( II )
A. Kalina: USP # 4,604,867 ~ Implementing a Thermodynamic Cycle with Intercooling
A. Kalina: USP # 4,732,005 ~ Direct Fired Power Cycle
A. Kalina: USP # 4,763,480 ~ Implementing a Thermodynamic Cycle with Recuperative Preheating
A. Kalina: USP # 4,899,545 ~ Thermodynamic Cycle
A. Kalina: USP # 4,982,568 ~ Converting Heat from Geothermal Fluid to Electric Power
A. Kalina: USP # 5,029,444 ~ Converting Low Temperature Heat to Electric Power
A. Kalina: USP # 5,095,708 ~ Converting Thermal Energy into Electric Power
A. Kalina: USP # 5,440,882 ~ Converting Heat from Geothermal Liquid and Geothermal Steam to Electric Power
A. Kalina: USP # 5,450,821 ~ Multi-Stage Combustion System for Externally Fired Power Plants
A. Kalina: USP # 5,572,871 ~ Conversion of Thermal Energy into Mechanical and Electrical Power
A. Kalina: USP # 5,588,298 ~ Supplying Heat to an Externally Fired Power System
A. Kalina & Richard Pelletier: USP # 5,649,426 ~ Implementing a Thermodynamic Cycle
A. Kalina & Lawrence Rhodes:USP # 5,822,990 ~ Converting Heat into Useful Energy Using Separate Closed Loops
A. Kalina: USP # 5,950,443 ~ Method and System of Converting Thermal Energy into a Useful Form
A. Kalina & R. Pelletier ~ USP # 5,953,918 ~ Method and Apparatus of Converting Heat to Useful Energy
 
 

---

US Patent # 4,604,867

 
Abstract --- A method and apparatus for implementing a thermodynamic cycle with intercooling, includes a condensing subsystem, a boiler, and a turbine. The boiler may include a preheater, an evaporator, and a superheater. After initial expansion in the turbine, the fluid may be diverted to a reheater to increase the temperature available for superheating. After return to the turbine and additional expansion, the fluid may be withdrawn from the turbine and cooled in an intercooler. Thereafter the fluid is returned to the turbine for additional expansion. The cooling of the turbine gas may provide additional heat for evaporation. Intercooling may provide compensation for the heat used in reheating and may provide recuperation of available heat which would otherwise remain unused following final turbine expansion.

US Cl. 60/653 ; 60/670 ; 60/649
Intl. Cl. F01K 007/38; F01K 025/00

References
U.S. Patent Documents

USP # 3,979,914 ~ Sep., 1976 ~ Weber ~ 60/678
USP # 4,433,545 ~ Feb., 1984 ~ Chang ~ 60/677

Description

BACKGROUND OF THE INVENTION

1. Field of the Invention

This invention relates generally to methods and apparatus for transforming energy from a heat source into usable form using a working fluid that is expanded and regenerated. This invention further relates to a method and apparatus for improving the heat utilization efficiency of a thermodynamic cycle.

2. Brief Description of the Background Art

In the Rankine cycle, a working fluid such as water, ammonia or a freon is evaporated in an evaporator utilizing an available heat source. The evaporated gaseous working fluid is expanded across a turbine to transform its energy into usable form. The spent gaseous working fluid is then condensed in a condenser using an available cooling medium. The pressure of the condensed working medium is increased by pumping, followed by evaporation and so on to continue the cycle.

The Exergy cycle, described in U.S. Pat. No. 4,346,561, utilizes a binary or multi-component working fluid. This cycle operates generally on the principle that a binary working fluid is pumped as a liquid to a high working pressure and is heated to partially vaporize the working fluid. The fluid is then flashed to separate high and low boiling working fluids. The low boiling component is expanded through a turbine, to drive the turbine, while the high boiling component has heat recovered for use in heating the binary working fluid prior to evaporation. The high boiling component is then mixed with the spent low boiling working fluid to absorb the spent working fluid in a condenser in the presence of a cooking medium.

The theoretical comparison of the conventional Rankine cycle and the Exergy cycle demonstrates the improved efficiency of the new cycle over the Rankine cycle when an available, relatively low temperature heat source such as ocean water, geothermal energy or the like is employed.

In applicant's further invention, referred to as the Basic Kalina cycle, the subject of U.S. Pat. No. 4,489,563, relatively lower temperature available heat is utilized to effect partial distillation of at least a portion of a multi-component fluid stream at an intermediate pressure to generate working fluid fractions of differing compositions. The fractions are used to produce at least one main rich solution which is relatively enriched with respect to the lower boiling component, and to produce one lean solution which is relatively impoverished with respect to the lower boiling component. The pressure of the main rich solution is increased; thereafter, it is evaporated to produce a charged gaseous main working fluid. The main working fluid is expanded to a low pressure level to convert energy to usable form. The spent low pressure level working fluid is condensed in a main absorption stage by dissolving with cooling in the lean solution to regenerate an initial working fluid for reuse.

In any process of converting thermal energy to a usable form, the major loss of available energy in the heat source occurs in the process of boiling or evaporating the working fluid. This loss of available energy (known as exergy or essergy) is due to the mismatch of the enthalpy-temperature characteristics of the heat source and the working fluid in the boiler. Simply put, for any given enthalpy the temperature of the heat source is always greater than the temperature of the working fluid. Ideally, this temperature difference would be almost, but not quite, zero.

This mismatch occurs both in the classical Rankine cycle, using a pure substance as a working fluid, as well as in the Kalina and Exergy cycles described above, using a mixture as the working fluid. The use of a mixture as a working fluid in the manner of the Kalina and Exergy cycles reduces these losses to a significant extent. However, it would be highly desirable to further reduce these losses in any cycle.

In the conventional Rankine cycle, the losses arising from mismatching of the enthalpy-temperature characteristics of the heat source and the working fluid would constitute about 25% of the available exergy. With a cycle such as that described in U.S. Pat. No. 4,489,563, the loss of exergy in the boiler due to enthalpy-temperature characteristics mismatching would constitute about 14% of all of the available exergy.

The overall boiling process in a thermodynamic cycle can be viewed for discussion purposes as consisting of three distinct parts: preheating, evaporation, and superheating. With conventional technology, the matching of a heat source and the working fluid is reasonably adequate during preheating. However, the quantity of heat in the temperature range suitable for superheating is generally much greater than necessary, while the quantity of heat in the temperature range suitable for evaporation is much smaller than necessary. The inventor of the present invention has appreciated that a portion of the high temperature heat which would be suitable for high temperature superheating is used for evaporation in previously known processes. This causes very large temperature differences between the two streams, and as a result, irreversible losses of exergy.

These irreversible losses may be lessened by reheating the stream of working fluid after it has been partially expanded in a turbine. However, reheating results in repeated superheating. As a result, reheating increases the necessary quantity of heat for superheating. This increase in the required heat provides better matching between the heat source and the working fluid enthalpy-temperature characteristics. However, reheating has no beneficial effect with respect to the quantity of heat necessary for evaporation. Thus, the total quantity of heat necessary per unit of weight of working fluid significantly increases with reheating. Therefore, the total weight flow rate of working fluid through the boiler turbine is reduced. Thus, the benefits of reheating are largely transitory in that the reduced weight flow rate limits the possible increase in overall efficiency that may be derived.

The ideal solution to the age old dilemma of poorly matched heat source and working fluid enthalpy-temperature characteristics would be one that makes high temperature heat available from the heat source for use in superheating thereby reducing the temperature differences during superheating, but at the same time provides lower temperature heat which minimizes the temperature differences in the process of evaporation. It should be evident that these two goals are apparently mutually inconsistent since increasing the superheating heat would appear to require either increasing the overall heating source temperature or using reheating. As discussed above, reheating has certain drawbacks, which to a large degree mitigate the partly transitory gains achieved.

Moreover, the greater the available heat for superheating, the greater would be the output temperature of the gaseous spent working fluid from the turbine. This is undesirable from an efficiency standpoint since the superheating of the exiting steam makes subsequent condensing more difficult and causes additional losses of exergy. Thus, any effort to improve efficiency with respect to one part of the cycle seems to eventually cause lower efficiency in another part of the cycle.

SUMMARY OF THE INVENTION

It is one feature of the present invention to provide a significant improvement in the efficiency of a thermodynamic cycle by permitting closer matching of the working fluid and the heat source enthalpy-temperature characteristics in the boiler. It is also a feature of the present invention to provide a system which both increases the efficiency of superheating while providing concommitant advantages during evaporation. Another feature of the present invention is to enable these advantages to be attained without necessarily adversely reducing the mass flow rate of the cycle.

In accordance with one embodiment of the present invention, a method of implementing a thermodynamic cycle includes the step of expanding a gaseous working fluid to transform its energy into a usable form. The expanded gaseous working fluid is cooled and subsequently expanded to a spent low pressure level to transform its energy into a usable form. The spent working fluid is condensed. The condensed fluid is then evaporated using the heat transferred during the cooling of the expanded gaseous working fluid.

In accordance with another embodiment of the present invention, a method of implementing a thermodynamic cycle includes the step of superheating an evaporated working fluid. The superheated fluid is expanded to transform its energy into usable form. The expanded fluid is then reheated and subsequently further expanded to transform additional energy into a usable form. The expanded, reheated fluid is cooled and again expanded, this time to a spent low pressure level to transform its energy into a usable form. The spent working fluid is condensed and subsequently evaporated using heat transferred during cooling from the expanded, reheated fluid.

In accordance with yet another embodiment of the present invention, a method for implementing a thermodynamic cycle includes the step of preheating an initial working fluid to a temperature approaching its boiling temperature. The preheated initial working fluid is split into first and second fluid streams. The first fluid stream is evaporated using a first heat source while a second fluid stream is evaporated using a second heat source. The first and second evaporated fluid streams are combined and subsequently superheated to produce a charged gaseous main working fluid. The charged gaseous main working fluid is expanded to transform its energy into a usable form. Then the expanded, charged main working fluid is reheated and again expanded. The expanded, reheated, charged main working fluid is cooled to provide the heat source for evaporating the second fluid stream. The cooled main working fluid is again expanded, this time to a spent low pressure level to transform its energy into a usable form. The spent main working fluid is cooled and condensed to form the intial working fluid.

In accordance with still another embodiment of the present invention, an apparatus for implementing a thermodynamic cycle includes a turbine device. The turbine device has first and second turbine sets each including at least one turbine stage. Each of the turbine sets has a gas inlet and a gas outlet. A turbine gas cooler is connected between the gas outlet of the first set and the gas inlet of the second set, such that most of the fluid passing through the turbine would pass through the turbine gas cooler and then back to said turbine device.

BRIEF DESCRIPTION OF THE DRAWING

FIG. 1 is a schematic representation of one system for carrying out one embodiment of the method and apparatus of the present invention;

 
FIG. 2  is a schematic representation of one exemplary embodiment of Applicant's previous invention, showing within dashed lines a schematic representation of one exemplary condensing subsystem for use in the system shown in FIG. 1;

FIG. 3 is a graph of calculated temperature in degrees Fahrenheit versus boiler heat duty or enthalpy in BTU's per hour for the exemplary embodiment of Applicant's previous invention shown in FIG. 2; and

FIG. 4  is a graph of calculated temperature in degrees Fahrenheit versus boiler heat duty or enthalpy in BTU's per hour in accordance with one exemplary embodiment of the present invention.

