art systems depicted in FIGS. 5 and 6, the same possibilities exist for MP ORC systems. The condenser subsystem 232 may be replaced, in whole or in part, by an alternate subsystem that utilizes the residual heat energy present in the post-expansion working fluid for any other useful purpose.

(54) The description of this invention is intended to be enabling and not limiting. It will be evident to those skilled in the art that numerous combinations of the embodiments described above may be implemented together as well as separately, and all such combinations constitute embodiments effectively described herein.