RU2248453C2 - Electric power station and method of power generation with combination of cycles - Google Patents

Abstract

FIELD: power engineering.

SUBSTANCE: invention relates to power station and method of power generation basing on principle of combination of cycles providing higher specific capacity and flexibility of operating parameters of power station without corresponding increase of its thermal power rating. Process of improving thermal efficiency of combination cycles can be provided owing to use additional fuel and/or supply of heat. Particularly, gas turbines whose exhaust gases are delivered into steam generator for regeneration of heat can be additionally heated. Proposed system, as a whole, provides high efficiency of power station with combination of cycles mainly owing to Rankine (lower) cycle. Models of implementation of invention include load driven by engine of upper cycle set into action by liquid from upper cycle delivered into heat regeneration device. Said heat regeneration device is heated by additional fuel and/or from additional heat source to form higher energy consuming liquid of lower cycle and/or increase of amount of said liquid used as energy source for engine of lower cycle setting into motion load (potentially the same load as that in engine of upper cycle). Energy of liquid of upper cycle and liquid of lower cycle sets into action engines of upper and lower cycles, respectively, however, said liquids and/or exhaust of corresponding engines can found wide application at combined generation of heat and electric energy.

EFFECT: considerable increased of power rating of steam turbines and total efficiency of power station.

33 cl; 51 dwg

Description

FIELD OF TECHNOLOGY

The present invention relates to power plants using equipment that can operate with or without combination of duty cycles. As will be shown below, the ever-increasing need for more efficient and environmentally friendly methods of energy production has led to large-scale research in order to create systems and methods for solving problems in these areas. However, modern technologies mainly fairly short-sightedly assess the combined effect of increasing the efficiency of energy production methods on the economy and the environment.

In the framework of the present invention, it is proposed to move away from tradition and consider, taking into account the analysis of the impact on the economy and the environment, the full range of issues necessary to solve the problem of creating a power unit of the required capacity. The present invention, through the use of this analysis and the application of the method of additional heat and fuel recovery, can significantly reduce the economic costs and environmental impact of existing power plants by reducing units of equipment while maintaining or maintaining an acceptable level of electricity production. As a result of such an improvement, cheaper electric energy will be obtained in a more environmentally friendly way than is possible with traditional power units using a combination of duty cycles currently known in the art.

BACKGROUND

Short review

Power plants and heat and electricity plants use gas turbines as primary engines for generating energy. Gas turbines operate on the Brayton thermodynamic cycle and typically discharge a large number of exhausts at sufficiently high temperatures into the atmosphere. These exhausts sent to a steam boiler for regeneration (the so-called steam generator with heat recovery - HRSG) form steam that can be used to obtain additional energy and / or for the current process. To generate additional energy, the steam can be sent to additional steam turbine (ST) .Thus, gas turbines operate according to the Braiton cycle, and steam turbines generate energy according to the Rankine cycle. This cycle is called the "combined cycle." In this scheme, the Brae steam turbine cycle tone called "topping cycle," a Rankine cycle steam turbines - "bottoming cycle", as in the "topping cycle" is generated energy needed for lower loop thus ensures the functional connection of said loops within the prior art..

Rankine cycle

Steam has been used for energy for over a century. At earlier stages, a pump was used to obtain the required water pressure, a steam boiler to heat the water before it was converted to steam, and a steam engine, usually a piston type, to produce mechanical energy. Such plants were used in factories, locomotives, steamboats, and in other cases where mechanical energy was required.

With the development of technology, the trend of using steam engines has lost its significance and the use of steam turbines has become more relevant. One of the advantages of using steam turbines was the efficiency of their full cycle when they were used in conjunction with a condenser. In this case, the vapor reaches atmospheric pressure, which is only slightly higher than the absolute vacuum (from 0.5 to 2 psi, absolute pressure (psia). This allows the steam to expand more than in the configuration with the exhaust into the atmosphere, to extract more energy from mass of steam, thus generating more energy and, consequently, increasing the overall efficiency of the steam cycle.This full steam cycle, from the thermodynamic point of view, is called the Rankine cycle.

Figure 1 shows the thermodynamic regime of the Rankine cycle. The graph (100) shows a diagram of the Rankine cycle in the coordinates of pressure and volume. From point (101) to point (102), water is under pressure at a constant volume. From point (102) to point (103), the water is boiled until it turns into steam at constant pressure. The segment from point (103) to point (104) shows the process in which the steam isentropically expands and directly performs the work. Then, from point (104) to point (101), the steam at low pressure condenses back to the state of water, and the cycle ends.

Graph (110) shows a diagram of the Rankine cycle versus temperature and entropy. From point (111) to point (112), water is under pressure. From point (112), water is boiled at a constant temperature until all of it turns into steam, and then it overheats to point (113). From point (113) to point (114), a process is shown in which the vapor isentropically expands and does the work. From point (114) to point (111), steam at low pressure undergoes condensation to a state of water at a constant temperature, and the cycle ends. Information on the Rankine cycle can be found in the source: Eugene A. Avallone, Theodore Baumeister III. - Marks' Standard Handbook For Mechanical Engineers (Ninth Edition) / ISBN 0-07-004127-X, 1987, section 4-20.

Power plant duty cycle

For several decades, the Rankine cycle has been used to generate most of the electricity in the United States and many other countries. Figure 2 shows the basic Rankine cycle diagram, including four main components: a feed pump (201), a steam boiler with an evaporator (203) and a superheater (205), a steam turbine (207) and a condenser (209). It should be noted that one or more units of each unit may be used, but for ease of demonstration, FIG. 2 shows one unit of each device. The Rankine cycle begins (steam pressure is less than 3206.2 psia) when water enters the inlet (211) of the feed pump (201). Water under pressure (202) is sent to the evaporator (203), in which the water under pressure is heated. Usually the heating is provided by the heat obtained by burning fuel in the boiler, the heat from the combustion process is transferred to the water that passes through the pipes and / or passages of the boiler. When water receives enough heat, it boils and turns into steam (204). Steam at this stage is in a two-phase state: both steam and water coexist at the same pressure and temperature, called saturation pressure and saturation temperature. In most designs created in recent decades, such steam (204) is sent to a superheater (205), where it is heated to a temperature above the saturation temperature. Such steam (206) is called "superheated" steam. Superheated steam reduces (but does not completely remove) the possibility of water entering the steam turbine (207), causing undesirable effects, as this can significantly damage the inside of the steam turbine. More important, however, is that superheated steam provides higher cycle efficiency. This is very important in the case of large central power plants.

After formation, superheated steam (206) is sent through one or more pipes to a steam turbine (207). In a steam turbine, steam expands and rotates the rotor. After the turbine passes, the steam loses pressure, leaves the steam turbine (208) and is sent to the condenser (209), in which it is converted back to water. Typically, this device is made in the form of a tubular heat exchanger, but can be made in another form, for example, in the form of a spray chamber, an air-cooled condenser, or in the form of other devices for heat exchange used for similar purposes. After heat is taken from low pressure steam and converted to water by condensation, the condenser collects water into compartments, which are usually called condensate collectors (210), from which condensate is usually passed through a line (211) using a pump and sent back to the boiler’s feed pump ( 201). The mechanical energy generated by the steam turbine is converted in the generator (212) into electrical energy. Such a cycle of water passing from the starting point through the above system and returning back to the starting point is a simple Rankine cycle.

In modern power units using steam as a working fluid, a boiler is usually used to produce steam. The boiler can be heated by various types of fuel, including oil, natural gas, coal, biological mass, as well as others - such as nuclear fuel. In boilers, a combination of several fuels can be used. Depending on the cost of production, the cost of fuel, maintenance issues and other factors, owners and engineers choose the steam pressure and temperature level at which the boiler generates steam.

Due to the impressive size and weight of steam turbines, considerable time is required for warming up. This is due to the thick metal case and large heavy screws that are used in their design. Machines take a long time for their heavy parts to heat up equally to avoid interference between stationary and moving parts, which occurs due to different thermal expansion.

Despite the fact that the heavy construction causes a slow heating process, it provides additional strength and stability of the level of energy production. Even after four years of almost constant operation, the performance of a steam turbine is reduced by less than 2%. This decrease in performance, together with the fact that the feed pumps consume only about 2% of the energy generated by the steam turbine, means that the level of performance of steam turbines remains almost at the optimal level for a long period even with a decrease in the pressure created by the feed pump. In other words, if the efficiency of the feed pump is reduced from 75 to 65%, the energy costs for water supply increase from 2 to 2.31%. This has a slight effect on the overall performance of a power unit using the Rankine cycle, which is one of its main advantages.

The Brighton cycle

The Brighton cycle is significantly different from the Rankine cycle, since the main component of the cycle is the compression of the working fluid, which is used as a compressible gas. During this process, a large amount of energy is consumed, therefore, effective compression of the working fluid is an integral condition for the effectiveness of the Brighton cycle. Common engines that use the Braithon cycle are turboprop aircraft engines, jet aircraft engines, and stationary gas turbines. The stages of engine operation include: gas suction, compression to a higher pressure, usually 3-30 times higher than the ambient air pressure, a subsequent increase in temperature through direct combustion (although heating by an external source can be carried out), expansion of the obtained hot gases under high pressure in the turbine compartment. Aircraft engines first create traction to raise the aircraft into the air, so the energy they produce is almost never mechanical energy (a turboprop gas turbine engine can propel a propeller, and can also provide traction due to the high speed of the exhaust gases).

In the case of stationary gas turbines, the purpose of the engine is to generate mechanical energy. Approximately 2/3 of the energy generated by the turbine is spent on ensuring the operation of the compressor, and 1/3 of the remaining generated energy is left to work with an external load. This drawback of gas turbine systems can be used as an advantage in the present invention, as will be shown later in this document.

Aircraft engines use the Braiton cycle because these engines have high thrust to weight ratios. This is necessary to minimize the mass of the aircraft so that it can fly. In stationary applications, gas turbines are used to generate electrical energy at maximum load. Another advantage of the engines operating on the Braithon cycle, compared with the Rankine cycle, are relatively quick start and stop. Since steam turbines are large and heavy, they have to be started slowly, since it is necessary to warm up a powerful casing in order to avoid thermal deformation and the occurrence of friction between stationary and rotating elements. A power plant’s steam turbine may take 24 hours to warm up and start from scratch until it reaches full capacity. Due to their lower operating pressure and lighter weight, gas turbines are easier to start and can reach full power in a few minutes.

For this reason, many utilities in the United States and other countries use gas turbines to generate electricity during a spike in consumption. Gas turbines are not effective for continuous use as part of the daily cycle (from 25% to 30%), but satisfy the demand for electricity for several hours every day.

Steam Turbine Design

When designing a steam turbine for its use in a power plant (at a constant speed), the design engineer first analyzes the required turbine parameters determined by the customer. This is because the steam turbine will be designed in accordance with customer requirements and produced for the customer based on its specifications. A steam turbine is designed not from scratch, as can be understood from the reference to design "on order", but using components from the "family" of well-known technical means, but it will have a unique steam distribution system in each case. After studying the parameters of the turbine, the design engineer determines the conditions for the use of steam within a particular power plant and, based on the data obtained, determines how the working fluid will enter the turbine’s high-pressure unit. Based on this, the design engineer selects the optimal casing for the high pressure unit in each case. In the same way, he chooses the optimal casing for intermediate and low pressure chambers.

Knowing which casing to use, the engineer selects the appropriate set of turbine blades (stationary and moving). The size of the turbine blades is determined mainly in accordance with the volume of the stream (compared to the mass of the stream) of the steam passing through the turbine. After selecting the casing and blades of the turbine, the engineer completes the design of the steam turbine by selecting the appropriate valves, controls, equipment and other devices necessary for the operation of the steam turbine. The final project is the development of a high-performance steam turbine according to customer requirements.

With respect to this design scheme, it is interesting to note that two steam turbines with the same steam parameters, but with very different power parameters (for example, 200 MW and 400 MW), may look almost the same. This is due to the fact that optimally matched casing is suitable for both turbines. However, due to the large difference in the volume of the steam flow, blades that are twice as large in size (weight) will be used in a large turbine. It is also interesting to note that in both turbines the same cladding of high and intermediate pressure units can be used. This means that manufacturing a larger steam turbine even with significantly larger operating parameters may be slightly more expensive than manufacturing a steam turbine with lower operating parameters. This fact can be considered as an advantage in the present invention, which will be shown later in this document.

Gas turbine design

Unlike a steam turbine, a gas turbine is not produced individually according to the needs of each customer. Although related accessories such as starting means, engine oil cooling devices, and monitoring devices can be ordered for any particular case, the main engine always remains rated. Much of the above is explained by the fact that gas turbines are monoblock power plants that require only fuel to produce energy. Conversely, steam turbines are just a component of a power plant, and they need a boiler, feed pump and condenser to make up an existing power plant. The compressor compartment, the combustion chamber and the turbine compartment of the gas turbine must be designed in such a way that they are combined in a single design. Since the design of a gas turbine is a laborious engineering task, after completion of the design and large-scale testing, the turbines are launched into mass production without any changes in the drawings of the main engine. This makes it impossible for the customer to specifically indicate the power of the generated energy using only gas turbines or with a combination of cycles within the framework of the current level of technology. When constructing power plants with combined cycles, the customer has to choose from the nominal offers of the manufacturer that best suit his needs in terms of power unit, its productivity and cost.