DESCRIPTION OF A PREFERRED EMBODIMENT

Referring to the drawing wherein like reference characters are utilized for like parts throughout the several views, a system 10, shown in FIG. 1, implements a thermodynamic cycle, in accordance with one embodiment of the present invention. The system 10 includes a boiler 102, in turn made up of a preheater 104, an evaporator 106, and a superheater 108. In addition, the system 10 includes a turbine 120, a reheater 122, an intercooler 124, and a condensing subsystem 126.

The condenser 126 may be any type of known heat rejection device. In the Rankine cycle, heat rejection occurs in a simple heat exchanger and thus, for Rankine applications, the condensing subsystem 126 may take the form of a heat exchanger or condenser. In the Kalina cycle, described in U.S. Pat. No. 4,489,563 to Kalina, the heat rejection system requires that gases leaving the turbine be mixed with a multi-component fluid stream, for example, comprised of water and ammonia, condensed and then distilled to produce the original state of the working fluid. Thus, when the present invention is used with a Kalina cycle, the distillation subsystem described in U.S. Pat. No. 4,489,563 may be utilized as the condensing subsystem 126. U.S. Pat. No. 4,489,563 is hereby expressly incorporated by reference herein.

Various types of heat sources may be used to drive the cycle of this invention. Thus, for example, heat sources with temperatures as high as, say 1000.degree. C. or more, down to low heat sources such as those obtained from ocean thermal gradients may be utilized. Heat sources such as, for example, low grade primary fuel, waste heat, geothermal heat, solar heat or ocean thermal energy conversion systems may be implemented with the present invention.

A variety of working fluids may be used in conjunction with this system depending on the kind of condensing subsystem 126 utilized. In conjunction with a condensing subsystem 126 as described in the U.S. patent incorporated by reference herein, any multi-component working fluid that comprises a lower boiling point fluid and a relatively higher boiling point fluid may be utilized. Thus, for example, the working fluid employed may be an ammonia-water mixture, two or more hydrocarbons, two or more freons, mixtures of hydrocarbons and freons or the like. In general, the fluid may be mixtures of any number of compounds with favorable thermodynamic characteristics and solubility. However, when implementing the conventional Rankine cycle, a conventional single component working fluid such as water, ammonia, or freon may be utilized.

As shown in FIG. 1, a completely condensed working fluid passes through a preheater 104 where it is heated to a temperature a few degrees below its boiling temperature. This preheating is provided by the cooling of all streams of a heat source indicated in dashed lines through the preheater 104. The working fluid which exits the preheater 104 is divided at point 128 into two separate streams.

A first stream, separated at point 128, enters the evaporator 106 while the second stream enters the intercooler 124. The first stream is heated in the evaporator 106 by the countercurrent heating fluid flow indicated in dashed lines through the evaporator 106 and communicating with the heating fluid flow through the preheater 104. The second fluid stream passing through the intercooler 124 is heated by the fluid flow proceeding along line 130. Both the first and second streams are completely evaporated and initially superheated. Each of the streams has approximately the same pressure and temperature but the streams may have different flow rates. The fluid streams from the evaporator 106 and intercooler 124 are then recombined at point 132.

The combined stream of working fluid is sent into the superheater 108 where it is finally superheated by heat exchange with only part of the heat source stream indicated by dashed lines extending through the superheater 108. Thus, the heat source stream extending from point 25 to point 26 passes first through the superheater 108, then through the evaporator 106 and finally through the preheater 104. The enthalpy-temperature characteristics of the illustrated heating fluid stream, indicated by the line A in FIG. 4, is linear.

From the superheater 108, the total stream of working fluid enters the first turbine set 134 of turbine 120. The turbine set 134 includes one or more stages 136 and, in the illustrated embodiment, the first turbine set 134 includes three stages 136. In the first turbine set 134 the working fluid expands to a first intermediate pressure thereby converting thermal energy into mechanical energy.

The whole working fluid stream from the first turbine set 134 is reheated in the reheater 122. The reheater 122 is a conventional superheater or heat exchanger. With this reheating process the remaining portion of the heat source stream, split at point 138 from the flow from point 25 to point 26, is utilized. Having been reheated to a high temperature, the stream of working fluid leaves the reheater 122 and travels to the second turbine set 140. At the same time the heating fluid flow from point 51 to point 53 is returned to the main heating fluid flow at point 142 to contribute to the processes in the evaporator 106 and preheater 104. The second turbine set 140 may include a number of stages 136. In the illustrated embodiment, the second turbine set 140 is shown as having four stages, however, the number of stages in each of the turbine sets described herein may be varied widely depending on particular circumstances.

The working fluid in the second turbine set 140 is expanded from the first intermediate pressure to a second intermediate pressure, thus generating power. The total stream of working fluid is then sent to the intercooler 124 where it is cooled, providing the heat necessary for the evaporation of the second working fluid stream. The intercooler 124 may be a simple heat exchanger. The fluid stream travels along the line 130 to the last turbine set 144.

The last turbine set 144 is illustrated as having only a single stage 136. However, the number of stages in the last turbine set 144 may be subject to considerable variation depending on specific circumstances. The working fluid expands to the final spent fluid pressure level thus producing additional power. From the last turbine set 144 the fluid stream is passed through the condensing subsystem 126 where it is condensed, pumped to a higher pressure and sent to the preheater 104 to continue the cycle.

A Kalina cycle condensing subsystem 126', shown in FIG. 2, may be used as the condensing subsystem 126 in the system shown in FIG. 1. In analyzing the condensing subsystem 126', it is useful to commence with the point in the subsystem identified by reference numeral 1 comprising the initial composite stream having an initial composition of higher and lower boiling components in the form of ammonia and water. At point 1 the initial composite stream is at a spent low pressure level. lt is pumped by means of a pump 151 to an intermediate pressure level where its pressure parameters will be as at point 2 following the pump 151.

From point 2 of the flow line, the initial composite stream at an intermediate pressure is heated consecutively in the heat exchanger 154, in the recuperator 156 and in the main heat exchanger 158.

The initial composite stream is heated in the heat exchanger 154, in the recuperator 156 and in the main heat exchanger 158 by heat exchange with the spent composite working fluid from the turbine 120'. When the system of FIG. 1 is being implemented with the condensing subsystem 126' the turbine 120 may be used in place of the turbine 120'. In addition, in the heat exchanger 154 the initial composite stream is heated by the condensation stream as will be hereinafter described. In the recuperator 156 the initial composite stream is further heated by the condensation stream and by heat exchange with lean and rich working fluid fractions as will be hereinafter described.

The heating in the main heat exchanger 158 is performed only by the heat of the flow from the turbine outlet and, as such, is essentially compensation for under recuperation.

At point 5 between the main heat exchanger 158 and the separator stage 160 the initial composite stream has been subjected to distillation at the intermediate pressure in the distillation system comprising the heat exchangers 154 and 158 and the recuperator 156. If desired, auxiliary heating means from any suitable or available heat source may be employed in any one of the heat exchangers 154 or 158 or in the recuperator 156.

At point 5 the initial composite stream has been partially evaporated in the distillation system and is sent to the gravity separator stage 160. In this stage 160 the enriched vapor faction which has been generated in the distillation system, and which is enriched with the low boiling component, namely ammonia, is separated from the remainder of the initial composite stream to produce an enriched vapor fraction at point 6 and a stripped liquid fraction at point 7 from which the enriched vapor fraction has been stripped.

Further, the stripped liquid fraction from point 7 is divided into first and second stripped liquid fraction streams having parameters as at points 8 and 10 respectively.

The enriched fraction at point 6 is enriched with the lower boiling component, namely ammonia, relatively to a lean working fluid fraction as discussed below.

The first enriched vapor fraction stream from point 6 is mixed with the first stripped liquid fraction stream at point 8 to provide a rich working fluid fraction at point 9.

The rich working fluid fraction is enriched relatively to the composite working fluid (as hereinafter discussed) with the lower boiling component comprising ammonia. The lean working fluid fraction, on the other hand, is impoverished relatively to the composite working fluid (as hereinafter discussed) with respect to the lower boiling component.

The second stripped liquid fraction at point 10 comprises the remaining part of the initial composite stream and is used to constitute the condensation stream.

The rich working fluid fraction at point 9 is partially condensed in the recuperator 156 to point 11. Thereafter the rich working fluid fraction is further cooled and condensed in the preheater 162 (from point 11 to 13), and is finally condensed in the absorption stage 152 by means of heat exchange with a cooling water supply through points 23 to 24.

The rich working fluid fraction is pumped to a charged high pressure level by means of the pump 166. Thereafter it passes through the preheater 162 to arrive at point 22. From point 22 it may continue through the system shown in FIG. 1.

When a Kalina cycle is implemented, the composite working fluid at point 38 exiting from the turbine 120 has such a low pressure that it cannot be condensed at this pressure and at the available ambient temperature. From point 38 the spent composite working fluid flows through the main heat exchanger 158, through the recuperator 156 and through the heat exchanger 154. Here it is partially condensed and the released heat is used to preheat the incoming flow as previously discussed.

The spent composite working fluid at point 17 is then mixed with the condensation stream at point 19. At point 19 the condensation stream has been throttled from point 20 to reduce its presure to the low presure level of the spent composite working fluid at point 17. The resultant mixture is then fed from point 18 through the absorption stage 152 where the spent composite working fluid is absorbed in the condensation stream to regenerate the initial composite stream at point 1.

The intercooling process accomplished by the intercooler 124, shown in FIG. 1, reduces the output of the last turbine stage per pound of working fluid. However, intercooling also enables reheating without sacrificing the quantity of working fluid per pound. Thus, compared to reheating without intercooling, the use of intercooling achieves significant advantages.

The heat returned by the intercooler 124 to the evaporation process is advantageously approximately equal the heat consumed in the reheater 122. This assures that the weight flow rate of the working fluid is restored. Then it is not necessary to decrease the mass flow rate of the working fluid to accommodate the higher temperature reheating process.

The parameters of flow at points 40, 41, 42 and 43 are design variables and can be chosen in a way to obtain the maximum advantage from the system 10. One skilled in the art will be able to select the design variables to maximize performance under the various circumstances that may be encountered.

The parameters of the various process points, shown in FIG. 1, are subject to considerable variation depending on specific circumstances. However, as a general guide or rule of thumb to the design of systems of this type, it can be pointed out that it may often be advantageous to make the temperature at point 40 as close as possible to the temperature of point 37 so that the efficiencies of the first turbine set 134 and the second turbine set 140 are close to equal. In addition, it may be desirable in many situations to design the system so that the temperature at point 42 is generally higher than the temperature of the saturated vapor of the working fluid in the evaporator 106. It may also often be desirable to make the temperature at point 43 generally higher than the temperature of a saturated liquid of the working fluid in the boiler 102.

