Abstract:
Reheat of reheat regenerative steam power cycle increases its efficiency by increasing the average temperature of heat reception. In spite of such an increase in efficiency, reheating increases the irreversibility of feed water heaters by using superheated steam of a greater temperature difference in the regenerative cycle. This invention introduces some modifications to the regular reheat regenerative steam power cycle that reduces the irreversibility of the regenerative process. The invention applies reversible reheating in addition to the regular reheating and uses smaller temperature differences across feed water heaters than the regular cycle. A comparison study between the regular reheat regenerative cycle and the invented cycle is done. The results indicate that a gain in efficiency of up to 2.5% is obtained when applying invented cycle at the same conditions of pressure, temperatures, number of reheating stages, and feed water heaters. In addition, the invented cycle has some practical advantages associated with up to 50% reduction in the mass flow rate that is regularly reheated for the same output power. Such advantages such as less pressure drop and heat transfer loss. Such advantages allow us to use a greater number of reheating stages of the invented cycle for the same pressure drop and heat transfer losses of the reheater pipes of the regular cycle. Another practical advantage of the invented cycle over the regular cycle is higher heat transfer coefficients for the heat exchangers of the feed water heaters because they are mainly operated in the two-phase region. Such practical advantage results in smaller sizes for the heat exchangers of the invented cycle compared with the ones for the regular cycle.

Description:
REFERENCES CITED 
     Bassily, A. M., 1999, “Improving the Efficiency and Availability Analysis of a Modified Reheat Regenerative Rankine Cycle” Proceedings of the Renewable and Advanced Energy Systems for the 21 st  Century, Lahaina, Maui, Ha. April 11-15. 
     Moran, M. J., and Shapiro, H. N., 1995 , Fundamentals of Engineering Thermodynamics , John Wiley &amp; Sons, Inc., New York, 3 rd  Edition, pp. 590-610. 
     TECHNICAL FIELD 
     The present invention relates to the field of power generation system of the continuous combustion type using steam as the working medium. The general objective of the invention is to provide a system of power generation, having higher efficiency than the current systems while maintaining low capital cost, leading to a total running cost that is lower than the total running cost of the existing systems. 
     BACKGROUND OF THE INVENTION 
     Increasing the efficiency of power generation can be done by increasing the average temperature of heat reception through regeneration or reheating. The main purpose of reheating is to ensure high efficiency of expansion through steam turbines. The average temperature of heat reception can be increased through raising the steam generator pressure (P x ). As P x  increases, there will be need for more stages of reheating to ensure high efficiency of expansion in steam turbines. As the number of reheating stages grows, more steam will be extracted for regeneration at high superheat temperature that has high temperature difference of heat transfer. Such a high temperature difference of heat transfer increases the irreversibility of feed water heaters. There is no feasible method is known to reduce the irreversibility of feed water heaters in case of using superheated steam for feed heating. This invention introduces some modifications to the Rankine Reheat Regenerative cycles that reduce the regeneration irreversibility and increase the cycle efficiency. 
     BRIEF SUMMARY OF THE INVENTION 
     The invention is particularly advantageous for use in systems that use steam as a working medium; however, the invention is also advantageous for power systems that use any other fluids as working media. The invention can also be applied to the combined cycle power systems and Binary cycle power systems. 
     In general, it may be said that I attain the principal object of the invention, as well as the other objects thereof which will hereinafter appear, by further expanding the required amount of the working medium to be reheated just for the purpose of further expanding it in rotary turbines to produce power. The required amount of the working medium to heat the fluid entering each feed heater is extracted at almost the same pressure that corresponds to that heater. The remainder amount of that required for feed generation of the working medium after expansion if it is in a two-phase condition is allowed to enter a separator to convert the inlet two phase of the working medium to two outlets. The first outlet is dry gas and the second outlet is liquid. The dry gas will either be reheated to higher temperature just for the purpose of effective expansion in the following stage of expansion in a rotary turbine, or will be allowed to expand in the following stage of expansion without reheating. The liquid working medium out of the separator will mix with the outlet of that feed heater. If the remainder amount of that required for feed regeneration after expansion was in a gas phase condition, it is allowed to expand further in the same rotary turbine to the pressure that equal to the pressure of the next feed heater. By this process, I am enable to use working medium in a two-phase region to heat the feed heater at a pressure that is almost equal to the pressure of that heater, resulting: 
     First, a reduction in the feed water heater irreversibility since the temperature difference of heat transfer is minimum, resulting in a higher efficiency for the power system. 
     Second, a higher heat transfer coefficient since the heat transfer coefficient of the condensing two-phase working medium used to heat the working medium entering feed heater is up to 200 times that of a gas-phase working medium, resulting in a smaller and cheaper heat exchange units for feed generation. 