Comparison of gas and steam turbine performance

The largest and most productive gas turbine with 60 cycles available today has a capacity of 250 MW and an efficiency of 40% LHV (lower heat transfer coefficient). An example of such a gas turbine is the Westinghouse, Model 501G. Parameters of steam turbines, for comparison, can reach 1,500 MW, and the efficiency exceeds 45% LHV. If we compare the power plants of the Rankine cycle with the power plants of the Brighton cycle, each of which uses the largest and most efficient turbines, we can conclude that one steam turbine of the Rankine cycle has a capacity of about six times higher and an efficiency of 12.5% more than the cycle Brighton using the best gas turbines. This fact can be cited as an advantage of the present invention, as shown later in this document.

Heat and energy production cycle

One of the features of a gas turbine is that it emits large volumes of exhaust gas at high temperature. During the Arab oil embargo in 1973 and rising prices in the energy sector, much attention was paid to finding ways to use the energy of high-temperature exhaust gases.

High prices in the early 1970s marked the wave of construction of small power plants, which used the principle of cogeneration. Cogeneration can be defined as the simultaneous production of mechanical or electrical energy together with thermal energy. In other words, the engine (gas turbine or otherwise) is used to produce energy, and at this time the excess heat received from the engine is used for another process, which replaces the fuel necessary for this process. In terms of fuel use, this method was very effective and was invariably promoted by the United States Public Utilities Policy and Regulation Act of 1978, which ordered local utilities to purchase electricity from qualified producers of associated electricity at a price that excluded the cost of new power plants .

At first, the stations for generating heat and electric energy were small, usually less than 50 MW. They consisted of small gas turbines with PGRT for steam production. In many cases, the vapor pressure was relatively low (less than 600 psi) since steam was used in the process. Some stations included a steam turbine, others did not. With the development of industry, larger power plants with a higher vapor pressure began to be designed to take into account the increase in lower cycle productivity. In addition, the largest gas turbine manufacturers have designed and built larger and more productive gas turbines to meet market demands for the simultaneous production of heat and electricity. Soon, due to high productivity, low emissions and low capital costs, cogeneration power plants gave way to combined cycle power plants (i.e. power plants that produce only energy and do not produce useful thermal energy). And until now, sometimes there are proposals for the construction of cogeneration power plants, only now they are usually called thermal power plants.

Although there was a gradual transition from small power plants that produce thermal and electric energy to large power plants with combined cycles, the overall system and method of generating electricity remained mostly unchanged. Gas turbines were the main engine, and the heat of the exhaust gases was used to produce steam using PGRT. Steam turbines optimized for maximum power generation generated additional electricity, which equaled approximately 50% of the electricity generated by gas turbines. PGRT was usually a boiler with two or three pressure levels to ensure maximum heat recovery, and the steam turbine was designed so that it could use steam of all pressure levels. A review of manufacturers of rated power units with combined cycles will demonstrate this trend. The 1997 Turbo Equipment Manual (1997 TURBOMACHINERY HANDBOOK / USPS 871-500, ISSN 0149-4147) describes rated combined cycle power plants offered by various manufacturers, including ABB, General Electric, and Westinghouse. In almost every case, the generated power of a steam turbine is approximately 40% to 60% of the power of a gas turbine. A 1996 General Electric informative document (General Electric informative document GER-3567G, 1996, GE Heavy-Duty Gas Turbine Performance Characteristics,) by Frank J. Brooks describes gas turbines used in power plants with a combination of cycles.

In short, most manufacturers of turbo-equipment have switched from manufacturing small cogeneration power units that produce thermal and electric energy at the same time, to manufacturing power units with combined cycles. The sizes of power units with combined cycles have grown from small cogeneration projects to generate less than 50 MW of electricity to large structured power units that generate more than 700 MW of electricity (as, for example, in the case of a power unit with combined cycles Westinghouse model 2X1 501G). Such power units are mainly gas turbine, where the steam turbine produces additional electricity, nominally constituting from 40% to 60% of the electricity generated by the gas turbine. Since gas turbines are the main engine, the requirements for the parameters of rated power units with combined cycles are very stringent, since gas turbines are in-line products, and steam turbines are often designed and manufactured according to customer requirements. The proposed power plant and a method of generating energy with a combination of cycles are characterized by more flexible approaches, while at the same time allowing you to save such advantages of power units with a combination of cycles, such as high productivity, low emissions and low capital costs, which will find a favorable response in the industry.

BACKGROUND

Performance optimization

Feed water heating

Power plants using the Rankine cycle generate billions of dollars of electricity annually and consume huge amounts of fuel every year, which leads to the fact that much attention is paid to development and analysis in the direction of optimizing the Rankine cycle by introducing some small changes to existing configurations. Figure 3 shows some general changes that are used to achieve optimal Renkine cycle performance. Reference numeral 303 in FIG. 3 is a schematic illustration of a first feed water heater. This device is usually made in the form of a shell and a tubular heat exchanger, but it can be a plate and a frame of a heat exchanger, a swirl mixer of a heat exchanger that mixes incoming water with small amounts of steam, or another device for providing heat exchange. As a result of studies, it was found that the use of steam obtained from a steam turbine for preheating water before it enters the boiler increases the cycle productivity.

The feed water heater uses steam coming from a steam turbine in the optimum section for pre-heating water between the condenser outlet (319) and the boiler inlet (306). A second feed water heater (305) is also shown in this example. The number of heaters and their optimal characteristics of the steam flow are determined, but not limited, by a number of indicators, including the pressure at the entrance to the steam turbine, the temperature at the entrance to the steam turbine, the conditions of steam overheating, the efficiency of the feed water heater, and other factors. Typically, the number of feed water heaters, their design and the parameters of the steam supplied to the turbine should be determined for each power plant individually, taking into account the variety of designs of each power unit and the operating characteristics.

Reheat

One of the varieties of the Rankine cycle is characterized by the use of reheating to increase the efficiency of the cycle. This option involves the expansion of steam in a steam turbine in the design mode from the inlet to some predetermined superheat pressure. At the same time, a certain amount of steam energy is already used for the production of mechanical energy (shaft power in hp). The spent low-energy steam is then returned to the boiler, where it is reheated. The reheated high-energy steam is again sent to a steam turbine to produce additional energy. In a cycle, reheating steam can be used more than once. Given the specified design conditions, pressure and temperature at the entrance to the turbine, as well as under other conditions, such overheating is used to increase the efficiency and effectiveness of the cycle.

Other factors

Other factors affecting the efficiency of the cycle are the pressure and temperature of the steam at the turbine inlet, as well as the exhaust pressure. Typically, an increase in inlet pressure and temperature provides a corresponding increase in efficiency. Lowering the exhaust pressure also leads to increased efficiency. The exhaust pressure is usually determined by external factors, such as the temperature of river water, air or other liquid used to cool the condenser. These factors set limits on the exhaust pressure, and the condenser and associated equipment are designed to meet these limitations and evaluate the required parameters — such as the size of the cooling devices, environmental factors and costs.

Design limitations

The pressure and temperature at the turbine inlet are usually determined by the developer. However, there are design restrictions. With an increase in the inlet pressure, the pressure on the boiler tubes, casing and internal parts of the steam turbine increases. The pressure is limited by the ability to produce equipment of the necessary strength or economic feasibility. In addition, at vapor pressures above 3206 psi, the coexistence of the liquid and gaseous state can no longer continue. This state is called the critical point of steam, since vaporization does not occur at a higher pressure. And for the production of steam at a higher pressure, the coexistence of water and steam is required, as well as a more complex supercritical boiler. With increasing temperature, the permissible pressure on the boiler tubes, casing and internal parts of the steam turbine decreases and approaches the existing restrictions, but traditional materials of the steam turbine quickly lose their properties even with a slight increase in temperature (50 degrees Fahrenheit - ° F). Conventional large steam turbines constructed with the latest technology have a maximum temperature limit of 1050 ° F at the inlet of the working fluid to the turbine.

Steam Cycle Optimization

After selecting the temperature and pressure in the boiler, it is necessary to optimize the duty cycle. A typical high-performance steam cycle involves the use of feed water heaters, a superheater, a steam turbine with intermediate overheating, boiler feed pumps and a condenser. The source describing cycle optimization is an information document issued by General Electric entitled “Optimization, Evaluation, and Testing of a Steam Turbine Cycle” by James S. Wright (General Electric Reference GER-3642E, 1996 / James S. Wright). This document describes the operating options for various cycle parameters such as pressure, temperature, number of reheats and number of feed water heaters.

Renkine cycle example

Figure 3 schematically depicts the Rankine cycle, including both heating the supply water and overheating. Such a subcritical Rankine cycle works by supplying water to the inlet of the boiler feed pump (301). Then, using the feed pump (301), the water pressure is increased to the value necessary to pump water. Water under pressure is supplied through a water supply line (302) to a feed water heater (303), and then through a water supply line (304) to a feed water heater (305). Feed water heaters (303, 305) preheat the water before it enters the boiler through the inlet pipe (306). Preheated water enters the boiler evaporator (307), where the water is additionally heated under pressure.

Steam leaves the evaporator (308) and continues to move to the superheater (309) and leaves line (310). Superheated steam is directed to the intermediate pressure zone of the steam turbine (311). Steam expands in the intermediate pressure zone along line (312) and then returns to the boiler overheating zone (313), where it is brought to the temperature at or close to the input temperature of the steam with the help of additional heat. This reheated steam is then sent via line (314) to the intermediate pressure section (PD) of the steam turbine to create additional pressure. In the PD section (315), the steam expands, producing mechanical energy. After that, the steam leaves the intermediate pressure section and through the bypass pipe (316) enters the low pressure section (LP) of the steam turbine (317).

Due to the large volume flows at low pressure in the operation of large aggregates, the low pressure section usually has two guide compartments, so the steam enters the middle of the casing and moves through the vanes forward and backward to produce more mechanical energy. Then this steam leaves line (318) and enters the capacitor (319). Condensed steam, bypassing the condensate collector (330), is sent along line (320) to the inlet of the boiler feed pump (301). To heat the feed water, steam from the intermediate and low pressure sections (321) and (324) of the steam turbine is directed to the feed water heaters (305) and (303), respectively, along lines (323) and (326). In these lines, as well as lines (322) and (325), non-return valves are used to prevent the outflow of steam into the steam turbine in case of emergency (for example, sudden shutdown), when the pressure in the turbine rapidly drops to the pressure level in the condenser. Such valves can be either open or closed and are intended only for safety. Steam from these steam extraction pipelines, heading into the boiler, preheats the feed water. At the same time, it condenses in the feed water heaters, and the condensate (327, 328) returns to the inlet of the feed pump (301). The mechanical energy produced by a steam turbine is converted into electrical energy by a generator (329).

For more powerful central power plants, the common pressure for subcritical cases is 1800 and 2400 psia. For supercritical conditions, pressures from 3500 psi or higher can also be used. The inlet steam temperature for most high-power steam turbines is limited to 1050 ° F for both inlet and superheated flows. However, some advanced-technology steam turbines use an inlet temperature of 1070 ° F for high-pressure inlets and 1112 ° F for overheating, as described in detail in a steam turbine published by General Electric under the title “Steam Turbines for ultrascritical power plants ”by Klaus M. Retzlaff and W. Anthony Rugger (“ Steam Turbines for Ultrasupercritical Power Plants ”by Klaus M. Retzlaff and W. Anthony Ruegger (General Electric Reference GER-3945,1996).

Rankine cycle performance

In the table in FIG. 4, for an example of a steam turbine with an efficiency of 90%, a comparison of the basic Rankine cycle (Option 1) is carried out (except for the boiler performance and interference standards for power), a cycle in which only reheating is used (Option 2, Option 3) , and a cycle using reheating along with heating feed water (Option 4 and Option 5). Examples are also included for varying inlet pressures and overheating (Option 3), as well as heating feed water (Option 5). Option 6 and Option 7 apply to supercritical steam cycle conditions. Option 6 reflects a supercritical steam cycle with ultra-supercritical (temperature at the inlet or superheat temperature above 1050 Fahrenheit) characteristics of the steam flow and double reheating (the steam is heated twice at two different pressure levels in the boiler). Option 7 is identical to Option 6 with the addition of heating feed water. For the purposes of this comparison, only two steam extractions were used, and it was assumed that the extraction pressure was at the level of cold reheat pressure and transient pressure (second cold reheat for supercritical cycles). The use of more feed water heaters will achieve greater cycle performance. General Electric Reference Paper GER-3642E, 1996, entitled “Steam Turbine Cycle Optimization, Evaluation, and Performance Testing” by James S. Wright (1996) by James S. Wright)) provides data on the selection of the optimal number of feed water heaters, indicating that a 1.5% penalty specific heat consumption was estimated based on three feed water heaters versus seven. Thus, there are opportunities for improving the performance of the feed water heating cycle shown in FIG. 4 (Options 4, 5 and 7). Using reheating, optimal heating of the supply water and ultra-supercritical conditions of steam flows, it is possible to achieve a total capacity of the power plant cycle of more than 45%.