While a single pressure in the evaporator 106 and intercooler 124 is utilized in the illustrated embodiment, one skillled in the art will appreciate that dual, triple and even higher numbers of boiler pressures may be selected for specific circumstances. The present invention is also applicable to multiple boiling cycles. While special advantages may be achieved through the use of intercooler 124 heat in the evaporation process, the use of the intercooler 124 between turbine sets can be applied to any portion of a thermodynamic system where there is a shortage of adequate temperature heat. Intercooling could provide heat to supplement boiling or to supplement heating in a superheater.

It should be understood that the present invention is not limited to the use of intercooling in combination with reheating. Although this combination results in significant advantages, many advantages can be achieved with intercooling without reheating. For example, intercooling may be utilized without reheating whenever the fluid exiting from the final turbine stage is superheated. In general, it is important that intercooling be taken between turbine stages in order to obtain a sufficiently high fluid temperature.

It is generally advantageous that at least most of the fluid flow through the turbine be passed through the intercooler. Even more advantageously, substantially all of the flow through the turbine is passed through the intercooler. Advantageously, substantially all of the cooled fluid is returned to the turbine for further expansion.

The advantages of the present invention may be appreciated by comparison of FIGS. 3 and 4. In FIG. 3 a boiler heat duty cycle for a thermodynamic cycle is illustrated for a system of the type shown in FIG. 2, pursuant to the teachings of U.S. Pat. No. 4,489,563, previously incorporated herein. The heat source is indicated by the line A while the working fluid is indicated by the line B. The enthalpy-temperature characteristics of the working fluid during preheating are represented by the curve portion B1. Similarly, evaporation is indicated by the portion B2 and superheating is indicated by the portion B3. The pinch point is located in the region of the intersection of the portions B1 and B2. The extent of the gap between the curves A and B represents irreversible inefficiencies in the system which are sought to be minimized by the present invention. During superheating, excessive heat is available, while during evaporation insufficient heat is available.

Referring now to FIG. 4, calculated temperature versus enthalpy or heat duty in a boiler is shown for an illustrative embodiment of the present invention. The working fluid is represented by curve C while the heat source fluid is represented by the curve A. The points on the graph correspond to points on FIG. 1. Instead of having three approximately linear regions, the graph shows that the working fluid has approximately four linear regions with the present invention. In the region between points 22 and 44, 46, preheating is occuring in the manner generally identical to that occuring with Applicant's previous invention, represented by portion B1 in FIG. 3. Evaporation is represented by the curve portion between the points 44, 46 and 48, 49 and the saturated liquid point is indicated as "SL" while the saturated vapor point is indicated as "SV". The curve portion between points 48, 49 and 30, 41 represents superheating with reheating following efficient evaporation. It can be seen that the curve portion between points 40 and 30, 41 closely follows the heat source line A and therefore results in close temperature matching. In general, the overall configuration of the curve, particularly, the portion between points SV and 30, 41 more closely approximates the heat source line A than was previously possible so that greater efficiencies may be realized with the present invention.

In order to further illustrate the advantages that can be obtained by the present invention, two sets of calculations were performed. In both sets, the same heat source was utilized. The first set of calculations is related to an illustrative power cycle in accordance with the system shown in FIG. 2. In this illustrative cycle the working fluid is a water-ammonia mixture with a concentration of 72.5 weight percent of ammonia (weight of ammonia to total weight). The parameters for the theoretical calculations which were performed utilizing standard ammonia-water enthalpy/concentration diagrams are set forth in Table 1 below. In this table the points set forth in the first column correspond to points set forth in FIG. 2.

                  TABLE 1
    ______________________________________
                                  NH.sub.4 Con-
                                  centration
    Point
         Temp.    Press.   Enthalpy
                                  lbs NH.sub.4 /
                                          W
    No.  (.degree.F.)
                  (psia)   (BTU/lb)
                                  total wt.
                                          lb/hr
    ______________________________________
     1    60.00   23.40    -79.72 .4392   104639.19
     2-17
          60.00   74.61    -79.72 .4392   52073.66
     2-20
          60.00   74.61    -79.72 .4392   52565.53
     2    60.00   74.61    -79.72 .4392   104639.19
     3-17
         115.87   74.31    -16.82 .4392   52073.66
     3-20
         115.87   74.31    -16.82 .4392   52565.53
     3   115.87   74.31    -16.82 .4392   104639.19
     3-11
         115.87   74.31    -16.82 .4392   26111.02
     3-12
         115.87   74.31    -16.82 .4392   37736.67
     3-16
         115.87   74.31    -16.82 .4392   40791.51
     4-11
         134.02   74.11    45.97  .4392   26111.02
     4-12
         134.02   74.11    45.97  .4392   37736.67
     4-16
         134.02   74.11    45.97  .4392   40791.51
     4   134.02   74.11    45.97  .4392   104639.19
     5   148.23   73.91    104.42 .4392   104639.19
     6   148.23   73.91    625.12 .9688   13821.00
     7   148.23   73.91    25.19  .3586   90818.19
     8   148.23   73.91    25.19  .3586    9197.34
     9   148.23   73.91    385.41 .7250   23018.34
    10   148.23   73.91    25.19  .3586   81620.85
    11   123.01   73.71    314.18 .7250   23018.34
    12   122.52   73.91    -3.84  .3586   81620.85
    13   101.31   73.61    245.97 .7250   23018.34
    14    60.00   73.51    -48.36 .7250   23018.34
    15   148.23   23.90    548.21 .7250   23018.34
    16   122.01   23.70    436.94 .7250   23018.34
    17    75.00   23.60    294.63 .7250   23018.34
    18    84.37   23.60    30.22  .4392   104639.19
    19    86.01   23.60    -44.35 .3586   81620.85
    20    86.71   73.91    -44.35 .3586   81620.85
    21    60.00   1574.00  -48.36 .7250   23018.34
    22    119.01  1573.00  19.85  .7250   23018.34
    23-14
          55.00   --       --     WATER   741492.81
    23-1  55.00   --       --     WATER   485596.48
    23    55.00   --       --     WATER   1227089.29
    24-13
          64.14   --       --     WATER   741492.81
    24-18
          78.69   --       --     WATER   485596.48
    24    69.90   --       --     WATER   1227089.29
    25   1040.00  --       235.95 GAS     125248.00
    26   152.82   --       13.26  GAS     125248.00
    30   990.00   1570.00  1231.52
                                  .7250   23018.34
    31   918.46   1090.00  1187.99
                                  .7250   23018.34
    32   841.93   734.00   1141.40
                                  .7250   23018.34
    33   756.84   470.00   1090.03
                                  .7250   23018.34
    34   664.37   288.00   1035.14
                                  .7250   23018.34
    35   565.61   168.00   978.08 .7250   23018.34
    36   453.43   87.00    915.46 .7250   23018.34
    37   367.12   50.00    868.77 .7250   23018.34
    38   262.47   24.10    813.91 .7250   23018.34
    ______________________________________

The above cycle had an output of 2595.78 KWe with a cycle efficiency of 31.78%.

In the second case study, an illustrative power cycle in accordance with the present invention was added to the apparatus which was the subject of the aforementioned case study. The same pressure in the boiler, the same composition of working fluid, and the same temperature of cooling water were employed. The parameters for the theoretical calculations which were performed again utilizing standard ammonia-water and enthalpy/concentration diagrams are set out in Table 2 below. In Table 2 below, points 1-21 correspond with the specifically marked points in FIG. 2. Points 23-55 correspond with the specifically marked points in FIG. 1 herein.

In relation to this second case study, the following data was calculated:

                  TABLE 2
    ______________________________________
                                   NH.sub.4 Con-
                                   centration
    Point
         Temp.    Press.   Enthalpy
                                   lbs NH.sub.4 /
                                           W
    No.  (.degree.F.)
                  (psia)   (BTU/lb)
                                   total wt.
                                           lb/hr
    ______________________________________
     1   60.00    25.60    -79.85  .4536   105580.76
     2-17
         60.00    74.61    -79.85  .4536   50589.80
     2-20
         60.00    74.61    -79.85  .4536   54990.97
     2   60.00    74.61    -79.85  .4536   105580.76
     3-17
         111.28   74.31    -22.07  .4536   50589.80
     3-20
         111.28   74.31    -22.07  .4536   54990.97
     3   111.28   74.31    -22.07  .4536   105580.76
     3-11
         111.28   74.31    -22.07  .4536   28091.82
     3-12
         111.28   74.31    -22.07  .4536   40205.78
     3-16
         111.28   74.31    -22.07  .4536   37283.16
     4-11
         127.49   74.11    33.90   .4536   28091.82
     4-12
         127.49   74.11    33.90   .4536   40205.78
     4-16
         127.49   74.11    33.90   .4536   37283.16
     4   127.49   74.11    33.90   .4536   105580.76
     5   142.00   73.91    93.93   .4536   105580.76
     6   142.00   73.91    618.89  .9741   13639.05
     7   142.00   73.91    16.07   .3764   91941.71
     8   142.00   73.91    16.07   .3764   9745.95
     9   142.00   73.91    367.65  .7250   23385.00
    10   142.00   73.91    16.07   .3764   82195.76
    11   118.33   73.71    300.43  .7250   23385.00
    12   117.83   73.91    -11.31  .3764   82195.76
    13   99.03    73.61    237.69  .7250   23385.00
    14   60.00    73.51    -48.36  .7250   23385.00
    15   142.00   26.10    500.68  .7250   23385.00
    16   117.49   25.90    411.45  .7250   23385.00
    17   75.00    25.80    286.44  .7250   23385.00
    18   82.86    25.80    24.54   0.4536  105,580.76
    19   83.66    25.80    -49.97  0.3764  82,195.76
    20   83.66    73.91    -49.97  0.3764  82,195.76
    21   60.00    75.40    -48.36  0.7250  23,385.00
    22   114.33   1,574.40 14.38   0.7250  23,385.00
    23-14
         55.00    --       --      WATER   --
    23-1 55.00    --       --      WATER   --
    23   55.00    --       --      WATER   --
    24-13
         63.88    --       --      WATER   --
    24-18
         76.79    --       --      WATER   --
    24   69.07    --       --      WATER   --
    25   1,040.00 --       235.95  GAS     125,248.00
    26   147.30   --       11.85   --      125,248.00
    30   990.00   1,570.00 1,231.518
                                   0.725   23,385.00
    31   925.50   1,140.00 1,192.105
                                   0.725   23,385.00
    32   848.91   768.00   1,145.497
                                   0.725   23,385.00
    33   769.84   510.00   1,097.707
                                   0.725   23,385.00
    34   896.96   330.00   1,182.850
                                   0.725   23,385.00
    35   803.24   210.00   1,123.792
                                   0.725   23,385.00
    36   708.98   130.00   1,065.948
                                   0.725   23,385.00
    37   602.31   72.40    1,002.486
                                   0.725   23,385.00
    38   181.56   26.30    771.740 0.725   23,385.00
    40   769.84   510.00   1,097.707
                                   0.725   23,385.00
    41   990.00   509.00   1,243.062
                                   0.725   23,385.00
    42   602.31   72.40    1,002.486
                                   0.725   23,385.00
    43   318.15   71.40    840.260 0.725   23,385.00
    44   293.55   1,570.00 233.915 0.725   23,385.00
    45   293.55   1,570.00 233.915 0.725   5,448.71
    46   293.55   1,570.00 233.915 0.725   17,936.30
    47   562.00   1,570.00 930.164 0.725   5,448.71
    48   562.00   1,570.00 930.164 0.725   17,936.30
    49   562.00   1,570.00 930.164 0.725   23,385.00
    50   1,040.00 --       235.950 GAS     --
    51   1,040.00 --       235.950 GAS     --
    52   618.65   --       130.184 GAS     --
    53   809.00   --       177.962 GAS     --
    54   707.73   --       152.545 GAS     --
    55   310.50   --        52.838 GAS     --