     Third, the amount of working medium that is expanded further for the purpose of power generation is reduced significantly. The results show that up to 50% reduction in the mass flow rate of the reheater pipes of the invented cycle over the regular current Rankine reheat regenerative cycle at the same conditions of temperatures, pressures, number of feed water heaters, and reheating stages. Such results lead to up to 75% reduction in the pressure drop of the reheater pipes and significant reductions in the heat transfer losses from such pipes (assuming the same pipe sizes and coefficients of friction), resulting in further improvement in thermal efficiency. 
     Therefore, implementing the invention is expected to reduce the capital cost of the equipment and the cost of energy to run it, resulting in a reduction of the total cost. The invention is applicable to many different arrangements of power systems and for the purpose of illustration I have shown in the accompanying drawing several schematic diagrams for carrying the invention into effect, together with the corresponding illustrations of the thermal characteristics of those cycles. 
     In the systems illustrated, the working medium is water in the liquid phase, steam in the gas phase. Any kind of fuel can be applied to those systems such as fossil fuel (oil, natural gas, coal), nuclear fuel. For convenience, I will refer, but without limitation to the working fluid as water in a liquid form and steam in a gas form. It is understood that other media having equivalent functions may be employed instead. 
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS 
     FIG. 1 a  is a schematic diagram of a simple power cycle embodying the invention; 
     FIG. 1 b  is an illustrative diagram of the cycle shown in FIG. 1 a  as it is ideally represented on a temperature-entropy diagram; 
     FIG. 2 a  is a schematic diagram of a steam power cycle in a system employing a plurality of turbines, feed heaters, pumps, and separators and embodying the invention; 
     FIG. 2 b  is an illustrative diagram of the cycle shown in FIG. 2 a  as it is ideally represented on a temperature-entropy diagram; 
     FIG. 3 is a schematic diagram similar to FIG. 2 a  showing a second arrangement of a steam power cycle in a system employing a plurality of turbines, feed heaters, pumps, and separators and embodying the invention; 
     FIG. 4 is an illustrative diagram of the cycle shown in FIG. 3 as it is ideally represented on a temperature-entropy diagram; 
     FIG. 5 is a schematic diagram similar to FIG. 3 showing a third arrangement of a steam power cycle in a system employing a plurality of turbines, feed heaters, pumps, and separators and embodying the invention; 
     FIG. 6 is an illustrative diagram of the cycle shown in FIG. 5 as it is ideally represented on a temperature-entropy diagram; 
     FIG. 7 is a schematic diagram similar to FIG. 3 showing a fourth arrangement of a steam power cycle in a system employing a plurality of turbines, feed heaters, pumps, and separators and embodying the invention; 
     FIG. 8 is an illustrative diagram of the cycle shown in FIG. 7 as it is ideally represented on a temperature-entropy diagram; 
     FIG. 9 b  is a schematic diagram showing a fifth arrangement of a steam power cycle in a system employing a plurality of turbines, feed heaters, pumps, and separators and embodying the invention; 
     FIG. 10 is an illustrative diagram of the cycle shown in FIG. 9 b  as it is ideally represented on a temperature-entropy diagram; 
     FIG. 11 is a schematic diagram similar to FIG. 9 b  showing a sixth arrangement of a steam power cycle in a system employing a plurality of turbines, feed heaters, pumps, and separators and embodying the invention; 
     FIG. 12 is an illustrative diagram of the cycle shown in FIG. 11 as it is ideally represented on a temperature-entropy diagram; 
     FIG. 13 is a schematic diagram similar to FIG. 9 b  showing a seventh arrangement of a steam power cycle in a system employing a plurality of turbines, feed heaters, pumps, and separators and embodying the invention; 
     FIG. 14 is an illustrative diagram of the cycle shown in FIG. 13 as it is ideally represented on a temperature-entropy diagram; 
     FIG. 15 is a schematic diagram similar to FIG. 13 showing an eightieth arrangement of a steam power cycle in a system employing a plurality of turbines, feed heaters, pumps, and separators and embodying the invention; 
     FIG. 16 is an illustrative diagram of the cycle shown in FIG. 15 as it is ideally represented on a temperature-entropy diagram; 
     FIG. 17 is diagrammatic illustration of a steam separator; and 
     FIG. 18 is a diagrammatic illustration of a multi-pass heat exchanger. 