The total capacity of the power plant cycle (efficiency) includes not only the performance of the basic steam cycle, as shown in FIG. 4, but also the productivity of the boiler and the norms of power interference in relation to, for example, boiler feed pumps and condenser pumps with circulating water . As stated in the magazine “POWER MAGAZINE” (ISSN 0032-5929, July / August 1998, page 26):

“Over the past few years, new developments have appeared to increase the productivity of steam power plants, and most of these developments are related to a steam turbine. Productivity at the level of 45% LHV (lower heat transfer coefficient) or higher is now possible with the advent of the latest steam power plants operating on fossil fuels using the highest steam parameters, advanced heating cycles of feed water, metallurgy achievements in the manufacture of boilers and turbines, etc. . ”.

Achieving a combined plant capacity of 45% LHV, including boiler performance and power interference standards, usually means that the basic steam cycle should even exceed 45%. With a boiler productivity of 85%, interference norms for power of 2.5%, a ratio of HHV coefficient (higher heat transfer coefficient) to LHV coefficient equal to 1.11 (which is typical for natural gas), and a power plant capacity of 45% (LHV) basic steam cycle productivity will be calculated as:

As shown in the table of FIG. 4, the use of a steam superheat cycle at pressure values according to the table can increase the base performance of the Rankine cycle by 4.79%. However, the use of overheating, as well as the use of increased inlet pressure and heating of the supply water, can increase the productivity by at least 10.3% for subcritical conditions of steam operation. (It should be noted that performance improvement is the ratio of the performance of a particular option to the base performance. Thus, a cycle with a 40 percent output converts 40% of the input energy into electricity. This is two times more than in the case of a 20 percent cycle. Therefore, performance improvement from 20% to 40% is 100% or double the amount of energy generated).

Fuel efficiency is critical for power plants, and powerful central coal burning plants can spend about $ 140 million annually on fuel based on a power plant capacity of 1000 MB, a thermal capacity of 45% LHV, $ 2.00 per million BTU (BTU) for fuel and 8500 working hours per year. Based on this, even a 1 percent increase in productivity will mean greater fuel savings ($ 1.4 million annually).

Combined cycle

Although the Renkine cycle has proven to be effective, today's more stringent energy and environmental standards require more attention to be paid to fuel productivity and reducing emissions from power plants. This has led to the need for the development and construction of new combined cycle power plants.

Figure 5 shows a conceptual diagram of the application of the principle of combining cycles. In general, a combined cycle is not limited to combining the upper Braithon cycle and the lower Rankine cycle, but may be any combination of cycles. The upper and lower cycles can be the same cycles using different fluids. In any case, the circuit shown in FIG. 5 is applicable. According to the diagram of Figure 5, the upper cycle fluid (501) enters the upper cycle engine (502), where fuel (503) is added to increase its temperature. The fluid performs the work, which is converted by the engine of the upper cycle into mechanical energy (horsepower on the shaft). This energy drives the load of the upper cycle (504). This load can be an electric generator, pump, compressor or other devices that require the application of mechanical energy. The liquid that leaves the upper cycle engine is routed through the exhaust pipe (505) to a heat recovery device (506), and then enters an open tank (507).

In this example, the top loop is an open loop. In other words, the upper cycle fluid comes from a large reservoir and is discharged into the same reservoir. A heat recovery device (506) absorbs part of the energy of the exhausts of the upper cycle and transfers it to the liquid of the lower cycle (508). In this example, the lower cycle fluid is heated at three separate pressure levels: on the high pressure line (509), the intermediate pressure line (510), and the low pressure line (511). These fluids then enter the lower cycle engine (512), where mechanical energy is generated to drive the load of the lower cycle (513). And again, an electric generator, pump, compressor, and other devices that require mechanical energy to function can act as a given load.

From the lower cycle engine, the lower cycle liquid enters the heat exchanger (514), in which heat is discharged. The lower cycle fluid then enters a pump or compressor or other fluid transfer device (515), from where it then returns to the heat recovery device (506). For this example, it is taken as the condition that the lower cycle is a closed cycle, which means that the liquid of the lower cycle is constantly circulating in a closed loop. In this process, you can use more than two cycles, and any of the cycles can be either open or closed loop. This characterizes the main features of the application of combining cycles.

PGRT in combined cycles

In the modern construction of cogeneration and gas / steam turbine power plants, the concept of a combined cycle power plant has come to mean using the Braithon cycle as the upper cycle and the Rankine cycle as the lower cycle. Such power plants use gas turbines or combustion turbines with a boiler at the outlet of a gas turbine as the primary propulsion mechanism (Braithon cycle mechanism) to regenerate the generated heat. This boiler is commonly referred to as a waste heat boiler or a heat recovery steam generator (PGRT). It may also include burners to increase the temperature of the exhaust gases and generate more steam than just the waste heat (in addition to burning fuel). PGRT produces steam, which is then fed to a steam turbine for energy production. Due to the high fluid temperatures in the gas turbine (approximately 2400 ° F for General Electric process “F” class devices and 2600 ° F for Westinghouse process “G” class devices) and waste heat recovery, combined cycle power plants have more higher fuel productivity than traditional steam power plants. In addition, with the advancement of gas turbine technology and the use of distillate oil products or natural gas fuels, combined cycle power plant emissions are not very significant. 6 shows a typical application of combining cycles.

PGRT is significantly different from a traditional boiler operating on the Rankine cycle. The Renkine cycle boiler runs on various fuels, including oil, natural gas, coal, biomass and others. In the boilers of the Renkine cycle, you can also use a combination of different types of fuel. PGRT in its work cannot use any type of fuel, but can only capture and use the waste heat of a gas turbine. If complementary fuel combustion is used, PGRT will require more refined fuels, such as natural gas or distillate oil products. Solid fuels, such as coal and biomass, are not commonly used in these types of boilers.

As can be seen from Fig.6, before PGRT there are several sections, including three sections of evaporation (one for each pressure level), economizers, superheaters and reheaters. Sections (601) and (602) are economizers. They are large tubular compartments in the HRSG, which preheat the water before converting it to steam in the evaporator. Compartments (603), (606) and (609) are low, intermediate and high pressure evaporators, respectively. Sections (604), (605) and (607) are feed water heaters. Section (608) is the superheater of the intermediate pressure evaporator, and sections (610) and (612) are the superheaters of high pressure evaporators. Section (611) is a re-heater compartment. Such PGRTs are usually large and large and have literally millions of tubes inside.

Steam from each pressure level is used at the power plant where it is required, but mostly most of the steam is used to produce additional electricity in a steam turbine. This means that steam at lower pressure levels must be introduced into the steam turbine in an appropriate place other than the steam inlet at high pressure. This also means that the steam turbine must have special devices (openings, branches, connections, valves, etc.) through which the steam can flow, and that in the operating mode the steam pressure in the steam turbine at the joints must be less than the steam pressure from the corresponding sections of the boiler PGRT. Otherwise, steam will not enter the steam turbine.

As can be seen from FIG. 6 when compared to FIG. 3, the traditional Rencine cycle involves the use of feed water heaters that draw steam from a steam turbine to pre-heat the feed water, while PGRT uses the heat generated by the gas turbine to provide this function. Therefore, conventional steam-powered feed water heaters are not commonly used in combined cycle systems. In the General Electric company information document entitled “Steam turbines for power plant systems with combined STAG cycles”, M. Boss (GE informative document GER-3582E (1996), entitled “Steam Turbines for STAG ™ Combined Cycle Power Systems”, M. ) it is confirmed that the feed water heaters were not used within the framework of the prior art:

“Sizing is important for any steam turbine, but it is especially important for combined cycle systems. There is usually no extraction of steam from a steam turbine, since heating of the supply water is usually provided inside the HRSG. ”

Another option commonly used in combined cycle systems is to use two boiler pumps (630) and (631), which are commonly referred to as low and high pressure boiler feed pumps, respectively (LP-BFP; HP-BFP). This arrangement allows the low pressure pump to supply pressurized water to the low and intermediate pressure levels, and the high pressure pump to supply water for the high pressure level, which saves the mechanical energy of the pump. For high-power systems with a combination of cycles, the location of the steam turbine / condenser is similar to their location in systems operating according to the Rankine cycle shown in Figure 3 (although, from the point of view of the internal circuit, the steam supply routes are completely different).

PGRT / Combined Cycle Disadvantages

Common disadvantages

In modern technology, the maximum inlet pressure of a steam turbine with a combination of cycles is nominally 1800 psia, and the temperature of the incoming steam reaches 1050 ° F for inlet and superheated steam. Some of the disadvantages of this organization PGRT systems with combined cycles are:

1. The productivity of the steam cycle is significantly lower than the performance of traditional steam power plants.

2. Numerous evaporation compartments are required to maximize heat recovery. This leads to an increase in equipment and maintenance costs.

3. Numerous evaporators require the involvement of power plant operators and the organization of control systems to monitor and control boiler cylinder levels.

4. PGRT with numerous compartments are large, they require the creation of a branched infrastructure of large-capacity buildings, many areas and powerful foundations to support the mass of the generator.

5. HRSGs are expensive (approximately US $ 10 million for a HRSG sufficient to recover the heat of one GE Frame 7 gas turbine).

6. With an increase in the number of components, evaporator sections, control and other devices, maintenance requirements increase.

7. Low pressure steam (ie steam other than the one characterized by the highest pressure) has a significantly lower ability to produce energy in a steam turbine than higher pressure steam.

8. Partial load, operations not stipulated by the project, and other conditions, along with those provided in the project, usually lead to less heat recovery and lower cycle productivity.

9. An increase in the number of PGRT tubes in order to increase heat recovery leads to additional restrictions on the movement of the exhaust stream of the gas turbine, which creates an increased back pressure that reduces the productivity and efficiency of the gas turbine.

10. The temperature levels of the exhaust gas turbine are insufficient to provide an increase in some parameters of the operating modes of steam, currently used in advanced steam cycles (600 ° C degrees Celsius, which is equivalent to 1112 ° F).

11. The existing problems in reheat lines with several gas turbines (usually three or more) make it difficult to use large steam turbines in power plants with combined cycles in the current state of the art. For modern, powerful and efficient combined cycle power plants such as the GE S207FA, steam turbine power typically reaches 190 MW, which is much lower than General Electric steam turbine power, which can exceed 1200 MW. For more information on large steam turbines, please refer to the information brochure issued by General Electric (GE) entitled “Steam Turbines for Large Power Plants” by John K. Reinker and Paul W. Mason (“Steam Turbines for Large Power Applications "by John K. Reinker and Paul B. Mason (General Electric Reference GER-3646D, 1996)).

Part-load inefficiency

Another drawback of systems with combined cycles is their operation under partial load. When the system to which the power plant is connected reduces load requirements, the power plant must respond by reducing production. Such load modulation provides for maintaining a constant speed of the equipment and a constant industrial voltage frequency (i.e. 60 Hz in the USA and 50 Hz in Europe). To modulate the load in a power plant with combined cycles in a gas turbine, less fuel is burned and energy production is reduced. This usually requires a decrease in the combustion temperature in the gas turbine and / or a decrease in the amount of air entering there.

Partial load operation reduces the efficiency of the gas turbine, thereby reducing the efficiency of the entire power plant with combined cycles. 7 shows a typical curve for a large modern gas turbine with inlet guide vanes for modulating the intensity of air intake into the inlet. Even with increased productivity under partial load conditions provided by inlet guide vanes at a load of 60% (Generator output - as a percentage of the projected amount), a gas turbine consumes more than 70% of the fuel required to operate at full load (Heat Consumption - percentage of the projected). This represents an increase of 17.5% in the amount of heat (specific heat consumption). For gas turbines without inlet guide vanes, this decrease in productivity will be more significant.

To try to avoid this kind of load reduction, as well as provide greater energy production for a given amount of hardware (sometimes called power density), manufacturers can install two gas turbines in power plants with combined cycles, each of which has its own HRS connected to one steam turbine (so-called “2-in-1 design”). With this design (2 gas turbines and 1 steam turbine), when the power plant load is reduced by more than 50%, one gas turbine can be turned off, and the other can return to production of about 100%. This mode of operation increases the capacity of the partial load below 50% of the total load of the power plant, as shown graphically in Fig. 8. The graph shows a typical comparison of two gas turbines, taken from the General Electric information brochure. Production line and performance with a combination of General Electric cycles by David L. Chase, Leroy O. Tomlinson, Thomas L. Davidson, Raub W. Smith and Chris E. Maslak (1996) (GE informative document GER-3574F (1996), entitled "GE Combined-Cycle Product Line and Performance" by David L. Chase, Leroy O. Tomlinson, Thomas L. Davidson, Raub W Smith, and Chris E. Maslak) on the partial load curve for a combination of cycles and a 2-in-1 design. switching from three to two gas turbines with a load of slightly less than 67%. Nevertheless, this still provides a significant increase in the amount of heat under partial load conditions. It should be noted that such an increase in partial load efficiency is the result of higher equipment costs. In the framework of the existing level of technology it is necessary to solve the problem of efficiency without the need to purchase additional equipment, thereby increasing the total cost of the power plant.