Abstract --- A method and apparatus for implementing a thermodynamic cycle, which includes the use of a composite stream, having a higher content of a high-boiling component than a working stream, to provide heat needed to evaporate the working stream. After being superheated, the working stream is expanded in a turbine. Thereafter, the expanded stream is separated into a spent stream and a withdrawal stream. The withdrawal stream is combined with a lean stream to produce a composite stream. The composite stream evaporates the working stream and preheats the working stream and the lean stream. The composite stream is then expanded to a reduced pressure. A first portion of this composite stream is fed into a gravity separator. The liquid stream flowing from the gravity separator forms a portion of the lean stream that is combined with the withdrawal stream. The vapor stream flowing from the separator combines with a second portion of the composite stream in a scrubber. The vapor stream from the scrubber combines with a third portion of the expanded composite stream to produce a pre-condensed working stream that is condensed forming a liquid working stream. The liquid streams from the scrubber and gravity separator combine to form the lean stream. The liquid working stream is preheated and evaporated transforming it into the gaseous working stream. The cycle is complete when the gaseous working stream is again superheated.

US Cl. 60/673 ; 60/649
Intl. Cl. F01K 025/06

References Cited
U.S. Patent Documents
USP # 4,548,043 ~ Oct., 1985 ~ Kalina ~ 60/673
USP # 4,604,867 ~ Aug., 1986 ~ Kalina ~ 60/649

Description

BACKGROUND OF THE INVENTION

1. Field of the Invention

This invention relates generally to methods and apparatus for transforming thermal energy from a heat source into mechanical and then electrical form using a working fluid that is expanded and regenerated. This invention further relates to a method and apparatus for improving the thermal efficiency of a thermodynamic cycle.

2. Brief Description of the Background Art

It is well known that, in accordance with the Second Law of thermodynamics, the exergy (energy potential) of any heat source is increased as the temperature of this heat source is increased. Because of this effect, technological improvements in power generation have been directed toward increasing the temperature of the heat released in the process of combustion. One such improvement is the counterflow preheating of the combustion air with combustion gases to increase the combustion temperature and the average temperature of heat released from the burning of fuel. This technique, referred to as "pulverized-coal combustion," is well known and widely established.

Unlike the energy potential of the heat source, the efficiency of a power cycle depends, not on the temperature of the heat source directly, but on the average temperature of the working fluid in the process of heat transfer from the heat source. If this temperature of heat acquisition is significantly lower than the temperature of the available heat source, irreversible losses of exergy occur in the process of heat transfer, and the efficiency of the cycle remains relatively low.

This effect explains the relatively low efficiency of conventional power plants. For example, the limit of efficiency of a power plant converting thermal energy into power is on the level of approximately 63% , even when the working fluid temperature is maintained at the 1,000.degree. to 1,100.degree. F. limit that the metallurgical properties of modern power plants dictate. Similarly, the efficiency of the best direct-fired plants, based on a turbine electrical-power output (from which the work of the circulating feed pumps is subtracted) does not exceed 41-42%. In other words, the thermodynamic efficiency of these plants does not exceed 65% (the ratio of the thermal efficiency to the thermodynamic limit of efficiency).

The theoretical reason for this phenomenon is that the bulk of the heat transferred to the working fluid, i.e., water, is acquired in the boiler, where water boils at a temperature of approximately 660.degree. F. (350.degree. C.), while the available heat has a much higher temperature. It is absolutely clear, from a thermodynamic point of view, that unless the temperature of the heat acquisition by the working fluid is increased drastically, the efficiency of the process of conversion of thermal energy into power, i.e., the efficiency of the thermodynamic cycle, cannot be increased.

Use of a working fluid with a boiling temperature higher than that of water would not as a practical matter improve efficiency of the cycle for the following reason. The pressure in the condenser must be maintained at deep vacuum, even when water is used as a working fluid. If fluid with a normally higher-than-water boiling temperature is used, an even deeper vacuum in the condenser would be required, which would be technically impractical. Unless this super-low pressure in the condenser was provided, the temperature of condensation of such a hypothetical high-boiling fluid would be high, and all the gains obtained in the boiler would be lost in the condenser. Because of this problem, very little progress has been made in improving the efficiency of direct-fired power plants in the last sixty to seventy years.

A promising way to increase the efficiency of a power cycle utilizing high-temperature heat sources would be to use the so-called "recuperative cycle". According to this idea, the working fluid should be preheated to a relatively high temperature by the returning streams of the same working fluid. Only after such preheating should the external heat be transferred to the working fluid. As a result, all heat acquisition would occur at a high temperature, and theoretically the efficiency of such a cycle would be increased.

The only practical example of such a cycle is the so-called "recuperative Brighton Cycle", which utilizes a gaseous working fluid. In this cycle, the working fluid is compressed at ambient temperature, preheated in a recuperator, additionally heated by a heat source, expanded in a turbine, and sent back into the recuperator, thus providing preheating.

Despite its theoretical advantages, the recuperative Brighton Cycle does not, in reality, provide a superior efficiency because of two factors:

(1) the "work of compression" of a gaseous working fluid is very high and cannot be performed isothermally or with a small rise in temperature; and

(2) because a gaseous working fluid is used, the temperature difference in the recuperator must be relatively high, thus causing irreversible exergy losses.

The ideal solution to a high-efficiency power cycle would be to combine a high degree of recuperation, characteristic of the Brighton Cycle, with a steam cycle wherein the working fluid pressure is increased while this fluid is in a liquid state. This allows the use of pumps, with a relatively minor work requirement (low "work of compression") to increase fluid pressure.

The direct realization of such a cycle unfortunately appears impossible, for a very simple reason. If the process of recuperative heating includes liquid preheating, evaporation, and some superheating, then the returning stream, which must have a lower pressure than the oncoming stream, would condense at a lower temperature than that at which the oncoming stream boils. This phenomenon appears to make the direct recuperation of heat in such a process impossible.

As indicated above, the overall boiling process in a thermodynamic cycle can be viewed for discussion purposes as consisting of three distinct parts: preheating, evaporation, and superheating. With conventional technology, the matching of a heat source and the working fluid is adequate only during the high temperature portion of superheating. The inventor of the present invention has appreciated, however, that in previously known processes a portion of the high temperature heat which would be suitable for high temperature superheating is used instead for evaporation and preheating. This causes very large temperature differences between the two streams, and as a result, irreversible losses of exergy. For example, in the conventional Rankine cycle, the losses arising from mismatching of the enthalpy-temperature characteristics of the heat source and the working fluid would constitute about 25% of the available exergy.

The ideal solution to the age old dilemma of poorly matched heat source and working fluid enthalpy-temperature characteristics would be one that makes high temperature heat available from the heat source for use in superheating thereby reducing the temperature differences during superheating, but at the same time provides lower temperature heat which minimizes the temperature differences in the process of evaporation.

Conventional steam-power systems provide a poor substitute for this ideal system. This is because the heat provided by the multiple withdrawal of steam, that has been partially expanded in a turbine, may only be used for the low temperature pre-heating of the incoming or feed water stream to the turbine. This use of the multiple withdrawal of steam to provide heat to the feedwater is known as feedwater preheating. Unlike its use in low temperature pre-heating, the withdrawal of partially expanded steam can not provide heat for the high temperature protion of the preheating process or for the evaporation of or for the low temperature portion of the superheating of the feedwater stream.

Because of technological limitations, the water usually boils at a pressure of approximately 2,500 psia and at a temperature of about 670.degree. F. Thus, the temperature of the heat source of these systems is generally substantially greater than the boiling temperature of the liquid working fluid. Because of the difference between the high temperature of the combustion gases and the relatively low boiling temperature of the working fluid, conventional steam systems use high-temperature heat predominantly for low-temperature purposes. Since the difference between the temperature of the available heat and the temperature required for the process is very large, very high thermodynamic losses result from an irreversible heat exchange. Such losses severely limit the efficiency of conventional steam systems.

Replacing conventional systems with a system that provides lower temperature heat for evaporation of the working fluid may substantially reduce thermodynamic losses resulting from evaporation. Reducing these losses can substantially increase the efficiency of the system.

SUMMARY OF THE INVENTION

It is one feature of the present invention to provide a significant improvement in the efficiency of a thermodynamic cycle by permitting closer matching of the working fluid and the heat source enthalpy-temperature characteristics in the boiler. It is also a feature of the present invention to provide a direct fired power cycle in which high temperature heat added to the cycle may be used predominately, if not entirely, for high temperature purposes.

This transfer of heat to a working fluid predominately or solely at relatively high temperatures creates the necessary conditions at which to achieve a high thermodynamic and thermal efficiency. Because the working fluid in this cycle is a mixture of at least two components, the cycle enables a large percentage of recuperative heat exchange, including recuperative preheating, recuperative boiling and partial recuperative superheating, to be achieved. Such recuperative boiling, although impossible in a single component system, is possible in this multicomponent working fluid cycle. Unlike a single component system, when two or more components are used, different compositions for the working fluid may be used in different locations in the cycle. This enables a returning stream of working fluid, having a lower pressure than an oncoming stream, to condense within a temperature range which is higher than the temperature range within which the oncoming stream boils, thus effecting recuperative boiling of the working fluid.

In accordance with one embodiment of the present invention, a method of implementing a thermodynamic cycle includes the step of expanding a gaseous working stream to transform its energy into a useable form. The expanded gaseous working stream is divided into a withdrawal stream and a spent stream. After dividing the expanded stream into the two streams, the withdrawal stream is combined with a lean stream, having a higher content of a high-boiling component than is contained in the withdrawal stream, to form a composite stream that condenses over a temperature range that is higher than the temperature range required to evaporate an oncoming liquid working stream.