    
    
     DETAILED DESCRIPTION 
     FIG. 1 a  shows a schematic diagram of a cycle that comprises three feed water heaters, three turbines, four water pumps, one steam separator, one steam generator, one condenser, and electric generators. That cycle carries the invention into effect. FIG. 1 b  shows the temperature-entropy diagram of the cycle shown in FIG. 1 a  (with no pressure drops or heat losses). At point a, steam exits steam generator  1 , at a superheated condition (about 110 bar and 450° C.) and expands adiapatically to a lower pressure of 40.1 bar at point b, in large turbine  2 , where steam still in a superheated condition. Such expansion generates mechanical power that is usually converted to electricity in an electrical generator. It is understood that every step of expansion in a steam turbine produces mechanical power that is converted to electricity using electrical generators. A predetermined amount of superheated steam at point b is extracted from large turbine  2 . Such extraction can be done by controlling a valve on an exhaust pipe at a section that corresponds to the pressure at b in large turbine  2 . If the superheated steam at b is relatively at high superheat temperature, additional heat exchanger  7  can be used to exchange heat between the superheated steam at b and the saturated steam at g. Steam at point g has a lower pressure than steam at b as shown in FIG. 1 b . The function of additional heat exchanger  7  is to raise the temperature of the reheated steam at g to a higher temperature at h. Steam at h enters the steam generator  1  for the purpose of reheating. The first function of additional heat exchanger  7  is to reduce the amount of heat added to the reheated steam in steam generator  1 , thus increasing the cycle efficiency. The second function of additional heat exchanger is to reduce the temperature difference of heat transfer across water heater  14 , thus reducing the irreversibility of water heater  14  and increasing the cycle efficiency. The superheated steam stream that exits additional heat exchanger  7 , at point c is at a saturated condition of about 250° C. where it enters water heater  14 . Water heater  14  could be closed feed water heater or an open feed water heater (direct contact heater). In such a direct contact heater, saturated steam mixes with the pumped hot water at p (165° C.), resulting in saturated water at a higher temperature of 250° C. at point q (limiting our discussion to only open feed water heaters). Hot water at q is pumped using pump  11  to a relatively higher pressure of 110 bar at r, where hot water enters steam generator  1 . The predetermined amount of steam at point b is determined using a heat exchange relation that would result in a saturated water condition at point q (the output of water heater  14 ). A predetermined amount of steam at almost the same pressure of the water entering feed water heater  13 , is extracted from large turbine  2 , at point d at a pressure of about 7.1 bar and a two-phase condition. A predetermined portion of the extracted steam at d (about 165° C.) enters feed water heater  13 , where it mixes with the pumped hot water at point n (about 100° C.), resulting in a saturated water exiting the heater at point o (162° C.). The remainder amount of wet steam at point d enters steam separator  6  that separates the entering wet steam to two outlets. The first outlet is a down stream of saturated water at point o and the second outlet is an upstream of dry saturated steam at point e. The separation process as all other processes that have been discussed so far is a continuous adiabatic process at almost constant pressure. The steam separator  6  can be located as close as possible to steam turbines to minimize any pressure drops in the steam piping system. The steam separator has two functions. The first function is to allow steam to be extracted in a two-phase region (at a lower temperature difference of heat transfer across water heater  2  than in the case of using superheated steam) for the purpose of the regeneration process in water heater  2 . The second function is to allow the dry steam output of the steam separator at point e to be expanded further in small turbine  3 . If steam at point d were allowed to expand in small turbine  3  without using the steam separator, the expansion process in small turbine  3  would be very inefficient. The reason for the inefficient expansion is that steam at point d is too wet for an efficient expansion process and needs to be dried in the steam separator first. The reduction of the temperature difference of heat transfer across water heaters reduces the irreversibility of water heaters and increases the cycle efficiency. The saturated steam exiting separator  6  at e enters small turbine  3  where it is expanded adiabatically to a lower pressure of about 0.92 bar at point f. Steam at point f is in a two-phase condition enters water heater  12 , where it mixes with the water exiting water pump  8 , at point  1  (about 27° C.), resulting in a saturated water exiting water heater  12 , at point m (about 97° C.). Saturated water output of steam separator  6  mixes with hot water output of feed water heater  13  at point o. The remainder portion of steam that enters large turbine  2  is expanded adiabatically to an intermediate pressure of about 30 bar (about ¼ of the absolute pressure value at point  1 ) at point g. If additional heat exchanger  7  was used, steam at point g will be heated to a higher temperature before it enters steam generator  1  to be reheated in reheater tubes  15 , at almost constant pressure to a high temperature of about 450° C. at point i. Superheated steam at i is expanded adiabatically to the condenser pressure at point j in large turbine  4 . Steam at point j is in a two-phase condition and a vacuum pressure of about 0.033 bar. Condenser  5  is usually water-cooled or air-cooled. It is a heat exchanger unit to condense steam in a continuous manner at almost a constant pressure. Water exiting the condenser at a vacuum pressure at point k is pumped using water pump  8 , to a pressure of about 0.91 bar which is the operating pressure of water heater  12 . Water exiting water heater  12  at a pressure of about 0.9 bar at point m is pumped using water pump  9 , to a pressure of about 7.1 bar which is the operating pressure of water heater  13 . Water exiting water heater  13  at a pressure of about 7 bar at point o is pumped using water pump  10 , to a pressure of about 40 bar which is the operating pressure of water heater  14 . The thermal characteristics of the cycle shown in FIG. 1 a  are ideally represented in FIG. 1 b , just for the sake of simplicity. It is understood that there will be minor pressure and heat transfer losses and the expansion processes in turbines will not be ideally adiabatic. 