Additional combustion of PGRT fuel

Another solution to increase the flexibility of the power plant with combined cycles is the use of additional fuel combustion in PGRT. This mode of operation of a gas turbine is achieved when the fuel is burned in a heat recovery steam generator (PGRT) immediately after the gas turbine (or at some intermediate stage inside the PGRT). This leads to an increase in the temperature of the exhaust gases entering the HRSG and the amount of steam that can enter the steam turbine. This allows the power plant to produce more energy. However, the amount of heat consumed is growing and fuel efficiency is correspondingly decreasing. Such a result is described by Moore, “General Electric,” in US Pat. No. 5,649,416. This patent, as well as US Pat. No. 5,428,950 to Tomlinson, are mentioned by Rice in US Pat. increase energy production, but at the expense of lower productivity. In the General Electric information brochure (1996), General Electric Production Line and Combined Cycle Productivity by David L. Chase, Leroy O. Tomlinson, Thomas L. Davidson, Raub W. Smith and Chris E. Maslak (GE informative document GER-3574F (1996) entitled "GE Combined-Cycle Product Line and Performance" by David L. Chase, Leroy O. Tomlinson, Thomas L. Davidson, Raub W. Smith, and Chris E. Maslak) states that “the increasing productivity for energy generated by additional combustion is 34-36% based on the lower calorific value of the fuel (LHV)”.

The same brochure in Table 14 indicates that additional fuel combustion in PGRT can increase the amount of product produced in a power plant with combined cycles at the current level of technology by 28%, but only with an increase of 9% in the total amount of heat in the combined cycle ( specific fuel consumption). In the current level of technology, no technological approaches have been developed in order to avoid heat losses associated with additional combustion.

In addition, additional fuel combustion in the framework of the prior art can be used to achieve higher ratios of the ratio of steam and gas turbines than with traditional combined cycles. However, operating at such high power levels of steam / gas turbines in order to satisfy the highest energy needs usually has a short duration, and long-term operation with such coefficients is not economical. Therefore, traditional power plants with combined cycles that were designed using steam / gas turbine ratio close to the same do not work like power plants of the Rankine cycle only, but operate in this mode to provide a temporary peak load, which always results in significant damage work performance.

Decline in gas turbine performance

As indicated in the discussion section of the Brighton cycle, approximately 2/3 of the energy generated by the turbine section of the gas turbine is needed to drive the compressor compartment, and the remaining 1/3 of the energy is enough for load driving operations. Such energy consumed by the compressor at 67% of the turbine power is significantly higher than in the example with the Rankine cycle, where the boiler pumps consume only 2% of the turbine power. Therefore, a gas turbine is prone to decline in operation if the compressor does not maintain an optimal level of performance.

For example, the typical capacity of an axial-flow compressor used on a large gas turbine can be 90%. Therefore, if the compressor requires 67% of the output power of the turbine compartment, only (0.67 × 0.90) = 0.603 or 60.3% would be the ideal power level (100% capacity). If the compressor capacity drops by 2.5%, its new capacity will reach (0.90 × 0.975) = 0.8775 or 87.75%. The required compressor power in this case would be (0.603: 0.8775) = 0.6872 or 68.72%. The net energy output of the turbine decreases from 33% (1.00-0.67) to 0.3128% (1.00-0.6872). This amounts to a loss of 5.2% in productivity (0.3128: 0.33 = 0.9479). Therefore, it is obvious that a slight decrease in the performance of a gas turbine compressor leads to a large decrease in the productivity and power of a gas turbine.

A 5% loss in productivity compared to the example above is typical of many gas turbines after a period of one to two years. Such a drop in productivity is mainly due to the deterioration of the compressor and erosion of the tips of the compressor blades. New blades and seals usually help restore compressor performance to the performance level of a new compressor. However, such repairs are expensive, time consuming and can be carried out after major inspections, which are planned every four years for modern gas turbines. Therefore, owners of power plants and operators should take into account the decline in productivity between the main overhauls of gas turbines.

Opportunities for Improvement in the Prior Art

It can be seen from the above that the parameters of certain modern technologies that can be improved are as follows:

Flexibility

Based on the needs of the load in the production of electricity in a particular region or market, a public power plant (distributing electricity among users) determines the demand for electricity taking into account current and expected needs. For example, if the expected load is 850 MW, in the traditional configuration according to the Rankine cycle, the utility / electricity company will contract with an architectural and engineering firm to develop and build such a power plant.

The boiler, pumps, condenser, steam turbine and all other auxiliary devices of the power plant are designed for a given power of 850 MW. Mostly the desired result can be achieved due to the fact that steam turbines are designed and manufactured in accordance with the needs of a particular customer. However, due to the fact that gas turbines are produced by the conveyor method, and the combined cycle is introduced at power plants using gas turbines to achieve the highest productivity and lowest capital costs, the public service and / or the electricity supplier cannot only declare the required level of power generation , but also must find the best option to meet their needs, based on existing proposals for combining cycles of different manufacturers. For example, an analysis of available power plants with combined cycles from the 1997 TURBOMACHINERY HANDBOOK BROCHURE (USPS 871-500, ISSN 0149-4147)) indicates that no power plants with combined power cycles are available on the market for 60 Hz 850 MW. Thus, the flexibility of a power plant project developer is limited by the capabilities of modern technology for combined cycle power plants. This suggests that in some circumstances a set of equipment for installation at a particular power plant will not be optimal due to the limitations imposed on the level of modern technology on the configuration of equipment for power plants.

Performance

Combined-cycle power plants are characterized by very high power generation in comparison with other traditional means of generating electricity. However, a large central power plant with combined cycles with an estimated capacity of 1000 MW at 55% of the LHV thermal efficiency (lower heat transfer coefficient), operating 8500 hours a year at full fuel load of $ 3 per million BTU of fuel, spends about 175 million US dollars annually for the purchase of fuel. Increasing productivity by even 1 percent will result in large fuel savings ($ 1.75 million annually).

US Pat. No. 4,333,310 to Robert Uram uses a control method that monitors the temperature of a steam in a steam turbine and controls the operation of the waste incineration chamber (afterburning) to control the temperature of superheated steam. On the one hand, the temperature of the inlet of the steam turbine is maintained at an optimal level, and on the other hand, this function does not significantly affect the load. In this patent, Yuram claims:

“It is desirable that the steam turbine operates in a so-called“ follow turbine ”mode, in which the power plant supplies electricity for the load so that the steam turbine follows the gas turbines and the afterburning positively follows the corresponding gas turbine. In other words, the heat generated by the afterburning follows the temperature of the exhaust gases of the gas turbine, and the steam obtained from the gases produced by the afterburning is used together by the steam turbine. ”

These lessons of the prior art are in direct contrast with the present invention, in which the heat contribution from additional fuel combustion is independent of gas turbines, and gas turbines are designed to operate at their full optimum power level.

Set value

The next major cost after fuel costs for combined cycle power plants is usually debt servicing. Manufacturers, engineering firms and owners are usually interested in reducing the cost of installing power plants. With 8% accrued interest and $ 450 per kilowatt of power plant with combined cycles of 1000 MW, debt servicing will be approximately $ 45 million per year for 20 years. Reducing production costs while fully utilizing production capabilities, in terms of cost / energy ratio, directly reduces debt service costs.

Temporary increase in peak output

The dilemma faced by power plant owners and utilities is the right choice of power plant capacity. Choosing a power plant that is too small leads to a shortage of electricity, a partial disruption of power supply and / or to the need to purchase expensive electricity from other manufacturers. Choosing a power plant that is too large will result in low productivity during part-time periods and an increase in capital costs per kWh of electricity generated. In many situations, the challenge for power plant designers is the need to provide both the ability to meet demand at the peak of power consumption and consumption under temporary loads. Peak consumption occurs only at certain times of the year for a limited period of time. Typically, peak hours on the hottest days of the summer months in terms of load are the most critical time for power producers and their systems. The ability to provide excess capacity during this period is highly desirable, and with emerging trends in the privatization of the electricity market, this can also be very useful and attractive.

For example, in the early summer of 1999, a lack of electricity in the northeastern United States raised concerns about the system's ability to meet the largest electricity demand. Some local news programs reported a cost of $ 30 per MWh during normal periods and up to $ 500-1,000 per MWh at peak season. However, as reported in POWER MAGAZINE (ISSN 0032-5929, March / April 1999, page 14), even greater costs were incurred: “Reserve stocks declined across the country from 27% in 1992 to 12% in 1998 based on Edison Institute of Electricity’s information in Washington, DC, as volatility due to privatization has caused a suspension in capacity expansion. A significant increase in prices in the Midwestern United States to $ 7,000 per MWh last summer was the focus of the press, but price increases to $ 6,000 MWh also took place in Alberta ... ”

However, ensuring the production of electricity at the peak of the season will not be attractive if the owners of the power plants are forced to pay for such capacity, as well as for servicing the debt, and also provide income for this additional capacity only for several months a year. Therefore, power plants are required that can generate more electricity than normal during peak season to help ensure that the system is at peak load.

FIG. 31B is a graph showing the relative amount of time that a typical power plant operates during the peak season, during the intermediate load period, and under basic load conditions. It can be seen from this graph that a power plant project for operation at peak congestion will not be profitable, since such a period accounts for less than 10% of the total time. Due to the fact that typical power plants of the current level of technology are not capable of various options for operating at peak power, the only practically applicable option for modern electricity producers is to purchase energy in excess of the plan in the season of the highest demand for it. The present invention provides a system and method that can satisfy these needs without purchasing additional electricity from the outside, and which are thus more economical in comparison with those known in the art.

Current reserve

The issue of increasing electricity production at peak season affects the concept of a working reserve. Working capital reserve meets the need of local public utilities to regulate overcoming power generation failures in crisis conditions through the use of spare fast-rotation generators operating in an automatic non-autonomous mode, but at the same time not generating energy. With this construction, if the electricity producer suffers losses of production capacity, the loaded reserve mode is automatically put into effect, compensating for the losses and, thus, preventing the collapse of the power system. With the privatization of enterprises, which is becoming more commonplace in the US electricity generation industry, the question of who provides the working capital reserve is crucial. In the era of tight regulation and monopolies in the energy industry, a monopolist representing the public service was required to ensure an automatic non-autonomous regime of the working reserve. In the era of privatization, the issue of paying for such collateral capacity is very important.

The present invention solves this problem by providing a significant increase in the capacity of a power plant by the public utility immediately in the event of equipment malfunction or failure. For example, according to the present invention, it is possible to ensure the operation of nine power plants with a capacity at peak load of up to 111% and compensate for the failure of one power generation station. In a number of cases, this would allow planning the energy system without a working reserve or with a significantly reduced rotation reserve, which would lead to a decrease in the total cost of electricity production as a whole.

Non-local power generation / distribution reliability issues

One of the serious problems of the current level of technology is that the capacity of a power plant mainly exists in a relatively constant and narrow range of power generation control. When the peak demand for electricity is superimposed on the grid plan, it is necessary to purchase it in other places where the demand for electricity is lower than the capacity of a remote power plant. With this mode of providing peak electricity needs, that is, by redirecting the power of remote power plants, there are several main problems.

Firstly, losses associated with the transfer of electricity from remote locations to the place of electricity demand. For example, one hot day in New York City may require the transfer of electricity from Canada or the western US states, which can lead to significant linear losses in power lines.

Secondly, there is uncertainty about the reliability of the acquisition of electricity in remote areas of the energy system during the period of the highest congestion. Despite the fact that, in principle, redistribution of electricity is possible, it leads to instability in the power system. Small malfunctions can occur in the entire power system, which will lead to an additional failure of the equipment or cause instability of the voltage in the network. Such events, which seem insignificant in remote regions of the country, can lead to serious accidents in large regions of the capital level through electric networks.

Thus, taking into account the above reasons, in general it is always considered the best opportunity to generate electricity on the spot and on demand. The current level of technology is based on the ideas of separation and distribution of electricity, while the present invention is based on the idea of developing a more reliable method for local production of electricity in an electricity production system, which is more efficient and reliable than known from the prior art. It must be emphasized that the limitations inherent in the current level of technology in terms of peak power plant performance generally impede the local production of energy required during peak seasons. This forces traditional power plants to incur significant costs (from 10-fold to 250-fold) when purchasing electricity from remote stations.

Operating and Maintenance Costs

The cost of labor, fuel, maintenance, water, chemicals, spare parts and other consumables, including other expenses for paying taxes and insurance, are part of the overall costs of operation and maintenance. With an increase in the size of a power plant, the number of units of equipment also increases, and with an increase in the number of units of equipment, the indicated costs increase as a whole. In the search for greater productivity, a more advanced and expensive gas turbine technology is used. Maintenance costs associated with the use of new exotic materials, complex vanes and hardware are planned to be significantly higher than those that would be required for the slightly less efficient hardware listed above, proven gas turbines and designed for them electricity generation systems.