After forming the composite stream, that stream is transported to a boiler where it is condensed to provide heat for the boiling of the oncoming liquid working stream. Evaporation of the liquid working stream produces the above mentioned gaseous working stream. Subsequently, the composite stream is separated to form a liquid stream and a vapor stream. Some or all of the liquid stream forms the above mentioned lean stream. The vapor stream is returned into the cycle, preferably by being combined with a portion of the composite stream to produce a pre-condensed working stream. The pre-condensed working stream is condensed to produce the liquid working stream that is transported to the boiler. The spent stream may be combined with this liquid working stream prior to the liquid working stream being sent to the boiler. Alternatively, the spent stream may be returned to the system at some other location. To complete the cycle, the heat, that the above mentioned composite stream transports to the boiler, is used to evaporate the liquid working stream to form the gaseous working stream.

In accordance with another embodiment of the present invention, the gaseous working stream, exiting from the boiler, may then by superheated in one or more heat exchangers by either the withdrawal stream or the spent stream or by both the withdrawal and spent streams. Following the superheating of the gaseous working stream in the heat exchangers, the gaseous working stream may be further superheated in a heater. The energy supplied to the heater is supplied from outside the thermodynamic cycle. After this superheating, expansion of the gaseous working stream takes place. This expanded gaseous working stream may be reheated and expanded one or more times before being divided into the spent and withdrawal streams. This embodiment may further include the step of reheating and expanding the spent stream one or more times after the spent stream has been separated from the withdrawal stream.

In addition, this embodiment may further include a series of recuperative heat exchangers used to recuperate heat from the withdrawal, composite and spent streams. These heat exchangers may allow the lean stream and the liquid working stream to absorb heat from the composite stream. Further, one or more of these heat exchangers may allow the spent stream to provide additional heat to the liquid working stream to aid in the preheating and boiling of the liquid working stream.

In accordance with yet another embodiment of the present invention, the methods for implementing a thermodynamic cycle described above may further include the step of reducing the pressure of the composite stream with a hydraulic turbine (or alternatively a throttle valve). After this reduction of pressure, a first portion of this composite stream may be partially evaporated in one or more heat exchangers with heat from the spent stream and with heat from this same composite stream as it flows toward the turbine. After the partial evaporation of this first portion of the composite stream, it is sent to a separator where it is separated into a vapor stream and a liquid stream.

In this embodiment, the liquid stream forms a portion of the lean stream which may be sent to a circulation pump to be pumped to a higher pressure. The circulation pump may be connected to the hydraulic turbine; the hydraulic turbine releasing energy used to operate the pump. After attaining this high pressure, the lean stream may be heated by the returning composite stream in one or more heat exchangers. After acquiring this additional heat, the lean stream is combined with the withdrawal stream to form the composite stream used to preheat and evaporate the liquid working stream.

The vapor stream may be combined with a second portion of the composite stream, that flows from the hydraulic turbine, in a direct contact heat exchanger or in a scrubber. The liquid stream flowing from the heat exchanger or scrubber may combine with the liquid stream from the separator to produce the lean stream. The vapor stream flowing from the heat exchanger or scrubber forms a super rich stream. In this embodiment, this super rich stream may be combined with a third portion of the composite stream, that flows from the hydraulic turbine, to form a pre-condensed working stream. This stream may then pass through a heat exchanger, to supply heat to the returning liquid working stream, before it is fed into a water-cooled condenser to be fully condensed to produce the liquid working stream.

The liquid working stream may be pumped to a high pressure by a feed pump. After obtaining this high pressure, the liquid working stream may be heated in a series of heat exchangers by the pre-condensed working stream, returning composite stream and the returning spent stream. This heat exchange, which may be accompanied by the pumping of the liquid working stream to progressively higher pressures, continues until the liquid working stream is evaporated to produce the gaseous working stream, thereby completing the cycle.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a schematic representation of one embodiment of the method and apparatus of the present invention.

FIG. 2 is a schematic representation of a second embodiment of the method and apparatus of the present invention.

DESCRIPTION OF PREFERRED EMBODIMENTS

The schematic shown in FIG. 1 shows an embodiment of preferred apparatus that may be used in the above described cycle. Specifically, FIG. 1 shows a system 100 that includes a boiler in the form of heat exchangers 112 and 127, a preheater in the form of heat exchangers 114 and 116, and a superheater in the form of heat exchangers 109 and 110. In addition, the system 100 includes turbines 102, 104 and 106, superheater 101, reheaters 103 and 105, gravity separator 120, scrubber 125, hydraulic turbine 119, pumps 122, 123, 138 and 139, heat exchangers 117, 118 and 128, and condenser 121. Further, the system 100 includes stream separators 131-137 and stream mixers 140-147.

The condenser 121 may be any type of known heat rejection device. For example, the condenser 121 may take the form of a heat exchanger, such as a water cooled system, or another type of condensing device. In the alternative, condenser 121 may be replaced with the heat rejection system described in U.S. Pat. Nos. 4,489,563 and 4,604,867 to Kalina. The Kalina system requires that the stream shown approaching condenser 121 in FIG. 1 be mixed with a multi-component fluid stream, for example, a fluid stream comprised of water and ammonia, condensed and then distilled to produce the original state of the working fluid. Thus, when the heat rejection system of the Kalina cycle is used in place of condenser 121, the distillation subsystem described in U.S. Pat. Nos. 4,489,563 and 4,604,867 may be utilized in place of condenser 121. U.S. Pat. Nos. 4,489,563 and 4,604,867 are hereby expressly incorporated by reference herein.

Various types of heat sources may be used to drive the cycle of this invention. Thus, for example, heat sources with temperatures as high as 1,000.degree. C. or more down to heat sources sufficient to superheat a gaseous working stream may be used to heat the gaseous working stream flowing through heater 101 and reheaters 103 and 105. The combustion gases resulting from the burning of fossil fuels is a preferred heat source. Any other heat source capable of superheating the gaseous working stream that is used in the described embodiment of the invention may also be used.

While the embodiment illustrated in FIG. 1 is related to pulverized coal combustion, this system may be used with a variety of combustion systems, including different types of fluidized bed combustion systems and waste incineration systems. One of ordinary skill can adjust the system by adding heat exchangers needed to accommodate a variety of different combustion systems.

The working fluid used in the system 100 may be any multi-component working fluid that comprises a lower boiling point fluid and a relatively higher boiling point fluid. Thus, for example, the working fluid employed may be an ammonia-water mixture, two or more hydrocarbons, two or more freons, mixtures of hydrocarbons and freons or the like. In general, the fluid may be mixtures of any number of compounds with favorable thermodynamic characteristics and solubility. In a preferred embodiment, a mixture of water and ammonia is used.

As shown in FIG. 1, a working stream circulates through system 100. The working stream includes a gaseous working stream that flows from stream mixer 142 until it is separated into a withdrawal stream and a spent stream at separator 131. In addition to the gaseous working stream, the withdrawal stream (that flows from separator 131 to stream mixer 141) and the spent stream (that flows from separator 131 to stream mixer 147) the working stream includes a pre-condensed working stream (that flows from mixer 146 to condenser 121) and a liquid working stream (that flows from condenser 121 to boilers 112, 127). Each portion of the working stream contains the same percentage of high boiling and low boiling components.

The gaseous working stream, that has been completely evaporated and superheated in previous stages of system 100, enters heater 101. While in heater 101, the gaseous working stream is superheated to the highest temperature that is reached at any stage in the process. After being superheated, this gaseous working stream is expanded in turbine 102 to an intermediate pressure. This expansion allows the heat contained in the gaseous working stream to be converted into energy that is in a useable form.

After expansion in turbine 102, the gaseous working stream is separated by separator 131 into two streams, a withdrawal stream and a spent stream. The spent stream is reheated in reheater 103, expanded in turbine 104, reheated a second time in reheater 105 and expanded a second time in turbine 106. Although FIG. 1 shows the system 100 as having two reheaters 103 and 105, for reheating the spent stream, and two turbines 104 and 106, for expandiang the spent stream, the optimum number of reheaters and turbines depends upon the desired efficiency of the system. The number of reheaters and turbines may be either increased or decreased from the number shown in FIG. 1. In addition, a single heater may be used to heat the gaseous working stream, prior to expansion, and the spent working stream, prior to the expansion of the spent stream. Therefore, the number of heaters and reheaters may be more than, less than or equal to the number of turbines.

Further, system 100 may include additional heaters and turbines for reheating and expanding the gaseous stream exiting from turbine 102 prior to that stream's separation into the withdrawal and spent streams. Thus, although the inclusion of reheaters 103 and 105 and turbines 104 and 106 to system 100 provides a preferred embodiment of the present invention, one may select a different number of reheaters and turbines without departing from the scope of the disclosed general inventive concept.

After these reheatings and expansions of the spent stream, the stream passes through a series of recuperative heat exchangers. As shown in FIG. 1, the spent stream, after expansion, passes through recuperative heat exchangers 110, 127 and 116. While passing through heat exchanger 110, the spent stream provides heat to superheat the gaseous working stream. While passing through heat exchanger 127, the spent stream provides heat to evaporate the oncoming high-pressure liquid working stream. Similarly, while passing through heat exchanger 116, the spent stream provides heat to preheat this oncoming high pressure liquid working stream.

Whether any or all of the heat exchangers 110, 127 and 116 are used or whether a number of additional heat exchangers are added to the system is a matter of design choice. Although the inclusion of heat exchangers 110, 127 and 116 to system 100 is preferred, the spent stream may pass through an increased number of heat exchangers, or not pass through any heat exchangers at all, without departing from the scope of the disclosed invention.

The withdrawal stream beginning at stream separator 131 initially passes through recuperative heat exchanger 109. While passing through heat exchanger 109, the withdrawal stream provides heat for the superheating of the oncoming high-pressure gaseous working stream. Although system 100 preferably includes heat exchanger 109, one may remove heat exchanger 109 or add additional heat exchangers. The preferred state of the withdrawal stream at point 42, after it has passed through heat exchanger 109, is that of a superheated vapor.

After heating the gaseous working stream, the withdrawal stream combines with a lean stream at stream mixer 141. This lean stream contains the same components as are contained in the working stream. The lean stream, however, contains a higher content of a high-boiling component than is contained in any part of the working stream. For example, if ammonia and water are the two components present in the working and lean streams, the water is the high-boiling component and the ammonia is the low-boiling component. In such a two component system, the lean stream contains a higher percentage of water than is contained in the working stream. As shown in FIG. 1, the lean stream flows from stream mixer 144 to stream mixer 141.

In this embodiment, the state of the lean stream at point 74, prior to mixing with the withdrawal stream at stream mixer 141, is preferably that of a subcooled liquid.

Mixing the lean stream with the withdrawal stream at stream mixer 141 provides a composite stream that has a lower boiling temperature range than the lean stream but a higher boiling temperature range than the withdrawal stream or any other portion of the working stream. The state of the composite stream as it flows from stream mixer 141 depends upon the states of the lean and withdrawal streams. It is preferably that of a vapor-liquid mixture. Preferably, the pressure of the withdrawal stream at point 42 and the lean stream at point 74, prior to mixing at stream mixer 141, will be the same as the pressure of the composite stream at point 50, that is formed at stream mixer 141. The temperature of the composite stream at this point is preferably higher than the temperature of the lean stream at point 74 and slightly lower than that of the withdrawal stream at point 42.