     To calculate the mass flow rate at each point of a cycle that has seven separator-heater couples, we write the energy balance for the separator-heater couple in a system of 7 separator-heater couples with maximum mass flow rate of unity shown.                    m   hn          h   sni       +       (     1   -       ∑     k   =   n     7                     m   hk       -       ∑     k   =   n     7                     m   sk         )                     h   hni         =       (     1   -       ∑     k   =     n   +   1       7                     m   hk       -       ∑     k   =   n     7                     m   sk         )          h   hno               (   1   )                     (         ∑     k   =   1       n   -   1                       m   hk       +       ∑     k   =   1       n   -   1                       m   sk         )                     h   sno       +       m   sn          h   hno         =       (         ∑     k   =   1       n   -   1                       m   hk       +       ∑     k   =   1     n                     m   sk         )                     h   sni               (   2   )                                
     Equation 1 is written for heater numbers n and Equation 2 for separator number n in a system of 7 heaters-separators where h is specific enthalpy [j/kg], m mass flow rate [kg/sec], and the subscripts hk is heater number k, hn is heater number n, sk is separator number k, sn is separator number n, hni is inlet to heater number n, hno is outlet of heater number n, sni is inlet to separator number n, sno is outlet of separator number n. Solving Equations 1 and 2 for each set of separator-heater simultaneously, we obtain the mass flow rates since the enthalpy at each point is known. 
     FIG. 2 a  shows a schematic diagram of a system that comprises 3 large scale turbines (T 1 , T 2 , &amp; T 3 ), 3 small scale turbines (T 4 , T 5  &amp; T 6 ), 7 feed water heaters (FWH 1 , FWH 2 , FWH 3 , FWH 4 , FWH 4 , FWH 5 , FWH 6  &amp; FWH 7 ), 3 steam separators (S 1 , S 2  &amp; S 3 ), one condenser (C 1 ), one steam generator, 8 water pumps (P 1 , P 2 , P 3 , P 4 , P 5 , P 6 , P 7  &amp; P 8 ), and electrical generators. FIG. 2 b  shows the thermal characteristics of the cycle shown in FIG. 2 a  on the temperature-entropy diagram. The thermal characteristics of the cycle are ideally represented on the temperature-entropy diagram (with no pressure drops or heat losses). Such a cycle carries the invention into effect. Steam exiting the steam generator at point  1  (a temperature of about 600° C. and a pressure of about 300 bar) is expanded in large turbine T 1  continuously and adiabatically to lower pressures providing mechanical power that is converted usually to electricity using an electrical generator. The amount of steam needed to heat the hot water at point  30  in feed water heater FWH 1  to point  31  is extracted from large turbine T 1  at a pressure of about 130.1 bar. The conditions at point  30  are a pressure of about 130 bar and a temperature of about 286° C. At point  31 , hot water is at almost the same pressure, but at 330° C. (saturated condition). The amount of steam needed to heat the hot water at point  28  in feed water heater FWH 2  to point  29  is extracted from large turbine T 1  at a pressure of about 70.1 bar. The conditions at point  28  are a pressure of about 71 bar and a temperature of about 242° C. At point  29 , hot water is at almost the same pressure, but at 286° C. (saturated condition). The amount of steam needed to heat the hot water at point  26  in feed water heater FWH 3  to point  27  is extracted from large turbine T 1  at a pressure of about 35.55 bar. The conditions at point  26  are a pressure of about 35.45 bar and a temperature of about 201° C. At point  27 , hot water is at almost the same pressure, but at 242° C. (saturated condition). The amount of steam needed to heat the hot water that enters feed water heaters FWH 4 , FWH 5 , FWH 6 , and FWH 7  is expanded adiabatically and continuously in large steam turbine T 1  to pressure of 15.7 bar at point  8 . The amount of steam needed to heat the hot water at point  24  in feed water heater FWH 4  to point  25  is extracted from large turbine T 1  at a pressure of about 15.75 bar. The conditions at point  24  are a pressure of about 15.65 bar and a temperature of about 158° C. At point  25 , hot water is at almost the same pressure, but at 201° C. (saturated condition). Equations 1 and 2 can be used to determine the mass flow rates entering every steam separator and feed water heater. By adding the mass flow rate entering separator S 1  to that entering feed water heater FWH 4 , the mass flow rate to be extracted from large turbine T 1  at point  8  can be determined as m 8 . By adding the mass flow rates of steam extracted at points  2 ,  4 , and  6  to m 8 , the total mass flow rate of steam extracted from large turbine T 1  can be determined as m e . By subtracting m e  from the mass flow rate entering large turbine T 1  at point  1 , the mass flow rate that is expanded adiabatically to a pressure of about 66 bar at point  33  can be determined. At point  33 , steam returns to the steam generator for reheating at almost a constant pressure of 66 bar to a high temperature of 600° C. At point  34 , steam enters large turbine T 2  and expands adiabatically