In order to be prepared for equipment malfunctions, owners of the power plant must have a large number of spare parts in stock. This leads to stockpiling and, consequently, high costs in terms of unused capital and taxes. Owners and operators of power plants are always committed to reducing operating and maintenance costs.

Gas compression

According to current forecasts, stable reserves and the price of natural gas will remain until 2010. This type of fuel is clean, efficient, inexpensive, and thus preferred for combined cycle power plants. However, if the power plants are located away from the main gas pipelines, the natural gas coming under reduced pressure must be compressed to achieve a pressure sufficient for use in a gas turbine. In addition, high-performance gas turbines, such as the Westinghouse 501G model, require higher natural gas pressures than gas turbines with lower pressure ratios, such as the General Electric model PG7241FA. Requirements for a higher natural gas pressure necessitate the use of expensive natural gas compressors and are mandatory for the operation of a power plant (operation of a power plant is impossible without them). These compressors require constant maintenance, as well as consume stray energy (the energy needed to operate the compressors reduces the net power supplied to the electrical network). Reducing the need for such components reduces the cost of installing a power plant, the requirements for real estate, improves reliability and increases the amount of generated useful electricity.

Power Plant Reliability

The issues of energy reliability have become an aspect of interest for both individual consumers of electricity and industrial users.

Therefore, the technology of power generation must be stable and reliable. US Pat. No. 5,628,183 to Rice proposes a power plant system with a combination of high capacity cycles. However, this system involves the use of divertors in PGRT, and installations for the reforming of natural gas and superheated steam to 1400 ° F. Such systems will significantly increase the cost of installation and the cost of operation and maintenance. In addition, to date, boiler valves, PGRT and steam turbines have not demonstrated their long-term reliable operation at temperatures above 1150 ° F, and PGRT with divertors and natural gas reforming do not yet have a stable reputation in the market.

Air consumption

Gas turbine engines consume large amounts of air. Typical combined cycle plants consume about 20 Ibs. air per kilowatt of electricity generated. This corresponds to approximately 260 cubic feet (at sea level) per kilowatt. At the same time, air must be passed through the filter before it enters the gas turbine, in order to avoid malfunctions that may occur if foreign objects enter the gas turbine. From time to time, air filters must be cleaned and / or replaced. This increases the costs of operation and maintenance, as well as the downtime of the power plant (the time when the power plant is not working and does not produce electricity).

In addition, the air consumed by the gas turbine enters the HRSG and then enters the atmosphere in the form of exhausts. With the consumption of more air, more air must enter the atmosphere as exhaust. This results in a loss of performance, since the temperature of the HRSG exhaust is typically 180 ° F. In addition, the emitted air stream heats the atmosphere and causes local problems with air quality.

Power Station Emissions

In order to obtain a permit for operation, the power plant must first obtain a license. The latter usually sets acceptable levels of pollutant emissions by the power plant. Combined-cycle power plants, in comparison with other traditionally used systems, are very clean energy producers, but their “scourge” is usually a pollutant with a set maximum permissible content - nitrous oxide (NOX). The latter is usually controlled by injecting steam and / or water into a gas turbine, dry low-combustion NOX systems and / or subsequent purification of exhaust gases. As the latter method, the so-called “selective catalytic reduction” (SCR) is usually used, which consists of injecting ammonia (NH ₃ ) into the exhaust stream in the presence of a catalyst at a certain temperature level to reverse convert the NOX obtained during combustion to N ₂ and H ₂ O.

In Patents 3,879,616 to Baker et al., 4,578,944 to Martens et al. And 5,269,930 to Finckh et al. The load of a power plant is controlled by changing the output of a gas turbine. However, in the case of a partial load of a gas turbine, NOX emissions usually increase. Therefore, in order to reduce emissions, it may be necessary to introduce more ammonia into the exhaust gas. This leads to an increase in operating and maintenance costs and can be such a serious obstacle that at the stage of development of a power plant, the desired gas turbines cannot be used due to the high level of emissions during operation under partial load conditions. Also, when operating at full capacity, some power plants may not need VKS, and at a partial load, VKS will be mandatory. Another factor associated with air emissions is air consumption. Gas turbines need a lot of air, and the more air is consumed, the greater the possibility of emissions into the atmosphere.

It should also be noted that scientists studying the current global warming trend have found an increase in carbon dioxide over the past century, and that since the beginning of the technical revolution, there has been an increase in CO ₂ levels of 1-3 quasi-peak levels per year, and today's level exceeds 350 quasi-peak levels. Such an increase in CO ₂ content is believed to directly affect the global environment, and also has a detrimental effect on the global climate. Scientists and governments are currently preoccupied with determining what CO ₂ restrictions should be imposed to solve this problem. Based on this trend, power plants of the future should have the highest possible efficiency in terms of not only NOX production, but also CO ₂ . The present invention pays particular attention to issues of such effectiveness.

In conclusion, it should be noted that there are certain significant legal aspects (including fines and sanctions) for power plants that violate clean air regulations in various states, mainly in the United States. See the article “Exploring 'New Sources'” in the National Law Journal of August 23, 1999 (Jonathan S. Martel, '"New Source' Scrutiny", THE NATIONAL LAW JOURNAL, at B6 (August 23,1999)).

Environmental issues

In addition to air emissions, other environmental issues arise. To operate a steam power station, a source of clean water is needed to provide make-up water. Such water is used to recover steam / water lost as a result of leaks, discharge (condensate, sludge) or other reasons. Water is discharged, which is released from the compartments of the PGRT evaporator and discharged into the sewer. Such water is usually taken from the bottom of the PGRT to remove feed water, which has a high content of minerals and sediments. This process helps keep the steam passage clean, provides minimal deposits in the steam turbine and blade malfunctions due to breakdowns caused by corrosion. The discharged water must be discharged into rivers, canals, etc., which requires permission, the receipt of which from government bodies is a difficult and lengthy process.

Power Plant Control Distribution System (SRK)

Modern combined cycle power plants typically use a control distribution system (DCS) to control the entire power plant. Such dispensers are integrated with individual control systems for steam and gas turbines. With the help of IBS, it is possible to observe and monitor and control many other parameters. The use of control to improve performance or operation is described in US Pat. Nos. 3,879,616 to Baker et al., 4,201,924 to Uram et al. And 4,578,944 to Martens et al. None of these patents, however, provide heat transfer control in PGRT. In US Pat. No. 5,269,930 to Finckh et al., A method for utilizing excess heat in a HRSG is used to operate under a partial load of a gas turbine. Such a method, however, does not provide complete control, but is only a means for the regeneration of low-temperature used heat. None of the aforementioned patents has proposed a method for controlling the temperature of the exhaust gas PGRT to the optimum temperature.

Energy Efficiency

Power plants with combined cycles within the framework of the current level of technology, designed to operate with maximum efficiency, usually use HRSG with different pressure levels and usually operate at three pressure levels. For each PGRT, as well as for each pressure level, operators have to monitor the level of the separator drum of the boiler. It is also necessary to monitor parameters such as water quality and chemical content for each HRS. Since the system load for each public power plant is different, it is necessary to use power plants with a combination of cycles, like other power plants, in accordance with the load distribution, or ensure the load in accordance with the needs of the power supply system. This means that power plants will not work with a fixed load, and their load requirements will constantly change to meet the needs of the system. To increase the load, you can use additional fuel combustion (additional fuel is burned in / or near the inlet in the HRSG to replenish the exhaust gas energy). However, this is very detrimental to the overall efficiency of the power plant. This is noted by Rice in Patent 5628183 in connection with analysis by Westinghouse and General Electric. Myp (Moore) in patent 5649416 indicates that

“The additional combustion of fuel from a steam heat recovery generator can increase the total amount of energy generated and the total amount of energy produced by a steam turbine, but only if the total thermal efficiency of the power plant decreases.” Therefore, usually in combined-cycle power plants, additional combustion is not often used or therefore, to change and correspond to various system loads, gas turbines are transferred from the operating mode with lnoy load operation to partial load.

In addition, as noted above, operation under partial load leads to an increase in air emissions, which also adversely affects performance. Figure 7 shows a curve of the ratio of the productivity of a gas turbine to its load. At 100% load, it consumes 100% of the fuel, but at 60% load it consumes 70.5% of the fuel consumed at full load. This represents a 17.5% increase in the consumption of a particular fuel. In the case of large central power plants, this corresponds to a significant increase in additional fuel costs. In addition, operation under partial load of a gas turbine usually leads to an increase in NOX emissions, the most difficult to meet the requirements for acceptable levels. The operation mode at partial load of the gas turbine also changes the flow of exhaust reptiles through PGRT. Such a change in flow negatively affects the heat transfer to the HRSG, since this device is designed based on the calculation of the fixed surface area of the heat exchanger. Such a phenomenon, together with a decrease in the efficiency of a gas turbine, is a factor that also adversely affects the efficiency of a partial load power plant. Since partial load operation significantly changes the temperature level in PGRT, this can lead to inefficient use of IBS.

Steam turbine exhaust end load

In addition to restrictions on pressure and inlet temperature, another common restriction for a steam turbine is the end load of the exhaust. Generally speaking, it is a function of two parameters: the end flow of exhaust gases and exhaust pressure. These two factors determine the volumetric flow rate of the stream passing through the last compartment of the blades of a steam turbine. For optimum performance, steam turbine manufacturers usually indicate volumetric flow rates. As volumetric flow increases, additional, larger blades and / or more exhaust compartments may be required.

However, due to mechanical restrictions (centrifugal force), after reaching the maximum volumetric flow rates of the restrictions for the largest blades, it is possible to add additional compartments and blades to the exhaust outlet of the steam turbine to ensure uninterrupted flow. This increases the installation cost and property requirements for the steam turbine. Due to its configuration, the traditional combined cycle sends high pressure steam to the high pressure inlet of the steam turbine, then adds the steam coming from the intermediate pressure compartment of the HRTG to this stream at the inlet of the intermediate pressure compartment of the steam turbine, and then adds more steam from the low pressure compartment of the HRTG to this stream in the inlet of the low pressure compartment of the steam turbine.

Therefore, in this embodiment, relatively smaller flows pass through the high and intermediate pressure compartments of the steam turbine and they have lower performance than the low pressure compartment. This option as a result leads to the operation of steam turbines in the mode when the exhaust level reaches a limit or approaches the limit of the end limit of the exhaust of a steam turbine. This ensures that there is no interference for a temporary increase in power to produce energy at peak load, and also leaves little or no room for improving the technical characteristics of a steam turbine in the future to achieve higher energy parameters. In general, this type of steam turbine is less efficient than traditional steam turbines for power plants, since the high and intermediate pressure compartments have low volumetric flows.

In the General Electric company’s brochure GER-3582E (1996) entitled “Steam Turbines for STAG Combined Cycles” by M. Boss (“Steam Turbines for STAG ™ Combined Cycle Power Systems”, by M. Boss), the load of the end exhaust of a steam turbine is considered, which reflects the current level of technology:

“The exhaust size is important for any steam turbine, but especially important for combining cycles. Usually, there is no extraction of a steam turbine, since the heating of the feed water is usually carried out inside the HRSG. The generation of steam at various pressure levels (supplying an intermediate level of pressure and / or low pressure to the throttle valve) increases the mass of the stream while the steam expands inside the turbine. The mass flow at the exhaust of a unit of the combination of cycles in a system with three pressure levels can be higher by 30% than the flow before the throttle. This is in direct contradiction with most links with boilers with heat input from fuel combustion, where the mass of the exhaust stream is about 25-30% less than the mass of the stream before the throttle due to extraction from the turbine for many stages of heating the feed water. ”

Real estate

The combined cycle plant, although usually smaller in comparison with traditional steam power plants, nevertheless occupies a large area. HRSG with risers are especially large and require a lot of space (HRSG for one Westinghouse 501G gas turbine model is approximately 40 feet wide, 70 feet high and 200 feet long). Given the trend towards privatization of the electricity generation sector, power plant owners will look for the ideal place for their power plants. In many cases, these places are located close to places of consumption, which are usually located in an urban or industrial area. Thus, power plants are located close to the final consumer of electricity, and they do not require high-voltage transmission lines (which also requires a large area for the placement of real estate). However, in such areas it can be difficult and expensive to acquire real estate needed for large power plants with combined cycles.

The paramount objects for such combined cycle power plants will be existing power plants that can be reconfigured as combined cycle production facilities. Such facilities have an advantage because they have the proper zone with the necessary electrical and mechanical infrastructure. The disadvantage is that such an object may lack space in order to carry out a project to put the power plant into operation with combined cycles. Therefore, from the point of view of the availability of space, as well as from the point of view of cost savings, it is desirable to limit the size of the power plant, as far as possible.