The composite stream will contain a higher percentage of a high-boiling component than is contained in the withdrawal stream or in other portions of the working stream. Because the composite stream contains a higher percentage of a high-boiling component, it may be condensed within a temperature range which exceeds the boiling temperature range of the liquid working stream. Further, in this preferred embodiment, the composite stream may be condensed at a higher temperature than the boiling temperature of the liquid working stream, even if the pressure of the composite stream is significantly lower than the pressure of the oncoming liquid working stream.

The composite stream produced by the mixing of the withdrawal stream with the lean stream flows into heat exchanger 112, where it is cooled and condensed. As it is being cooled and condensed, the composite stream provides heat to evaporate the oncoming liquid working stream and to provide heat to the oncoming lean stream, as those streams enter heat exchanger 112.

Using a composite stream, having a higher boiling temperature range than the boiling temperature range of the liquid working stream, provides one of the principle distinctions between the thermodynamic cycle disclosed in the present invention and conventionally used cycles. Unlike a conventional thermodynamic cycle, the cycle of the present invention withdraws part of the gaseous working stream, after it has been partially expanded, to provide heat for a composite stream comprising that withdrawn part of the gaseous working stream together with a lower temperature lean stream. This composite stream, preferably having a pressure that is lower than the pressure of the oncoming liquid working stream, is used to heat and completely or partially evaporate the oncoming liquid stream.

Because of the higher percentage of a high-boiling component contained in this composite stream, the composite stream condenses over a range of temperatures that are higher than the temperatures required to evaporate the oncoming liquid working stream, even though the liquid working stream may enter heat exchanger 112 at a higher pressure than the pressure of the composite stream.

Such a method of evaporating a liquid working stream can not be performed in conventional steam-power systems. In conventional systems, the condensation of the withdrawn stream must occur over a lower temperature range than the boiling temperature of the oncoming liquid working stream, if the withdrawn stream has a lower pressure than the pressure of the oncoming liquid working stream. Thus, heat released by condensation of a withdrawn stream in conventional systems can be used only for partial preheating of the oncoming working stream.

In contrast, in the method disclosed by the present invention, the presence of a higher percentage of a high-boiling component in the composite stream allows that stream to condense over a higher temperature range than the boiling temperature range of the oncoming liquid working stream, even if the pressure of the composite stream is substantially lower than the pressure of the liquid working stream. It should be appreciated that the described method uses a single withdrawal stream to form a composite stream that acts as the heat source effecting the complete preheating and evaporation of the working stream and also provides heat for the low temperature superheating of the working stream.

To create this composite stream, however, part of the expanded gaseous working stream must be withdrawn. It should be appreciated that withdrawing part of this superheated stream for combination with a lean stream to produce the composite stream results in thermodynamic losses because of the reduction in temperature of the withdrawn stream. The losses resulting from the removal of part of the gaseous stream and mixing that withdrawal stream with a lean stream are, however, more than compensated for by the losses that are prevented when the composite stream is used to evaporate the liquid working stream.

As the calculations in Table II show, using a portion of the expanded gaseous working stream to create a composite stream, having a higher percentage of a high-boiling component than is contained in the liquid working stream, allows the thermodynamic cycle of the present invention to have a substantially increased efficiency compared to conventional steam-power systems. Using this composite stream to provide low temperature heat for the low temperature evaporation process allows the available heat in the system to be more adequately matched with the liquid working stream's enthalpy-temperature characteristics. This matching prevents the very high thermodynamic losses that occur in conventional systems that use high temperature heat in low temperature evaportion processes. The enormous amount of exergy saved by using this composite stream to more closely match the temperature of the heat source with the liquid working stream's enthalpy-temperature characteristics substantially exceeds any losses caused from removing part of the gaseous working stream from its superheated state.

The pressure at which the withdrawal stream is mixed with the lean stream to produce the composite stream must be a pressure which insures that the temperature over which the composite stream condenses will be higher than the temperature over which the liquid working stream evaporates. The leaner the composite stream, the lower will be the pressure needed for condensation. The lower the pressure, the larger the expansion ratio of turbine 102, corresponding to an increase in the work that this turbine provides.

There is a practical limit to the amount of the high boiling component that can be used in the composite stream. This is because a leaner composite stream is more difficult to separate. Thus, to optimize the system's efficiency, the choice of pressure and composition for the composite stream must be carefully made. Table I provides one example of a composite stream pressure and composition that may be used to provide a highly efficient cycle.

It should be appreciated that heat exchanger 127, wherein the spent stream is used to evaporate part of the liquid working stream, may be removed from system 100 without departing from the scope of the described general inventive concept. The portion of the liquid working stream that had passed through heat exchanger 127 would then be diverted to heat exchanger 112, where it would be evaporated.

After passing through heat exchanger 112, the composite stream is sent into heat exchanger 114 to provide heat for preheating the lean stream and the liquid working stream. As the composite stream transfers heat to the lean stream and the liquid working stream, the composite stream is further cooled. Again, although limiting the number of heat exchangers in this part of system 100 to heat exchangers 112 and 114 is preferred, additional heat exchangers may be added or heat exchanger 114 may be removed from the system 100 without departing from the scope of the disclosed invention.

After the composite stream exits from heat exchanger 114, it is sent into heat exchanger 117, where its heat is used to partially evaporate a countercurrent portion of that same composite stream that flows from separator 135.

Even after exiting heat exchanger 117, the pressure of the composite stream at point 53, in this embodiment of the present invention, remains relatively high. Since the composite stream may not be able to produce the working stream and lean stream at this high pressure, this pressure may have to be reduced. This reduction in pressure occurs in the hydraulic turbine 119. A particular hydraulic turbine that may be used in a Pelton wheel.

During this pressure reduction step, all or part of the work needed to pump the lean solution at pump 122 may be recovered. Because the weight flow rate of the stream passing through Pelton wheel 119 is higher than the weight flow rate of the lean stream passing through pump 122, the energy released in Pelton wheel 119 is usually sufficient to provide the work of pump 122. If the energy that Pelton wheel 119 releases is insufficient, a supplementary electrical motor can be installed to supply the additional power that pump 122 requires.

A throttle valve may be used as an alternative to hydraulic turbine 119. If a throttle valve is used instead of the hydraulic turbine, work spent to pump the lean solution will, of course, not be recovered. Regardless of whether hydraulic turbine 119 or a throttle valve is used, however, the remainder of the process will not be affected. The choice of whether to use a hydraulic turbine or a throttle valve to reduce the pressure of the composite stream is strictly an economic one. Further, although the use of heat exchanger 117 and turbine 119 is preferred, one may decide not to use these devices, or may decide to add additional heat exchangers or other pressure reduction apparatus to the system 100.

The composite stream flowing from hydraulic turbine 119 preferably has a pressure at point 56 that is approximately equal to or slightly greater than the pressure of condensation. A portion of this composite stream, having this reduced pressure, is separated from the composite stream at separator 137. This stream is again divided at separator 136. A first portion of the composite stream separated at separator 136 is then split into two streams at separator 135. These two streams are then sent into heat exchangers 117 and 118, where the counterstream of the same composite stream is cooled and the returning spent stream is condensed, partially evaporating these two streams. The countercurrent composite stream adds heat in heat exchanger 117 and the condensing spent stream adds heat in heat exchanger 118. After exiting exchangers 117 and 118, the two streams flowing from separator 135 are combined at stream mixer 145. This partially evaporated stream is then sent to gravity separator 120.

The state of the stream entering gravity separator 120 is that of a vapor-liquid mixture. In order to provide heat for this partial evaporation, the spent stream, which had been condensed in heat exchanger 118, must have a pressure which will enable the spent stream to be condensed at an average temperature which is higher than the average temperature needed to evaporate the portion of the composite stream that is to be separated. The leaner the composite stream, the higher the temperature necessary for its evaporation, and thus the higher the pressure of the spent stream at point 37. Increasing the pressure at point 37 reduces the expansion ratio in turbines 104 and 106 and, as a result, reduces the work output of these turbines. This shows that, although making the composite stream leaner increases the power output of turbine 102, it reduces the power output of turbines 104 and 106.

To maximize the total output of all three turbines, an appropriate composition must be selected for the composite stream. One such composition is provided in Table I.

The embodiment shown in FIG. 1 uses the returned spent stream to preheat the liquid working stream and to partially evaporate the stream sent to gravity separator 120. At the same time, the spent stream is condensed as it passes through heat exchanger 118. It should be noted that, instead of condensing the spent stream in condenser 121, without simultaneously recovering heat from that condensing stream, system 100 uses the heat that the spent stream releases as it is being condensed in heat exchanger 118 to preheat the liquid working stream and partially evaporate the composite stream sent to separator 120.

Gravity separator 120 separates the first portion of the composite stream into a vapor stream and a liquid stream. The liquid stream flowing from the bottom of gravity separator 120 forms a portion of the lean stream that is mixed with the previously described withdrawal stream at mixer 141.

The vapor stream flowing from gravity separator 120 is sent to the bottom of scrubber 125. A second portion of the composite stream, flowing from separator 136, is sent into the top of scrubber 125. The liquid and vapor streams fed into scrubber 125 interact, providing heat and mass exchange. A direct contact heat exchanger or other means for effecting heat and mass exchange between the liquid and vapor streams, shown fed into scrubber 125 in FIG. 1, may be used in place of scrubber 125. Whether scrubber 125, a heat exchanger, or some other means is used in system 100 is a matter of design choice.

In the embodiment shown in FIG. 1, liquid and vapor streams exit scrubber 125. The liquid stream is combined with the liquid stream flowing from separator 120 at stream mixer 144 to form the lean stream that is mixed with the withdrawal stream at stream mixer 141 to produce the composite stream. The liquid streams flowing from scrubber 125 and separator 120 to form the lean stream preferably have the same, or nearly the same, composition.

The lean stream flows from stream mixer 144 into circulation pump 122. Pump 122 pumps the lean stream to a high pressure. In the embodiment shown in FIG. 1, the pressure of the lean stream at point 70, as it flows from pump 122, is higher than the pressure of the lean stream at point 74, as it flows from heat exchanger 112, as is shown in Table I.

As shown in FIG. 1, this high pressure lean stream passes through heat exchangers 114 and 112, where the countercurrent composite stream provides heat to the lean stream, and combines with the withdrawal stream at stream mixer 141.

The vapor stream exiting scrubber 125 is a stream having a high percentage of the lower boiling component. This super rich stream combines with a third portion of the composite stream, i.e., that portion flowing from separator 137, at stream mixer 146. This stream forms a pre-condensed working stream which flows through heat exchanger 128 and into condenser 121. While passing through heat exchanger 128, this pre-condensed working stream is further condensed while adding heat to the countercurrent liquid working stream flowing from condenser 121 and pump 123. After exiting heat exchanger 128, the pre-condensed working stream enters condenser 121, where it is fully condensed.