and continuously to a pressure of about 14.5 bar and a temperature of about 374° C. at point  35  producing mechanical power that is usually converted to electricity in an electrical generator. Steam exiting large turbine T 2  enters the steam generator for a second stage of reheating at almost constant pressure to a temperature of about 600° C. at point  36 . The reheated steam at point  36  enters large turbine T 3  to expand continuously and adiabatically to a vacuum pressure of about 0.033 bar at point  37 . Steam at point  37  enters steam condenser C 1  where usually water or air is used to condense the steam in a continuous process at a constant pressure to water at vacuum pressure at point  17 . Water at  17  is pumped in a continuous process to a pressure of about 0.306 bar at point  18  where water enters feed water heater FWH 7 . The rest of steam that is expanded adiabatically and continuously in large turbine T 1  at point  8  enters steam separator S 1  after satisfying the required steam for feed water heater FWH 4 . In steam separator S 1 , steam is separated in a continuous process adiabatically and at almost constant pressure to two outlets. The first outlet is dry saturated steam, leaving the top of separator S 1  at point  9  at a pressure of 15.7 bar. The second outlet is saturated water leaving the bottom of separator S 1  at the same pressure of 15.7 bar where it joins the hot water exiting feed water heater FWH 4  at point  25 . Dry steam at point  9  is expanded adiabatically and continuously in small turbine T 4  to a pressure of about 5.8 bar at point  10  to produce mechanical power that is usually converted to electricity using an electrical generator. The amount of steam needed to heat the hot water at point  22  (at a pressure of about 5.78 bar and a temperature of about 112° C.) in feed water heater FWH 5  to point  23  is drawn from the steam entering separator S 2  at point  10 . At point  23  the hot water exiting the heater is at almost the same pressure, but at a temperature of 158° C. The rest of steam that exits small turbine T 4  at point  10  enters separator S 2  where steam is separated in a continuous process adiabatically and at almost a constant pressure to two outlets. The first outlet is dry saturated steam, leaving the top of separator S 2  at point  11  at a pressure of 5.8 bar. The second outlet is saturated water leaving the bottom of separator S 2  at the same pressure of 5.8 bar where it joins the hot water exiting feed water heater FWH 5  at point  23 . Dry steam at point  11  is expanded adiabatically and continuously in small turbine T 5  to a pressure of about 1.57 bar at point  12  to produce mechanical power that is usually converted to electricity using an electrical generator. The amount of steam needed to heat the hot water at point  20  (at a pressure of about 1.57 bar and a temperature of about 70° C.) in feed water heater FWH 6  to point  21  is drawn from the steam entering separator S 3  at point  12 . At point  21  the hot water exiting the heater is at almost the same pressure, but at a temperature of 112° C. The rest of steam that exits small turbine T 5  at point  12  enters separator S 3  where steam is separated in a continuous process adiabatically and at almost constant pressure to two outlets. The first outlet is dry saturated steam, leaving the top of separator S 3  at point  13  at a pressure of 1.57 bar. The second outlet is saturated water leaving the bottom of separator S 3  at the same pressure of 1.57 bar where it joins the hot water exiting feed water heater FWH 6  at point  21 . Dry steam at point  13  is expanded adiabatically and continuously in small turbine T 6  to a pressure of about 0.307 bar at point  14  to produce mechanical power that is usually converted to electricity using an electrical generator. The amount of steam needed to heat the hot water at point  18  (at a pressure of about 0.306 bar and a temperature of about 27° C.) in feed water heater FWH 7  to point  19  is drawn from the steam exiting small turbine T 6  at point  14 . At point  19  the hot water exiting the heater is at almost the same pressure, but at a temperature of 70° C. FIG. 2 b  shows the thermal characteristics of the cycle shown in FIG. 2 a  as they are represented ideally on the temperature-entropy diagram. 
     FIG. 3 shows the exact same cycle that is shown in FIG. 2 a  except that there is an additional heat exchanger to reduce the superheat temperature of the superheated steam at  2  extracted from large turbine T 1  for the purpose of heating the hot water of feed heater FWH 1 . As the superheated steam at  2  is cooled as it passes through heat exchanger HE 1 , the steam extracted from large turbine T 1  at point  33  is heated as it passes through heat exchanger HE 1  to a temperature of about 392° C. at point  33   x . The conditions at point  2  are a pressure of about 129.7 bar and a temperature of about 455° C. The conditions at point  33  are a temperature of about 357° C. and at a lower pressure than that at point  2 . The superheated steam at  2  that enters heat exchanger HE 1  exits the heat exchanger at point  3   x  where its temperature is about 367° C. FIG. 4 shows the thermal characteristics of the cycle shown in FIG. 3 as they are represented ideally on the temperature-entropy diagram. 