Noise Level / Public Opinion

Public approval is becoming increasingly difficult to obtain for many public power plant projects. Factors such as noise, increased traffic, unsightly terrain, pollution, hazardous waste, and more, increase dissatisfaction with power plants located near populated areas. Preferred are smaller, quieter, smaller equipment, lower emissions and inconspicuous compared to larger and more flashy power plants. Therefore, power plants with a combination of cycles, which are more compact, have a greater "density of electricity" (energy per unit volume), are desirable.

However, traditional power plants that are being built today to meet modern energy needs simply repeat the proven designs of power plants. No attention is paid to the question of whether it is possible to reconstruct power plants taking into account public opinion regarding the power plants themselves.

Heat dissipation

Conventional steam and combined cycle power plants require some kind of heat removal. This is usually necessary to condense the low pressure steam from the steam turbine exhaust. Such heat removal can take place in air, a river, a lake or a “reservoir” capable of absorbing heat. Since this heat release influences the local ecology and possibly the biological environment (for example, the presence of fish in the river), there is always a demand for methods to reduce heat release.

Decline in gas turbine performance

Although combined cycle power plants are highly efficient, this is the case with “new” power plants. Since in the current level of technology the combined cycle is usually based on the use of gas turbines, the level of its performance is very prone to decline, as a result of which the gas turbine's performance decreases significantly (from 2% to 6%) within one or two years of operation. This can become significant from the point of view of fuel cost, since the combined capacity of a power plant with combined cycles also decreases along with a decline in the characteristics of a gas turbine.

SUMMARY OF THE INVENTION

Accordingly, the objectives of the present invention are aimed at solving the problems of the existing level of technology and are reduced to the following:

1. To ensure the development of a power plant with a combination of cycles, which is more flexible from the point of view of its design than modern proposals, and has the ability to use modern technologies, acquired as necessary.

2. Reduce the total fuel consumption at rated power, but especially under partial load conditions, since most likely the power plant will operate only a small part of the time under full rated load conditions.

3. Reduce the cost of installing a power plant so that debt servicing is significantly reduced and financing by a bank or other lending institution would be easier for the owner.

4. Balance the ratio of time and cost in terms of investment, maintenance and fuel costs, so that the creation of power plants is more economical, and if possible reduce the total cost of electricity production.

5. Provide the ability to meet the demand for electricity during the peak season when operating at full load without reducing the usual level of performance or significantly increasing the cost of installation.

6. To reduce losses and losses associated with the transmission of electricity over long distances.

7. To increase the overall reliability of the electricity supply network by providing the possibility of generating electricity near the place of consumption during the period of greatest electricity demand.

8. Reduce the cost of operation and technical support. In addition to the cost of acquiring fuel, also reduce maintenance costs, supplies, supplies, insurance and other operating costs.

9. Reduce the need for gas fuel compression.

10. Increase the level of reliability.

11. Reduce air consumption and air cleaning requirements.

12. Reduce air emissions of pollutants, especially NOX.

13. To reduce discharge into water of discharge from PGRT and other sources.

14. Use control mechanisms to maximize productivity, reliability and heat recovery.

15. Simplify the way of operation and develop methods and / or strategies to increase the performance at part load and reduce the level of harmful emissions.

16. Optimize the performance of a steam turbine through the use of projects with improved volumetric performance and additional capacity to meet the demand for electricity in peak season.

17. To ensure an economical attitude to the land areas necessary for the location of the power plant by designing a compact power plant with a high energy density.

18. To reduce the noise level, size, area and volume of equipment to reduce the impact of the power plant on the local population.

19. Achieve a minimum level of heat dissipation.

20. To provide more cost-effective economic and territorial reconstruction of existing steam power plants and equipment with combined cycles to reduce capital costs and economic consequences associated with the installation of a large number of new equipment and real estate.

21. Create economic incentives for the use of environmentally friendly projects in the construction of new power plants.

22. Ensure the design of power plants with a combination of cycles that are less susceptible to the deterioration of the characteristics of gas turbines.

These problems can be solved within the framework of the invention, which is discussed in the following sections.

SUMMARY OF THE INVENTION

In short, the present invention is a system and method for using fewer and / or smaller gas turbines (GTs) and heat recovery steam generators (HRSGs) in combined cycle power plants. Traditional equipment of this type, when using a combination of cycles, is replaced by large-sized steam turbines (ST) and heat recovery steam generators with constant fuel combustion, which provides a number of economic, energy-intensive and environmental benefits.

The presented technology uses HRSG with several pressure levels to maximize heat recovery from the exhaust gases of a gas turbine. This option is usually used, since the current level of technology teaches not to use HRSG with constant combustion maintenance due to the common widespread view that such configurations have lower thermal performance. Despite the general conviction, the present invention shows that PGRT with constant combustion maintenance can be performed with a thermal capacity equal to or higher than known from modern practice of combining cycles. However, in order to achieve such a performance level of HRSG with constant combustion support, steam turbines must be designed and implemented differently.

In some preferred embodiments of the use of the present invention, gas turbines remain unchanged with respect to the prior art, but the exhaust is carried out in the HRSG. Such PGRT, however, is designed as a steam generator with a single pressure level (or initially having a single pressure level), which is optimized to ensure constant combustion and the formation of steam with a higher pressure, in contrast to the cases of the traditional use of combination cycles. In addition, the named PGRT is designed to control feed water flows through the economizer / feed water compartments to maximize heat recovery. In addition, steam turbines are designed to be large in size, which is typical of a traditional power plant in the Rankine cycle, where overheating and the resulting steam from a traditional steam turbine are fed to feed water heaters to maximize thermal performance. This is the advantage of a larger steam turbine, which is commonly used in a traditional steam power plant, which is described by Moore in Patent 5649416, owned by General Electric:

“Traditional steam power plants have advantages in terms of both lower cost and higher productivity due to the economical range of parameters. The traditional rule of thumb regarding cost states that doubling the parameters of a power plant results in a ten percent reduction in cost. According to this rule, the cost of one large production unit is supposed to be ten percent lower than the cost of a power plant consisting of two half-size units. ”

“The level of performance also improves with increasing size and energy parameters. As in the case of turbomachine engineering in general, the in-house productivity of a steam turbine is a strong function of the volumetric flow rates at the inlet, which are directly proportional to these parameters. It is also well known that the thermal performance of the Rankine cycle increases with the pressure at which steam is generated. An increase in pressure, however, reduces the volumetric flow rate of the steam stream at the turbine inlet, thereby reducing the internal expansion power. This as a result leads to the fact that the total productivity at low volumetric flow rates is significantly higher than the total productivity at high volumetric flows. Therefore, an additional advantage of increasing turbine size related to performance is that higher steam pressure in front of the throttle can be used more efficiently. ”

Using PGRT with sufficient fuel for additional combustion, the lower cycle of the present invention may be more independent of the operation of a gas turbine. Therefore, gas turbines can operate at full load, while the combined load of a power plant changes in accordance with its wide capabilities in full load operation only due to a change in the parameters of additional combustion and the load of a steam turbine. This increases the overall capacity of the power plant using a specific set of gas turbines, provides flexibility of the parameters of the power plant with a combination of cycles by changing the characteristics of additional combustion, and also increases the overall thermal performance of the power plant when working with partial load. In addition, it simplifies the work and can also potentially reduce emissions.

Design allows you to create PGRT, providing combustion at temperatures up to 2400 ° F, an exemplary simple design “2 in 1”, which presents two gas turbines and one large steam turbine, replaces two structures “2 in 1” (design “4 in 1” usually not used if overheating is used, due to balancing problems in the overheating profile). This exemplary design saves two gas turbines, two PGRT, one steam turbine, three switches, three transformers, as well as auxiliary equipment, real estate and technical support necessary to maintain this equipment. The capital costs for a power plant in US dollars / kW can thus be significantly reduced using the present invention.

All this can be achieved using technologies and hardware already tested in the field of turbomachine engineering. The use of PGRT with constant combustion support and a uniform pressure level is a new concept for this device, but it also does not go beyond the scope of modern technological practice or the possibility of implementation with the existing equipment. Therefore, from the point of view of reliability, there are no or almost no mutual contradictions. The general construction for several preferred embodiments of the invention is shown in FIG. 13, and some exemplary embodiments with a large number of details are shown in FIG. 9 and FIG. 15.

Improvements in comparison with the prior art

The present invention solves the problems of the prior art in this field by achieving the following objectives:

1. Providing greater flexibility in the design of a power plant with a combination of cycles in such a way that developers can achieve a high level of productivity, while choosing a specific power.

2. Reducing overall fuel consumption while improving performance in full and partial load operation.

3. Reducing the cost of placement and installation of the equipment fleet by increasing the energy density of the equipment (a greater amount of energy generated per unit of equipment).

4. Reducing the total cost of electricity production by reducing the three main factors associated with its production: energy consumption, capital costs and maintenance costs.

5. Providing temporary capacities to achieve peak loads when using additional combustion to generate more steam, as well as applying the steam turbine operating mode at overpressure (inlet pressure is slightly higher than the nominal pressure) and reducing the flow of generated steam to the feed water heaters.

6. Increasing the efficiency of the energy supply network by providing the possibility of local electricity production at peak season. Ensuring the production of electricity at the peak season in the field eliminates or reduces the shortcomings associated with the “import” of electricity from other parts of the country or from abroad. (Refers to transmission losses on power lines).

7. Increasing the reliability of the power supply network by reducing the transmission of electricity over long distances at peak season at maximum load.

8. Reduction of operating and maintenance costs mainly by reducing the number of equipment, systems, as well as the use of equipment requiring less maintenance for a kilowatt of generated electricity (steam turbines with low maintenance needs instead of gas turbines, whose similar needs are high).

9. Minimizing the need for compression of fuel gas through the use of fewer gas turbines and gas turbines with lower fuel gas pressure in the cycle, together with a larger steam turbine.

10. Improving reliability by simplifying the design of a power plant.

11. Reduced air consumption by using fewer gas turbines.

12. Reducing emissions of pollutants, especially NOX, through the use of gas turbines in a stable, almost waste-free mode and the use of cleaner gas turbines and in smaller quantities.

13. Reducing to a minimum discharge of condensate and other emissions into the water through the use of more productive cycles that require less steam per kilowatt of electricity generated.

14. The use of control mechanisms to increase productivity, reliability and heat recovery.

15. Simplification of work through the use of gas turbines at full load over a wide range of operation (the total generation of the power plant when combining cycles) and reducing the pressure levels of PGRT to the only one.

16. The maximum increase in the efficiency of the steam turbine by increasing the volumetric flow rate, especially in the compartments of high and intermediate pressure.

17. The economical consumption of space and land using smaller equipment and projects with a higher density of electricity.

18. Reduce noise, size and space requirements by using fewer pieces of equipment.

19. Keeping heat removal to a minimum due to the use of high efficiency cycles with less heat removal per kilowatt of generated electricity.

20. Creation of a project based on the principle of combination of cycles, more compatible with existing steam power plants, allowing more compact and efficient transfer of existing power plants to the use of combination technology of higher productivity cycles.

21. Minimal reduction of air consumption, pollutant emissions, as well as heat emissions into the atmosphere, but providing these benefits at a lower cost than traditional combined cycles.

22. Reducing the impact of a decline in gas turbine performance by using a combined cycle power plant that is less dependent on gas turbines and their efficiency.

BRIEF DESCRIPTION OF THE DRAWINGS

For a more complete understanding of the advantages of the invention, the following drawings and their description are provided:

Figure 1 shows the basic thermodynamic cycle of Rankine.

Figure 2 shows a system for the production of electricity in accordance with the current level of technology, using the basic Renkine cycle.

Figure 3 shows a diagram of the Rankine cycle, including a cycle of overheating and steam generation for heating the feed water, according to the use of the prior art in the case of a conventional power plant.

Figure 4 shows a comparative table of the performance of the base Rankine cycle, namely, the Rankine cycle with various improvement options to increase productivity.

Figure 5 shows a diagram of the basic principles of combining cycles.

Figure 6 shows a prior art diagram for a combined cycle power plant using gas turbines, PGRT and steam turbines.

Fig. 7 shows a heat consumption curve with respect to an energy generation generator for a General Electric industrial model PG7241 standard gas turbine.

Fig. 8 shows the dependences for operating at partial load of a power plant with combined cycles of the General Electric company with two GE C207 gas turbines using graphs corresponding to the operating characteristics for one or two gas turbines.

Figure 9 shows a General diagram of the installation in accordance with one embodiment of the present invention with respect to its use in a power plant.

Figure 10 shows a table of comparative performance data in accordance with the present invention and the prior art.

Figure 11 shows a graph of a typical temperature dependence of vapor enthalpy at a pressure of 1800 pounds per square inch, based on the assumption that water is a moving fluid.

On Fig shows a graph of a typical dependence of the enthalpy of the exhaust gases of a gas turbine on the temperature of the exhaust gases.

13 is a block diagram illustrating the general principles of the present invention in their use in combined cycle plants.

On Fig shows a graph of a typical dependence of a given diagram of the average temperature difference from the intensity of the fluid flow for options with superheating and overheating.

On Fig shows a diagram illustrating a model implementation of a power plant with combined cycles using the advantages of the present invention.