This pre-condensed working stream has the same composition as the above described withdrawal stream. It should be noted that only this pre-condensed working stream is condensed, minimizing the exergy losses at the condenser. As described above, the spent stream does not pass through the condenser. Instead, the heat released from the condensation of the spent stream is used to preheat the liquid working stream and to partially evaporate the composite stream sent to separator 120. The use of the spent stream in this manner ensures that the liquid working stream sent to heat exchangers 112 and 127 will be completely evaporated in a recuperative way, ensuring that system 100 will have a greater efficiency than the best conventional Rankine cycles.

Condenser 121 is preferably a water-cooled condenser. When such a condenser is used, a stream of cooling water flowing through condenser 121 completely condenses this working stream to produce the liquid working stream.

This liquid working stream flows into feed pump 123, where it is pumped to an increased pressure. This liquid working stream then flows into heat exchanger 128, where heat transferred from the pre-condensed working stream preheats the liquid working stream. After being preheated in heat exchanger 128, the liquid working stream is combined with the spent stream at stream mixer 147. This mixed stream is pumped to an intermediate pressure by pump 138. It then passes through heat exchanger 118, where it is preheated by heat transferred by the condensing returning spent stream. After exiting heat exchanger 118, the liquid working stream is pumped to a high pressure by pump 139. This high pressure, preferably subcooled, liquid working stream is then separated at separator 134 into two streams. One of the streams passes through heat exchanger 114, where heat transferred from the composite stream preheats this portion of the liquid working stream. The other stream flowing from separator 134 flows into exchanger 116, where heat from the returning spent stream is transferred to this portion of the liquid working stream, preheating this portion of the liquid working stream. The spent stream as it exits from exchanger 116 is preferably in the state of a saturated vapor, but alternatively may be in the state of a superheated vapor or may be partially condensed.

The portion of the liquid working stream passing through heat exchanger 116 is combined with the stream flowing from heat exchanger 114 at stream mixer 143. This stream is preferably in a state of a saturated, or slightly subcooled, liquid. The stream flowing from stream mixer 143 then is separated into two streams at separator 133. One stream flows into heat exchanger 112. The liquid working stream passing through heat exchanger 112 is evaporated with heat transferred from the composite stream flowing from stream mixer 141.

The other stream flowing from separator 133 then flows into heat exchanger 127, where it is evaporated with heat transferred from the spent stream.

The streams exiting heat exchangers 112 and 127 are combined at stream mixer 142. As described above, heat exchanger 127 could be removed, with all of the liquid working stream flowing from stream mixer 143 diverted to heat exchanger 112, without departing from the described general inventive concept.

In this embodiment, the stream flowing from stream mixer 142 is in the vapor state and makes up the cycle's gaseous working stream. The gaseous working stream flowing from stream mixer 142, which might even be slightly superheated, is divided into two streams at stream separator 132. One of these streams passes through heat exchanger 109, where it is superheated by the withdrawal stream passing from stream separator 131 through heat exchanger 109 to stream mixer 141. The other portion of the gaseous working stream passes through heat exchanger 110, where heat from the spent stream flowing from turbine 106 is used to superheat this portion of the gaseous working stream. The two streams flowing from stream separator 132 and through heat exchangers 109 and 110 are recombined at stream mixer 140. This recombined gaseous working stream flows into heater 101 to complete this thermodynamic cycle.

In the embodiment of system 200, shown in FIG. 2, the process of absorption, i.e., of adding the lean stream to the withdrawal stream to make the composite stream, is performed in two steps. The withdrawal stream is divided into first and second withdrawal streams at stream separator 150. The first withdrawal stream is combined with the lean stream at stream mixer 141, producing a first composite stream, which is leaner than it would be if the withdrawal stream with parameters as at point 42 was combined with the lean stream (as was done in the embodiment shown in FIG. 1).

Because the first composite stream in FIG. 2 is now leaner than the composite stream of FIG. 1, its pressure can be reduced, which will increase the work output from turbine 102. The first composite stream is then condensed in boiler 112. Thereafter, the first composite stream is combined with the second withdrawal stream at mixer 151, creating a second composite stream. The second composite stream is richer than the first composite stream. As a result, it is easier to provide for its separation.

The first composite stream provides heat for boiler 112, and enables the pressure of absorption to be reduced thus increasing the output of turbine 102. At the same time, the embodiment in FIG. 2 enables an enriched second composite stream to be sent into separator 120. This FIG. 2 embodiment thus provides the benefits of a lower pressure composite stream which does not at the same time prevent the composite stream from being easily separated.

Both the cycle shown in FIG. 1 and the cycle shown in FIG. 2 are substantially more efficient than conventional steam-power systems. The decision to use one of these preferred systems instead of the other is a matter of design choice.

In the above described thermodynamic cycles of the present invention, all of the heating and evaporating of the liquid working stream may be provided in a recuperative way, i.e., the returning composite and spent streams transfer heat to the liquid working stream as these two streams cool. Further, even part of the superheating of the gaseous working stream may be provided in this recuperative manner, i.e., the withdrawal stream and spent stream may transfer heat to the gaseous working stream as these two streams cool.

Use of a withdrawal stream to preheat an oncoming working stream is common in conventional steam-power systems. Such a practice is commonly known as "feed water heating". Feedwater heating is conventional systems is useful only for preheating the incoming liquid working stream, because the pressure and temperature of condensation of the withdrawal stream is too low for it to

---

Abstract --- A method and apparatus for implementing a thermodynamic cycle with preheating, involves expanding a gaseous working fluid to a medium pressure to transform its energy into usable form. The expanded gaseous working fluid is split into two different streams. One stream is further expanded to a spent low pressure level to produce further usable energy. This stream is then condensed. The other of the two streams is used to preheat the condensed stream and is mixed with the condensed stream at a point upstream of the point of preheating. This decreases the irreversibilities in the preheating process and enables greater efficiencies to be achieved.

References Cited
US Patent Documents

USP # 3,277,651 ~ Oct., 1966 ~Augsburger ~ 60/679
USP # 3,842,605 ~ Oct., 1974 ~ Tegtmeyer ~ 60/678
USP # 3,921,406 ~ Nov., 1975 ~ Teranishi, et al. ~ 60/678
USP # 4,003,205 ~ Jan., 1977 ~ Matsumura ~ 60/678
USP # 4,047,386 ~ Sep., 1977 ~ Frondorf ~ 60/678

Description

BACKGROUND OF THE INVENTION

1. Field of the Invention

This invention relates generally to methods and apparatus for transforming energy from a heat source into a useable form using a working fluid that is expanded and regenerated. This invention further relates to a method and apparatus for improving the heat utilization efficiency of a thermodynamic cycle.

2. Brief Description of the Background Art

In the Rankine cycle, a working fluid such as water, ammonia or freon is evaporated in an evaporator utilizing an available heat source. The evaporated gaseous working fluid is expanded across a turbine to transform its energy into useable form. The spent gaseous working fluid is then condensed in a condenser using an available cooling medium. The pressure of the condensed working medium is increased by pumping, followed by evaporation and so on to continue the cycle.

The Exergy cycle, described in U.S. Pat. No. 4,346,561, utilizes a binary or multi-component working fluid. This cycle operates generally on the principle that a binary working fluid is pumped as a liquid to a high working pressure and is heated to partially vaporize the working fluid. The fluid is then flashed to separate high and low boiling working fluids. The low boiling component is expanded through a turbine, to drive the turbine, while the high boiling component has heat recovered for use in heating the binary working fluid prior to evaporation. The high boiling component is then mixed with the spent low boiling working fluid to absorb the spent working fluid in a condenser in the presence of a cooling medium.

A theoretical comparison of the conventional Rankine cycle and the Exergy cycle demonstrates the improved efficiency of the new cycle over the Rankine cycle when an available, relatively low temperature heat source such as ocean water, geothermal energy or the like is employed.

In applicant's further invention referred to as the Basic Kalina cycle, the subject of U.S. Pat. No. 4,489,563, relatively lower temperature available heat is utilized to effect partial distillation of at least a portion of a multi-component fluid stream at an intermediate pressure to generate working fluid fractions of different compositions. The fractions are used to produce at least one main rich solution which is relatively enriched with respect to the lower boiling component, and to produce one lean solution which is relatively impoverished with respect to the lower boiling component. The pressure of the main rich solution is increased; thereafter, it is evaporated to produce a charged gaseous main working fluid. The main working fluid is expanded to a low pressure level to convert energy to useable form. The spent low pressure level working fluid is condensed in a main absorption stage by dissolving with cooling in the lean solution to regenerate an initial working fluid for reuse.

In any process of converting thermal energy to a useable form, a major loss of available energy in the heat source occurs in the process of boiling or evaporating the working fluid. This loss of available energy (known as exergy or essergy) is due to the mismatch of the enthalpy-temperature characteristics of the heat source and the working fluid in the boiler. Simply put, for any given enthalpy the temperature of the heat source is always greater than the temperature of the working fluid. Ideally, this temperature difference would be almost, but not quite, zero. This mismatch occurs both in the classical Rankine cycle, using a pure substance as a working fluid, as well as in the Kalina and Exergy cycles described above, using a mixture as a working fluid. The use of a mixture as a working fluid in the manner of the Kalina and Exergy cycles reduces these losses to a significant extent. However, it would be highly desireable to further reduce these losses in any cycle.

In the conventional Rankine cycle the losses arising from mismatching of the enthalpy-temperature characteristics of the heat source and the working fluid constitute about 25% of the available energy. With a cycle such as that described in U.S. Pat. No. 4,489,563, the loss of exergy in the boiler due to enthalpy-temperature characteristics mismatching would constitute about 14% of all of the available exergy.

The overall boiling process in a thermodynamic cycle can be viewed for discussion purposes as consisting of three distinct parts: preheating, evaporation and superheating. The quantity of heat in the temperature range suitable for superheating is generally much greater than necessary, or the quantity of heat in the temperature range suitable for evaporation is much smaller than necessary. A portion of the high temperature heat which would be suitable for high temperature superheating is used for evaporation in conventional processes. This causes very large temperature differences between the two streams, and as a result, irreversible losses of exergy.

In accordance with another invention of the applicant, the subject of U.S. Pat. No. 4,604,867, a fluid may be diverted to a reheater after initial expansion in the turbine to increase the temperature available for superheating. After return to the turbine, and additional expansion, the fluid is withdrawn from the turbine and cooled in an intercooler. Afterwards, the fluid is returned to the turbine for additional expansion. The cooling of the turbine gas may provide additional heat for evaporation. Intercooling provides compensation for the heat used in reheating and may provide recuperation of heat available which would otherwise remain unused following final turbine expansion.

In the past preheating of a working fluid is usually performed by extraction of part of the working fluid stream between turbine stages. This is followed by injection of the extracted stream or streams into the stream of feed water to the turbine. As a result heat of a lower temperature level may perform preheating, which occurs at relatively low temperature levels. Therefore, in general, this process increases the efficiency of the power plant.