     FIG. 5 shows a schematic diagram of the exact same cycle that is shown in FIG. 2 a  except that there is an additional steam separator and a stage of expansion in a small steam turbine. The mass flow rate of steam that expands in small turbine T 7  will affect the mass flow rate of the reheater pipes so that such mass of small turbine T 7  can be chosen to maximize cycle efficiency or output power whatever is required. Determining such a mass flow rate, the mass flow rate of the two-phase steam that enters separator S 4  can be determined. Dry steam exits the top of separator S 4  at point  15  (at a temperature of about 70° C. saturated condition) to enters small turbine T 7  to expand to the condenser pressure. Steam exiting small turbine T 7  enters condenser C 1  to be condensed at a vacuum pressure. As steam expands in small turbine T 7  to produce mechanical power that is usually converted to electricity using an electrical generator. Separator S 4  converts the inlet two-phase steam to two outlets adiabatically, continuously and at almost a constant pressure. The first outlet is dry steam at the top of the separator at point  15  and the second outlet is saturated water out of the bottom of the separator at point  19  that joins the hot water outlet of feed heater FWH 7 . The amount of steam needed to heat the hot water at point  18  (at a pressure of about 0.306 bar and a temperature of about 27° C.) in feed water heater FWH 7  to point  19  is drawn from the steam entering separator S 4  at point  14 . At point  19  the hot water exiting the heater is at almost the same pressure, but at a temperature of about 70° C. The rest of steam that exits small turbine T 6  at point  14  enters separator S 4 . FIG. 6 shows the thermal characteristics of the cycle shown in FIG. 5 as they are represented ideally on the temperature-entropy diagram. 
     FIG. 7 shows the exact same cycle that is shown in FIG. 5 except that there is an additional heat exchanger to reduce the superheat temperature of the superheated steam at  2  that is extracted from large turbine T 1  for the purpose of heating the hot water of feed heater FWH 1 . As the superheated steam at  2  is cooled as it passes through heat exchanger HE  1 , the steam extracted from large turbine T 1  at point  33  is heated as it passes through heat exchanger HE 1  to a temperature of about 392° C. at point  33   x . The conditions at point  2  are a pressure of about 129.7 bar and a temperature of about 455° C. The conditions at point  33  are a temperature of about 357° C. and at a lower pressure than that at point  2 . The superheated steam at  2  that enters heat exchanger HE 1  exits the heat exchanger at point  3   x  where its temperature is about 367° C. FIG. 8 shows the thermal characteristics of the cycle shown in FIG. 7 as they are represented ideally on the temperature-entropy diagram. 
     FIG. 9 b  shows a schematic diagram of a cycle that is composed of 3 large scale turbines (T 1 , T 2 , &amp; T 3 ), 3 small scale turbines (T 4 , T 5  &amp; T 6 ), 7 feed water heaters (FWH 1 , FWH 2 , FWH 3 , FWH 4 , FWH 4 , FWH 5 , FWH 6  &amp; FWH 7 ), a condenser (C 1 ), a steam generator, 8 water pumps (P 1 , P 2 , P 3 , P 4 , P 5 , P 6 , P 7  &amp; P 8 ), a multi-pass heat exchanger and electrical generators. FIG. 10 shows the thermal characteristics of the cycle shown in FIG. 10 on the temperature-entropy diagram. Such a cycle carries the invention into effect. Steam exiting the steam generator at point  1  (a temperature of about 600° C. and a pressure of about 300 bar) is expanded in large turbine T 1  continuously and adiabatically to lower pressures providing mechanical power that is converted usually to electricity using an electrical generator. The amount of steam needed to heat the hot water at point  30  in feed water heater FWH 1  is extracted from large turbine T 1  at a pressure of about 130.1 bar (point  2 ). The conditions at point  30  are a pressure of about 130 bar and a temperature of about 286° C. Hot water in FWH 1  is heated to point  31  where hot water is at almost the same pressure, but at 330° C. (saturated condition). The amount of steam needed to heat the hot water at point  28  in feed water heater FWH 2  is extracted from large turbine T 1  at point  4  at a pressure of about 70.1 bar. The conditions at point  28  are a pressure of about 71 bar and a temperature of about 242° C. Hot water in FWH 2  is heated to point  29  where hot water is at almost the same pressure, but at 286° C. (saturated condition). The amount of steam needed to heat the hot water at point  26  in feed water heater FWH 3  is extracted from large turbine T 1  at a pressure of about 35.55 bar (point 6 ). The conditions at point  26  are a pressure of about 35.45 bar and a temperature of about 201° C. Hot water in FWH 3  is heated to point  27  where hot water is at almost the same pressure, but at 242° C. (saturated condition). The amounts of steam needed to heat the hot water that enters feed water heaters FWH 4 , FWH 5 , FWH 6 , and FWH 7  are added and denoted as m 9 . By applying the energy and mass balance equations on separator S 1 , the mass flow rate-entering separator S 1  can be determined as m s1 . The amount of steam needed to heat the hot water at point  24  in feed water heater FWH 4  to point  25  is extracted from large turbine T 1  at a pressure of about 15.75 bar and can be determined as m FWH4 . The conditions at point  24  are a pressure of about 15.65 bar and a temperature of about 158° C. At point  25 , hot water at almost the same pressure, but at 201° C. (saturated condition). By adding m s1  to m FWH4 , the mass flow rate that is expanded adiabatically and continuously in large steam turbine T 1  to a pressure of 15.7 bar at point  8  can be determined as m 8 . By adding m 8  to the mass flow rates extracted at  2 ,  4 , and  6 , the