On Fig shows a graph of an exemplary control system that can be used to control one or more PGRT relative to the lessons of the present invention.

On Fig shows a block diagram of a General model of the organization of the control system, which can be used to control the entire system of the power plant in accordance with the present invention.

On Fig shows a block diagram of a model organization of a control system that can be used to control and directly influence the increase in power generation in accordance with the present invention.

On Fig shows a block diagram of a General model of the organization of the control system, which can be used to control and directly affect the reduction of power generation in accordance with the present invention.

On Fig shows a block diagram of a General model of the organization of the control system, which can be used to control and directly regulate the transient power in accordance with the present invention.

On Fig graphically depicted sources of energy, losses and performance, which are taken into account in the total analysis of the energy flow.

On Fig shows a typical configuration of a power plant with combined cycles of the model GE 207FA.

23 shows operational data tables for a typical power plant with combined cycles of the GE 207FA 521 MW model, taking into account a typical multifunctional test data profile.

On Fig shows a typical configuration of a power plant with combined cycles, model Westinghouse 2X1 501 G 715 MW.

On Fig shows the table of operational data for a typical power plant with a combination of cycles of the model Westinghouse 2X1 501G 715 MW, taking into account a typical multifunctional profile of test data.

FIG. 26 shows a typical 725 MW combined cycle power plant described in a preferred embodiment of the present invention.

On Fig depicts the table of operational data for a variant of the system with a capacity of 725 MW in accordance with the present invention, using a shielded PGRT, based on a typical multifunctional profile of test data.

On Fig depicts a graph of the relative differences in operation under partial load when comparing a conventional system with a combination of cycles of a power plant and the system of the present invention.

On Fig graphically shows the data on several models of the power plant and their nominal energy characteristics, in accordance with the present invention.

30 is a graph of a baseline steam cycle productivity level required for a power plant model using two General Electric rated industrial gas turbines, model PG7241 (FA), to match prior art performance levels with respect to energy parameters.

On Figa shows a diagram of a typical hourly load curve of the regional system (according to the publication "Electricity Prices in Competitive Conditions: Extremely High Cost of Electricity Generation Services and the Financial Status of Electricity Utilities" ("Electricity Prices in a Competitive Environment: Marginal Cost Pricing of Generation Services and Financial Status of Electric Utilities "(DOE report number DOE / EIA-0614)));

FIG. 31B is a diagram of a typical load duration curve showing the total long-term use of a power plant with rated capacity (information from Duke Energy Power Services, Inc.).

On Fig shows a typical steady-state data on a weekly load chart based on the information Figa.

On Fig graphically shows the performance under partial load of several models of power plants using the present invention, as well as several examples using the existing level of technology.

Fig. 34 is a tabular view showing comparative economic data for a power plant model using the present invention for a power plant with combined cycles of the GE C207FA model and Westinghouse 2X1 501G model, both examples are from the prior art.

On Fig depicts the sequence of operations of a typical process of thermal balancing for a model of a power plant supercritical regime in accordance with the present invention according to Fig.26, 27, 28, 33 and 34.

On Fig, 37 and 38 are tables of some production data related to Fig. 35.

FIG. 39 is a flowchart of an exemplary heat balancing process for an ultra-supercritical mode power plant model in accordance with the present invention of FIG. 33.

On Fig, 41 and 42 are tables of some production data related to Fig. 39.

FIG. 43 is a graphical depiction of a power plant load control method that can be used in combining cycles of the present invention using two or more gas turbines.

On Fig shows a table of comparative data for the conversion of an existing steam power plant to the technology of combining cycles using the preferred embodiment of the present invention and the prior art.

On Fig shows a diagram of a preferred embodiment of a power plant with combined cycles using mixed fuels and a boiler that runs on combustible fuel.

Fig. 46 is a diagram of a preferred embodiment of a combined cycle power plant using mixed fuels, a nuclear reactor, a geothermal steam generator, or other vapor generating energy source.

On Fig shows a model of the design / financing of the process to determine the preferred and / or optimal variant of the invention for a particular power plant.

On Fig shows a model of the design / financing of the process to determine the preferred and / or optimal variant of the invention for a particular power plant.

On Fig shows a model of the process of readjustment of the power plant to determine the preferred and / or optimal variant of the invention for a particular power plant.

On Fig shows a model of calculations for fuel to determine the preferred and / or optimal variant of the invention for a particular power plant using mixed fuels.

On Fig shows a General Electric steam turbine with three blocks and four steam streams with a combined high / intermediate pressure compartment and two dual-flow low pressure compartments.

MODES FOR CARRYING OUT THE INVENTION

Disclosure of Examples of Use of the Invention

Although the present invention can be used in various forms, the drawings are provided in this application and various detailed embodiments of the invention will be further described, given that the disclosed information is to be regarded as exemplary principles of the invention and that the information provided does not in any way limit the breadth of use of the invention those options that are shown in this document.

Charts and graphs

It should be particularly noted in the context of the descriptions provided herein that circuits, graphs, diagrams, and the like may include components and / or steps that in no way limit the overall direction of the achievements of the present invention. Also, the components and / or steps can be removed and / or shifted elsewhere in the following descriptions without any loss in the general focus of the invention. This aspect is especially important with respect to process flow models in which the achievements of the present invention can be used by a computer technician to create control systems that are functionally similar, despite some replacements or changes in the disclosure of steps or processes and providing the same results. as using the advantages of the present invention.

Equipment

In the discussion of the invention throughout this document, equipment manufactured by General Electric, Westinghouse, and other manufacturers will be mentioned. Particularly often mentioned is the General Electric model C207FA of the power plant, consisting of gas turbines of the General Electric model PG7241FA, as well as similar equipment from Westinghouse or other manufacturers. Such references are merely models and are presented to a reader having certain knowledge in a given industry sector for understanding the conclusions of the present invention.

In this document, an attempt was made from a more practical side to describe the advantages of the present invention in terms of equipment familiar to specialists in this field and used in the energy industry. None of the provisions of this document can be construed as limiting the scope of the present invention before using any particular model of equipment or its manufacturer. On the contrary, with regard to equipment to which the present invention may be applicable, this document should be construed as broadly as possible.

OVERVIEW

Steam has been used to produce energy for decades since the advent of steam locomotives that burned solid fuels such as wood or coal for energy. Before and during the 1980s, steam power plants still produced the bulk of the electricity in the United States using coal, oil, or nuclear fuel.

However, by the 1980s, smaller power plants with parallel energy production had been developed and built. Such power plants used gas turbines as their main engine with PGRT connected to the exhaust gas turbine to recover heat waste (the temperature of the exhaust gas usually reached 900 ° F to 1200 ° F) and convert the exhaust into steam. Such steam was then used for various purposes, district heating, steam treatment or the generation of additional energy in a steam turbine. This configuration of a power plant, gas turbine, PGRT and steam turbine has become known as a combined cycle design, and due to its high productivity, low cost and light construction, it has become the preferred power plant option for new independent energy producers.

However, during the evolution of such a combined cycle power plant, it became a power plant that uses a gas turbine as the main engine, and the steam turbine is a secondary engine. A study of building blocks with combined cycles offered by gas turbine manufacturers today will confirm this conclusion, since in most power plants with combined cycles in the prior art, gas turbines generate about two-thirds of all generated energy, and steam turbines produce about one-third. A review of the proposals of manufacturers of building blocks with combined cycles illustrates this trend. The 1997 Turbomachine Brochure (TURBOMACHINERY HANDBOOK, (USPS 871-500, ISSN 0149-4147)) lists standard power plants with combined cycles offered by various manufacturers, including ABB, General Electric and Westinghouse. In almost every case, steam turbines generate between 40 and 60% of the amount of energy that is generated by a gas turbine. 1996 General Electric Information Brochure, General Electric informative document GER-3567G, 1996, GE Heavy-Duty Gas Turbine Performance Characteristics, by Frank J. Brooks) talks about the performance of gas turbines corresponding to combined cycle power plants.

Several preferred embodiments of the present invention pay tribute to the combined cycle unit for its high level of performance, low cost and ease of configuration. However, the present invention provides a slightly different view of the relative size of individual engines. Although modern gas turbines have a capacity of 30 to 40%, they require the use of PGRT and a steam turbine in order to achieve a combined cycle performance of 50 to 60%. In addition, for efficient heat recovery of exhaust gases, such HRSGs usually have three levels of vapor pressure: high, intermediate, and low. The use of low and medium pressure steam leads to a total steam cycle productivity of only 34 to 36%.

The steam cycle productivity of modern large power plants, however, ranges from 45 to 50%. To achieve this level, the use of low pressure steam, as in the case of conventional combination cycles, is unacceptable. Therefore, several embodiments of the present invention describe a method that uses only high-pressure steam to achieve high steam cycle productivity in a combined cycle configuration, and also regenerates as much heat from the exhaust gas of a gas turbine as with the prior art combined cycle technology.

In this way, a power plant with a combination of new technology cycles departs from a typical prior art version, where the gas turbine was the main (largest) engine and the steam / gas turbine ratio was approximately 1: 2. A preference is proposed for the option in which the steam turbine is usually the main (larger) engine, and the ratio of steam / gas turbines can usually be selected in the range from 0.75: 1 to 2.25: 1 or more. This coefficient is easily changed due to changes in the design of the steam turbine, additional combustion parameters and the steam cycle.

During the operation of any power plant, workers must modulate power generation by the power plant by the load on the system (energy consumption by all users of the power supply system). Along with changes in the system load, the amount of energy produced by all the power plants connected to this network must also change to meet such changes, otherwise the speed of the equipment will change, and the received electricity will not have a frequency of 60 Hz (60 cycles per second for US power plants, etc. .). This will seriously affect the equipment used by the end users of electricity (electric clocks will not show the exact time, electric motors will not work at the required speed, etc.). Therefore, utility workers and power plants are responsible for maintaining a constant frequency or speed on their equipment. To achieve this goal, they must constantly change the amount of energy generated to match this system. It should be noted that in European and other countries, the nominal frequency is 50 Hz, and in the USA and other countries of the Western hemisphere, for example, in Canada, -60 Hz.

On hot summer and very cold winter days, the system load almost reaches its peak. Also on weekdays, usually from 16:00 to 20:00, the system reaches its daily peak. However, at night and on weekends, the system load can average 60% of the weekly peak. In connection with such changes in the dynamics of the load of the system for power plants with electricity distribution (controlled by utilities to match the system loads), it is unusual to work for a long time with a nominal output or at a constant load. On the contrary, it is typical to work with a high level of work during peak hours during the week (which does not necessarily correspond to the nominal load) and with a relatively low load (approximately 60% of output) during night periods and weekends. More complete information is presented in FIGS. 31A, 31B, and 32 for more complete information about a typical load level.

Thus, in order to be effective, the power plant must be flexible for constant fluctuations in load from 50 to 100%. Traditional power plants with combined cycles are efficient, but they have to refuse to work at full capacity under partial load. This is especially true for power plants where gas turbines are the main engines. In such power plants, to reduce the load from a higher level, more complex gas turbines with inlet guide vanes reduce the air flow passing through the engine, thereby reducing the pressure ratio. In addition, in order to further reduce the load on such engines, the temperature at the turbine inlet (also called the combustion temperature) must be reduced to work with a partial load. A decrease in the mentioned pressure and temperature significantly reduces the operational power of the gas turbine engine.

To improve the performance of a power plant with a combination of cycles, reduce costs, reduce emissions and property requirements, as well as simplify the operation and maintenance of a power plant, the present invention proposes the use of HRSG optimized for continuous additional combustion, which uses an evaporator (boiler) with a single pressure level equal to or greater than the pressure at the inlet of the steam turbine, usually used in the combined cycle in the framework of the previous prior art. In addition, the present invention proposes to use some of the capabilities of traditional Rankine cycles that are not used in conventional combined cycles.

Turning to FIG. 9, illustrating a new cycle implementation model. As in the case of a typical unit with combining cycles of the prior art, the new version uses one or more gas turbines as an energy device of the upper cycle. Also, as in the case of a typical unit with a combination of cycles, known from the prior art, the exhaust gases of a gas turbine enter the PGRT. From now on, however, the cycle changes compared to the traditional use of combining cycles. HRSG with a single pressure level is used instead of HRSG with several pressure levels. To increase cycle efficiency, the pressure of the steam produced can be significantly higher than the typical pressure rating of 1800 psia. Such pressure may, if desired, be supercritical (above 3206 psia). For simplicity, subcritical options (2400 psia) will be considered with specific examples. However, performance curves under supercritical steam conditions will also be considered.