However, conventional preheating has a drawback, because the steam used for preheating has a temperature which is significantly higher than the temperature of the feed water into which it is injected. This steam may even have a temperature which is higher than the temperature of the feed water obtained after injection. This creates irreversibilities and lowers the potential efficiency of the power plant.

It would be highly desirable to provide a process and apparatus which avoids the creation of these irreversibilities and thereby increases the efficiency of the power plant.

SUMMARY OF THE INVENTION

It is one feature of the present invention to provide a significant improvement in the efficiency of a thermodynamic cycle by permitting closer matching of the working fluid and heat source enthalpy-temperature characteristics during preheating. It is also a feature of the present invention to provide a system of preheating which decreases the irreversibilities and therefore increases the efficiency of the entire system.

In accordance with one embodiment of the present invention, a method of implementing a thermodynamic cycle includes the step of expanding a gaseous working fluid to transform its energy into useable form. The expanded gaseous working fluid is then split into two streams. The first stream is expanded to a spent low pressure level to transform its energy into usable form. The first stream is then condensed. The first and second streams are mixed to form a mixed stream after the second stream is used to preheat at least a portion of the mixed stream. Then the working fluid stream is evaporated to form the gaseous working fluid.

In accordance with another embodiment of the present invention, an apparatus for implementing a thermodynamic cycle includes a first turbine having a fluid inlet path and a fluid outlet path. The fluid outlet path is split into first and second lines. A second turbine is connected for fluid communication with the first line. A heat exchanger is connected for fluid communication with the second line and the first turbine. A condensing system has its output connected for fluid communication with the second turbine. A mixing chamber is connected for fluid communication with the output of the condensing system. The heat exchanger is arranged to transfer heat from fluid flowing from the first turbine to the mixing chamber to fluid flowing from the mixing chamber to the first turbine.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a schematic representation of one system for carrying out one embodiment of the method and apparatus of the present invention; and

FIG. 2 is a schematic representation of one embodiment of a distillation-condensation subsystem for use in connection with the system shown in FIG. 1.

DESCRIPTION OF A PREFERRED EMBODIMENT

Referring to the drawing wherein like reference characters are utilized for like parts throughout the several views, a system 100, shown in FIG. 1, implements a thermodynamic cycle, in accordance with one embodiment of the present invention. The illustrated system 100 includes a series of three turbines 102, 104 and 106, a condensing subsystem 108, and heat exchangers 110-124.

The condensing subsystem 108 may be any type of known heat rejection device. In the Rankine cycle, heat rejection occurs in a simple heat exchanger and thus for Rankine applications, the subsystem 108 may take the form of a heat exchanger or condenser. In the Kalina cycle, described in U.S. Pat. No. 4,489,563 to Kalina, the heat rejection system requires that gas leaving the turbine be mixed with a multi-component fluid stream, for example, comprised of water and ammonia, condensed and then distilled to produce the original state of the working fluid. Thus, when the present invention is used with a Kalina cycle, the distillation subsystem described in U.S. Pat. No. 4,489,563 may be utilized as a system 108. U.S. Pat. No. 4,489,563 is hereby expressly incorporated by reference herein.

Various types of heat sources may be used to drive the cycle of this invention. For example, heat sources with temperatures as high as, say 1000.degree. F. or more, down to the low heat sources such as those obtained from ocean thermal gradients may be utilized. Heat sources such as for example, low grade primary fuel, waste heat, geothermal heat, solar heat or ocean thermal energy conversion systems may also be implemented with the present invention. However, the present invention is particularly suitable for use with heat produced by the burning of fuel in a fluidized bed or by the burning of municipal wastes or other low grade fuel. Normally in the burning of such fuel, to avoid corrosion, the combustion gases cannot be cooled below a temperature of 300.degree. to 400°.degree. F.

A variety of working fluids may be used in conjunction with the system 100 depending on the kind of condensing subsystem 108 utilized. In conjunction with a condensing system 108 described in the U.S. patent incorporated by reference herein, any multi-component working fluid that comprises a lower boiling point fluid and a relatively higher boiling point fluid may be utilized. Thus, for example, the working fluid employed may be an ammonia-water mixture, two or more hydrocarbons, two or more freons, mixtures of hydrocarbons and freons or the like. In general, the fluid may be mixtures of any number of compounds with favorable thermodynamic characteristics and solubility. However, when implementing the conventional Rankine cycle, conventional single component working fluids such as water, ammonia, or freon may be utilized.

As shown in FIG. 1, a completely condensed working fluid which has been slightly preheated and pumped to a high pressure, exits the condensing subsystem 108 and is combined with a returning stream from the pump 126. The fluid exiting the pump 126 is at a temperature, pressure, and mass flow rate relatively close to that of the fluid exiting the condensing subsystem 108. In an illustrative embodiment the pressure of the two streams are substantially the same before they are mixed. After the two streams from the subsystem 108 and the pump 126 are combined at point 128, the working fluid is divided into two streams 130 and 132. The stream 132 is heated in the heat exchanger 122 in counterflow with the fluid in the line 134 returning from the turbine 102. The flow along the path 130 is heated by counterflow in the heat exchanger 124 with the returning stream from the turbine 106.

The returning stream along the path 134 that exits from the turbine 102 is a medium pressure stream relative to the returning streams from the turbine 106. The medium pressure returning stream from the turbine 102 is pumped by the pump 126 as described previously. In the heat exchanger 122, the returning medium pressure stream is condensed, releasing heat of condensation, which heats the stream 132.

The returning stream from the turbine 106, progressing along the line 136, is at a lower pressure than the stream from the turbine 102 which progresses along line 134. This returning stream 136 gives up heat in heat exchanger 124 to heat the fluid flow along the path 130 as described previously.

At point 138, the streams progressing along the paths 130 and 132 are combined and then divided into three streams which pass through heat exchangers 116, 118 and 120 respectively. The stream passing through line 140 is heated by the return stream in the line 136 which exited from the turbine 106. The fluid stream progressing along line 142 is heated by the medium pressure returning stream in line 134 which exits from turbine 102. Finally, the fluid flow through the line 144 is heated by an external heat source in the heat exchanger 116. As a result of the processes occuring in the heat exchangers 116, 118 and 120, each of the exiting flows along the lines 144, 142 and 140 is evaporated and slightly superheated.

Each of these slightly superheated streams are combined and pass through a heat exchanger 110 with heating by an external heat source. The flow exiting from the heat exchanger 110 is sent into the high pressure turbine 102 where it is expanded to a medium pressure to produce work.

The flow exiting from the turbine 102 is divided into two streams. One stream progresses along the path 134 and the other stream progresses along the path 146. The fluid flow through the path 134 is cooled and condensed, as described previously, to provide heat for preheating.

The stream progressing along the path 146 is reheated in heat exchanger 112 and is then expanded in the intermediate pressure turbine 104 to produce work. Thereafter, the stream is reheated in the heat exchanger 114 by an external heat source and then expanded in the low pressure turbine 106 to produce work. The flow exiting from the turbine 106 is a relatively low pressure returning stream. This stream progresses along the path 136 to be cooled in the heat exchanger 120, providing heat for the stream 140 as described previously. Ultimately the stream passes to the subsystem 108.

While the present invention has been described with two stage cooling of the stream progressing along the path 134 and two stage heating of the turbine 102 feed water, those skilled in the art will appreciate that the present invention can be implemented with single, double, triple or multiple stage heating of the feed water and cooling of the flow through the path 134.

A Kalina cycle condensing subsystem 108', shown in FIG. 2, is advantageously used as a subsystem 108 in the system shown in FIG. 1. In order to condense the working fluid stream, a distillation-condensation subsystem is employed when the pressure of the incoming stream to the system 108 is substantially lower than the pressure necessary to provide condensation of the returning low pressure stream at normal ambient temperatures.

The stream from the path 136 is sent into a heat exchanger 200 where it is cooled and partially condensed, releasing heat. Thereafter the stream passes through the heat exchanger 210, where it is further cooled and condensed. The stream is then mixed with a stream of lean solution at the point 212. As will become apparent subsequently, the lean solution is a solution which contains a higher proportion of a higher boiling temperature component than the stream exiting from the heat exchanger 210. The new stream, called the basic solution, has an increased content of the higher boiling component in comparison with the returning low pressure stream and for this reason can be completely condensed by a cooling source such as water. After complete condensation in the condenser 214, the basic solution is pumped by a pump 216. The basic solution is then sent into the heat exchanger 210 where it is heated by the returning streams from the heat exchangers 200 and 218.

Usually the temperature of the flow heading from the heat exchanger 210 toward the heat exchanger 218 is slightly below the boiling temperature of the fluid. The stream is divided into three separate paths 220, 222 and 224. The fluid progressing along the path 222 is sent into the heat exchanger 200 where it is partially heated and partially evaporated. The stream progressing along the path 220 is sent into the heat exchanger 218. Thereafter, the streams 220 and 222 are recombined to form the stream 226.

The stream 226, a vapor-liquid mixture, passes through a gravity separator 228 where it is separated into lean stream 232 and rich stream 230. Both streams 230 and 232 are sent through the heat exchanger 218 counterflow to the stream 220. The rich stream 230 is enriched with the light (lower temperature boiling) component and is cooled and partially condensed in the heat exchanger 218.

The partially condensed rich stream is combined with the flow from the path 224 producing a working solution composition. The working solution composition passes through heat exchanger 234 where it is further cooled and condensed. From here it is finally sent into the condenser 214 where it is fully condensed by a cooling source.

The condensate is pumped by a pump 236 to an intermediate pressure. Thereafter, it is sent counterflow through heat exchanger 234 where it is preheated. After preheating the stream is finally pumped to a high pressure by the pump 238 where it exits from the subsystem 108'.

Returning now to the lean stream, which is enriched with the heavier (higher temperature boiling) component, exiting from the gravity separator 228 along the line 232, the lean stream is cooled in the heat exchanger 218. Then it is further cooled in the heat exchanger 210 providing heat for the output flow from the pump 216. Thereafter, the stream progressing along the path 232 is throttled by the throttle valve 240 and is mixed at 212 as described previously.

The parameters of flow at the various points indicated in FIGS. 1 and 2 are design variables that can be chosen in a way to obtain the maximum advantage from the system 100. One skilled in the art will be able to select the design variables to maximize performance under the various conditions and circumstances that may be encountered, while achieving a heat balance. The parameters of the various process points, shown in FIG. 1, are subject to considerable variation depending on specific circumstances.

In order to further illustrate the advantages that can be obtained by the present invention, a set of calculations were performed. In these calculations, an illustrative power cycle in accordance with the system shown in FIGS. 1 and 2 was selected wherein the working fluid was a water-ammonia mixture. The parameters for the theoretical calculations (assumed ambient temperature 60° F.) which were performed utilized standard ammonia-water enthalpy-concentration diagrams. In the following table the points set forth in the first column correspond to the points in FIGS. 1 and 2. The column headed by the letter "G" shows the weight of the fluid at each point in proportion to the weight of fluid at the point 38.

                  TABLE I
    ______________________________________
                                   NH.sub.4 Concen-
                                   tration (lbs
 &nbs