total mass flow rate extracted for the purpose of regeneration can be determined as m e . By subtracting me from the mass flow rate that enters large turbine T 1  at  1 , the mass flow rate that expands adiabatically to a pressure of about 66 bar at point  33  can be determined. At point  33 , steam returns to the steam generator for reheating at almost a constant pressure of 66 bar to a high temperature of 600° C. At point  34 , steam enters large turbine T 2  and expands adiabatically and continuously to a pressure of about 14.5 bar at point  35  producing mechanical power that is usually converted to electricity in an electrical generators. Steam exiting large turbine T 2  enters the steam generator for a second stage of reheating at almost constant pressure to a temperature of about 600° C. at point  36 . The reheated steam at point  36  enters large turbine T 3  to expand continuously and adiabatically to a vacuum pressure of about 0.033 bar at point  37 . Steam at point  37  enters steam condenser C 1  where usually water or air is used to condense steam in a continuous process at a constant pressure to water at vacuum pressure at point  17 . Water at  17  is pumped in a continuous process to a pressure of about 0.306 bar at point  18  where water enters feed water heater FWH 7 . The rest of steam that is expanded adiabatically and continuously in large turbine T 1  at point  8  after satisfying the required steam for feed water heater FWH 4  enters steam separator S 1 . In steam separator S 1 , steam is separated in a continuous process adiabatically and at almost a constant pressure to two outlets. The first outlet is dry saturated steam, leaving the top of separator S 1  at point  9  at a pressure of 15.7 bar. The second outlet is saturated water leaving the bottom of separator S 1  at the same pressure of 15.7 bar where it joins the hot water exiting feed water heater FWH 4  at point  25 . Dry steam at point  9  is expanded adiabatically and continuously in small turbine T 4  to a pressure of about 5.8 bar at point  10  to produce mechanical power that is usually converted to electricity using an electrical generator. The amount of steam needed to heat the hot water at point  22  in feed water heater FWH 5  to point  23  is drawn from the steam exiting small turbine T 4 . The conditions at point  22  are a pressure of about 5.78 bar and a temperature of about 112° C. In heat exchanger HE 2  steam is reheated for the purpose of a more efficient expansion in the following stage of expansion. Steam exits multi-pass heat exchanger HE 2  at point  11  in a superheated condition where it enters small turbine T 5  to be expanded to a lower pressure adiabatically and continuously to produce mechanical power that is usually converted to electricity using an electrical generator. The amount of steam needed to heat the hot water at point  20  (at a pressure of about 1.57 bar and a temperature of about 70° C.) in feed water heater FWH 6  to point  21  is drawn from the steam entering heat exchanger HE 2  at point  12 . At point  21  the hot water exiting the heater is at almost the same pressure, but at a temperature of 112° C. The rest of steam that exits small turbine T 5  at point  12  enters multi-pass heat exchanger HE 2  where steam is reheated in a continuous process adiabatically and at almost a constant pressure to superheated steam, leaving the heat exchanger at point  13  at a pressure of 1.57 bar. Superheated steam at point  13  is expanded adiabatically and continuously in small turbine T 6  to a pressure of about 0.307 bar at point  14  to produce mechanical power that is usually converted to electricity using an electrical generator. The amount of steam needed to heat the hot water at point  18  (at a pressure of about 0.306 bar and temperature of about 27° C.) in the feed water heater FWH 7  to point  19  is drawn from the steam exiting small turbine T 6  at point  14 . At point  19  the hot water exiting the heater is at almost the same pressure, but at a temperature of 70° C. FIG. 10 shows the thermal characteristics of the cycle shown in FIG. 9 b  as they are represented ideally on the temperature-entropy diagram. 
     FIG. 11 shows a schematic diagram of the exact same cycle that is shown in FIG. 9 b  except that there is an additional pass in multi-pass heat exchanger HE 2  to reheat the steam exiting small turbine T 5  and a stage of expansion in small steam turbine T 6 . The mass flow rate of steam that expands in small turbine T 7  will affect the mass flow rate of the regular reheater pipes so that such a mass flow rate through small turbine T 7  can be chosen to maximize cycle efficiency or output power whatever is required. Determining such a mass flow rate, the mass flow rate of the two-phase steam that enters the final passage of multi-pass heat exchanger HE 2  at point  14  can be determined. Superheated steam exits heat exchanger HE 2  at point  15  (at a temperature of about 70° C. and saturated condition) to enters small turbine T 7  to expand to the condenser pressure. Steam exiting small turbine T 7  enters condenser C 1  to be condensed at a vacuum pressure. As steam expands in small turbine T 7  to produce mechanical power that is usually converted to electricity using an electrical generator. The amount of steam needed to heat the hot water at point  18  (at a pressure of about 0.306 bar and temperature of about 27° C.) in feed water heater FWH 7  to point  19  is drawn from the steam entering multi-pass heat exchanger at point  14 . At point  19  the hot water exiting the heater is at almost the same pressure, but at a temperature of about 70° C. The rest of steam that exits small turbine T 6  at point  14  enters multi-pass heat exchanger HE 2 . FIG. 12 shows the thermal characteristics of the cycle shown in FIG. 11 as they are represented ideally on the temperature-entropy diagram. 
     FIG. 13 shows the exact same cycle that is shown in FIG. 9 b  except that there is an additional heat exchanger to reduce the superheat temperature of the superheated steam at  2 . Steam at point  2  is extracted from large turbine T 1  for the purpose of heating the hot water of feed heater FWH 1 . The conditions at  2  are a pressure of about 129.7 bar and a temperature of about 455° C. As the superheated steam at  2  is cooled as it passes through heat exchanger HE 1 , the steam extracted from large turbine T 1  at point  33  is heated as it passes through heat exchanger HE 1  to a temperature of about 392° C. The conditions at point  33   x  are a temperature of about 357° C. and at a lower pressure than that at point  2 . The superheated steam at  2  that enters heat exchanger HE 1  exits the heat exchanger at point  3   x  where its temperature is about 367° C. FIG. 14 shows the thermal characteristic of the cycle shown in FIG. 13 as they are represented ideally on the temperature-entropy diagram. 
     FIG. 15 shows the exact same cycle that is shown in FIG. 11 except that there is an additional heat exchanger to reduce the superheat temperature of the superheated steam at  2  that is extracted from large turbine T 1  for the purpose of heating the hot water of feed heater FWH 1 . As the superheated steam at  2  is cooled as it passes through heat exchanger HE 1 , the steam extracted from large turbine T 1  at point  33  is heated as it passes through heat exchanger HE 1  to a temperature of about 392° C. at point  33   x . The conditions at point  2  are a pressure of about 129.7 bar and a temperature of about 455° C. The conditions at point  33  are a temperature of about 357° C. and at a lower pressure than that at point  2 . The superheated steam at  2  that enters heat exchanger HE 1  exits the heat exchanger at point  3   x  where its temperature is about 367° C. FIG. 16 shows the thermal characteristics of the cycle shown in FIG. 15 as they are represented ideally on the temperature-entropy diagram. 
     Steam separators are used in all modern steam generators except once-through types. The steam separator is shown in FIG.  17 . The steam separator comprises a closed cylinder that has one inlet and two outlets. The steam separator separates the wet (two-phase steam) to dry saturated steam and saturated water. Wet steam enters the drum from its side. Saturated water has higher density than steam comes out of the downcomers. Saturated steam entrains water and exits the top of the drum. The shown screens increase the efficiency of separation by allowing only dry steam to go through. The water level inside the drum has to be controlled to be within a specific range for efficient operation. The level control can be done measuring the water level inside the drum instantaneously using a level measuring device that has instantaneous output signal connected to a level transmitter. The output of the transmitter is connected to a controller that is connected to a control valve that controls the inlet wet steam to the drum as shown in FIG.  17 . If the set value for the water level was lower than the measured value, the controller will send a signal to the control valve to open the valve (by exerting a greater pressure or a smaller pressure on the valve diaphragm depending on the kind of valve). If the set value for the valve level was higher than the measured value, the controller signal will be to close the valve to reduce the water level inside the drum. 
     FIG. 18 shows the multi-pass shell and tube heat exchanger. The heat exchanger comprises a shell that has many tubes through which high-pressure, hot water passes through. The spaces around the tubes have buffles that support the tubes and direct the steam flow around the tubes to be in counter directions to the water flow inside the tubes to achieve the highest temperature difference and heat transfer rate. The shell is divided to four sections for four passages. The first passage is for steam outlet of separator S 1  at  9  that enters that passage of the multi-pass heat exchanger where steam is superheated to enter turbine T 4  at point  9   b . The second passage for steam outlet of turbine T 4  at point  10  that enters that passage of the multi-pass heat exchanger where steam is superheated and exit the shell to enter turbine T 5  at point  11 . The third passage for steam outlet of turbine T 5  at point  12  that enters that passage of the multi-pass heat exchanger where steam is superheated and exit the shell to enter turbine T 6  at point  13 . The fourth passage is for steam outlet of turbine T 6  at point  14  that enters that passage of the multi-pass heat exchanger where steam is superheated and exit the shell to enter turbine T 7  at point  15 . 
     From the foregoing description it will be evident that the invention is applicable to a wide variety of arrangements of power systems and it is to be understood as embracing all such systems as may fall within the terms of the appended claims when construed as broadly as is consistent with the state of prior art.