Energy flow analysis

First, it will be correct to consider the total energy flow in the application of the traditional combination of cycles. Simple energy analysis FIG. 21 shows the energy flow in a combined cycle plant, while FIG. 10 determines the energy flow for a conventional combined cycle power plant based on the parameters of the prior art (see the following section on optimizing the preferred cycle for equations, used to calculate the parameters of Fig.10). Table 10 shows the performance data of a General Electric PG7241 (FA) gas turbine under ISO standard conditions at 3.0 inches of water inlet differential pressure inlet and 10 inches water column differential pressure in the exhaust. ISO conditions are defined as 59 ° F and 14.696 psia for ambient pressure. According to Fig.21, 32.31% (all percentages are given on the basis of the coefficient of higher heat transfer HHV) of the initial fuel at the entrance to the gas turbine (2101) is converted into electricity, which is the generation of a gas turbine (2105). Depending on the exhaust gas flow of the gas turbine and its enthalpy, only 56.21% of the supplied energy is supplied to the HRSG (2103), which means that 11.46% are lost between the gas turbine and the HRSG (2102). This is probably the loss of a gas turbine generator, the heat loss of a gas turbine, auxiliary devices powered by equipment, auxiliary devices powered by an engine, losses from air resistance and other losses. In this example, auxiliary combustion energy is not added; therefore, such energy (2104) is taken as the zero value. Of the remaining 56.21% of the input energy of the gas turbine sent to PGRT (2103), about 10.7% (which equals (0.107) (0.5621) or 6.04% of the initial energy of the gas turbine) are lost in the exhaust pipe (2107 )

Of the remaining energy of the exhaust gases of a gas turbine for steam production in PGRT, 1% is considered lost as heat released into the environment in 2106. In terms of energy at the inlet of a gas turbine, this means a loss of 6.04% of the input energy of the gas turbine for exhaust loss and 0, 50% of the input energy of a gas turbine for heat loss in HRSG. Thus, 49.67% of the input energy of the gas turbine remains as energy converted to steam (2108), which can be converted into electricity by the steam turbine after regeneration.

Based on published thermal power of 6040 BTU / kWh for the General Electric GE STAG C207 FA power plant with two GE Frame 7s and one General Electric steam turbine, the power plant’s performance, taking into account the highest heat transfer coefficient of natural gas (HHV) , reaches 50.90%. If a gas turbine converts 32.31% of the supplied fuel (2101) into electricity, minus this, the steam turbine must convert 18.59% of the supplied fuel (2101) into electricity. With a steam turbine generator capacity of 99% (or 1% loss, (2110) and with an auxiliary load factor of 97.5%, as well as 49.67% of the received fuel (2108) used in the steam turbine cycle, the basic steam cycle capacity is 38 , 78% ((18.59 / 49.67) / (0.975) (0.99)). This is significantly less than 46.78% of the effective operation of special steam cycles of a power plant using only the Rankine cycle (see Figure 4 )

This steam cycle performance is confirmed by General Electric in its 1996 brochure, Production Line and Combined Cycle Operations of General Electric, by David L. Chase, Leroy O. Tomlinson, Thomas L. Davidson, Raub W. Smith, and Chris E Maslaka ("GE Combined-Cycle Product Line and Performance" by David L. Chase, Leroy O. Tomlinson, Thomas L. Davidson, Raub W. Smith, and Chris E. Maslak). In the discussion of additional burning of HRSG, this document states:

“... the increment in energy production efficiency through the use of additional combustion is 34-36% based on the lowest heat transfer coefficient (LHV).”

Since the additional combustion of fuel adds heat only to the steam cycle, the performance of the steam cycle of power plants with combined cycles of General Electric is the same as described above.

Cycle performance

The next question that needs to be answered is how to achieve the level of productivity of a traditional cycle of a steam power plant using the steam part of the combined cycle. Referring to FIG. 4, it should be noted that overheating helps to improve steam cycle productivity. However, overheating is already used in many high-performance combined cycles, such as General Electric's C207FA STAG, which uses two GE Frame 7s and one General Electric steam turbine to achieve a heating level of 6040 BTU / kWh (see 1996 brochure " David L. Chase, Leroy O. Tomlinson, Thomas L. Davidson, Raub W. Smith and Chris E. Maslak's General Electric Production Line and Combined Cycle Operations (GE informative document GER-3574F (1996), entitled "GE Combined-Cycle Product Line and Performance "by David L. Chase, Leroy O. Tomlinson, Thomas L. Davidson, Raub W. Smith, and Chris E. Maslak) Therefore, other improvements in steam cycle performance are associated with the use of higher steam turbine inlet pressures with higher flow rates (for higher steam turbine performance) and heating feed water. in several embodiments of the present invention.

Most modern high-performance gas turbines, such as the General Electric model PG7241 (FA), have an internal combustion temperature of around 2400 ° F. Such gas turbines are designed to use exhaust gas temperatures when operating at a standard load of 1100 ° F. For PG7241 (FA) under ISO conditions, adjusted for the loss of 10.0 inches of H ₂ O on the PGRT exhaust, the exhaust temperature is 1123 ° F. The corresponding exhaust gas flow is 3,552,000 lb / hr (pound - force per hour). With an inlet pressure of 1800 psia, an inlet temperature of 1050 ° F, overheating to 1050 ° F, exhaust at 1.2 inches of mercury, this steam cycle would require 1642.4 BTU / lb of input heat (see Figure 4, option 2) . With an input enthalpy of exhaust gases of the order of 412.6 BTU / lb and an enthalpy of gas at the outlet of 159.2 BTU / lb, the exhaust gases carry enough energy to produce about 548,000 BTU / lb of vapor flow. However, a more detailed study reveals a flaw in this logic. At 1800 pounds per square meter. an inch of steam boils at 621 ° F. As heat flows from higher to lower temperatures, a value of 650 ° F would be a suitable temperature level for exhaust gases leaving the evaporator section. If water was supplied to the evaporator, preheated to an enthalpy of 648 BTU / lb (optimal assumption), the energy required to generate steam is 1642.4-648 = 994.4 BTU / lb.

However, at 650 ° F, the exhaust gases have an enthalpy of 281.3 BTU / lb. Therefore, exhaust gases can boil and overheat only 469,000 Ib / hr (3,552,000) (412.6-281.3) / 994.4. Thus, the issue turns into a heat transfer problem, since there is an insufficiently high level of energy (high temperature) to generate steam at higher pressures.

Conversely, from an exhaust gas temperature of 650 ° F to exhaust at 180 ° F, there is enough energy to preheat water in an amount of 729,000 Ib / hr ((3,552,000) (281,3-159,2) / (648-53)) of hot tank at 53 BTU / lb for a saturation enthalpy of 648 BTU / lb. Therefore, the use of a boiler with a uniform pressure level for heat recovery, as in the prior art, is inefficient and either generates an excess of hot water that is not used or has little use in the production of electricity, or the temperature of the exhaust gas heaters significantly exceeds the optimal one. This result entailed the introduction of HRSG with several levels of pressure in the framework of the prior art. This option uses the energy of the heated water or exhaust gases by supplying section (s) of the low pressure evaporator to the HRSG, which (e) converts unusable hot water / exhaust gases to low pressure steam. Although this steam contains less energy and is less able to produce energy in a steam turbine than high pressure steam, low pressure steam nevertheless contributes to the energy production of a steam turbine and is used to reduce heat consumption by a power plant to generate energy ( thermal power).

Additional burning

Another way to eliminate the disadvantages of heat transfer of PGRT with a single pressure level without the need to increase the number of pressure levels, as is known from the prior art, is to add additional energy at the inlet or near the inlet of the PGRT by providing additional combustion of fuel (additional combustion). However, the findings are such that additional combustion reduces the overall thermal performance of the power plant. This is noted by Moore in Patent 5649416, which states that

“Additional combustion of PGRT fuel can increase the total amount of energy generated and part of the total energy produced by a steam turbine, but only if the total thermal capacity of the power plant decreases.”

In addition, Rice (Rice) in patent 5628183 declares that

“Supplementary combustion before HRSG is not the right solution to the problem of increasing cycle throughput.”

Thus, according to the prior art, the use of additional combustion technology is not effective. In addition, Rice mentions other documents from General Electric and Westinghouse that support the same point. In the General Electric information brochure entitled "Production Line and Combined Cycle Operations of General Electric" by David L. Chase et al. (GE informative literature, GER-3574F (1996), entitled "GE Combined-Cycle Product Line and Performance "by David L. Chase, Leroy O. Tomlinson, Thomas L. Davidson, Raub W. Smith, and Chris E. Maslak) states that

“... productivity in energy production through the use of additional combustion is 34-36% based on the lowest heat transfer coefficient (LHV).”

It follows that, although the combined cycle capacity is 56% in accordance with the lowest heat transfer coefficient (LHV) at full load, the energy generated through additional combustion is added at a capacity equal to or less than 36% of the lower heat transfer coefficient (LHV).

In the same document there is information indicating additional combustion as a factor that reduces productivity (thermal power) - this is table 14, according to which additional burning of HRSG can increase the power of power plants known from the prior art with combined cycles by 28%, but only with an increase in the combined heat output of the combined cycle by 9% (for the consumption of a specific fuel).

The flow of energy in the framework of the present invention

It would be appropriate to re-examine the total energy flow in the combined cycle, as is done in several embodiments of the present invention. The simple energy analysis in FIG. 21 shows the energy flow in a combined cycle plant, while FIG. 10 determines the energy flow for a conventional combined cycle power plant based on the parameters of the prior art (see the next section on optimizing the preferred cycle for equations used to calculate the parameters of FIG. 10). As already mentioned, gas turbine operating data is provided for the General Electric model PG7241 (FA) under conditions meeting the ISO standard at 3.0 inches H ₂ O inlet air pressure difference and 10 inches H ₂ O exhaust pressure difference. As shown in FIG. 21, 32.31% (all percentages are based on the HHV coefficient of higher heat transfer) of the initial fuel at the inlet of the gas turbine (2101) is converted to electricity, which is the output of the gas turbine (2105). Depending on the exhaust gas flow of the gas turbine and its enthalpy, only 56.21% of the energy supplied is supplied to the HRSG (2103), which means that 11.548% is lost between the gas turbine and the HRSG (2102). Of these remaining 56.21% of the input energy of the gas turbine directed to PGRT, about 10.7% is lost when passing through the exhaust pipe (2107), leaving 50.17% of the input energy of the gas turbine for PGRT. In this part, there are no changes compared to the prior art.

To ensure maximum heat recovery of the HRSG, some embodiments of the present invention contemplate increasing the flow of feed water through the HRSG until a sufficient balance of heat received from the feed water is achieved to match the heat loss of the exhaust gases necessary for optimal heat recovery (i.e., reducing PGRT exhaust gas temperatures up to approximately 180 ° F). Secondly, by adding fuel to the intake of the PGRT (additional combustion), the energy of the exhaust gases of the PGRT increases until its level becomes sufficient to convert most or all of the supply water to steam. When using option 3 in FIG. 4, heat must be added at 1633 BTU / lb to create the necessary steam characteristics. Since the thermal capacity of the exhaust gas is approximately 0.25 BTU / lb / ° F, and the thermal capacity of the recirculating condensate is approximately 1.0, the steam flow should be approximately 0.25 lb of steam per lb of exhaust gas flow. For two GE frame 7 gas turbines, this means a steam return of 1,776,000 Ib / hr.

2894 MMBTU / hr (one million BTU / hr) is needed to produce this amount of steam. Considering 1% of environmental losses from PGRT (2106), the heat demand reaches 2923 MMBTU / hr. When exhaust emissions are lost, the PGRT energy needed to produce this steam is 87.99% of the input energy of the gas turbine. Since the input energy of the PGRT 2103 minus the loss of exhaust gases of the PGRT (2107) is 50.17% (56.21-6.04) of the input energy of the gas turbine (2101), it is necessary to add energy equal to 31.78% of the input energy of the gas turbine, as heat from additional fuel combustion (2104), which will total 81.95% of the input energy of a gas turbine (2101). Subject to an adjustment of 1% to environmental losses (2106), 31.3% of the input energy of the gas turbine (2101) is converted to steam. Such steam at this stage is ready for its conversion into electricity in a steam turbine.

When using a steam turbine in accordance with the invention, higher pressures, overheating and heating of the feed water can also be applied. In addition, the characteristics of a steam turbine will be evaluated as exceeding 2.5 times the characteristics of a standard power plant with combined cycles. This will help the designer to use the steam cycle productivity of 44.38%, as shown in the example of option 5 of Figure 4 (if we follow the Moore, the steam turbine productivity increases as the parameters increase, but for the purpose of demonstration, the total productivity of 90% was set aside in this example).

When using a basic steam cycle with a capacity of 44.39%, 36.01% of the heat is converted into mechanical energy (shaft power in hp), taking into account the generator efficiency of 99% and a factor of 97.5% for calculating the side load. The electricity generated by the steam turbine is 34.76% of the input energy of the gas turbine (2101). When generating a gas turbine (2105) equal to 32.31% of the input energy of a gas turbine (2101), the generation of a steam turbine is 34.76% of the input energy of a gas turbine (2101), and with the additional use of additional fuel at the input in the amount of 31.18% of the input energy of the gas turbine (2101), the combined cycle capacity will be (output divided by the input energy) ((0.3231 + 0.3476) / (1 + 0.3118)), i.e. 50.90%. When using only two feed water heaters in a cycle, the performance of this model is equal to the performance of a traditional power plant with combined cycles of General Electric. For supercritical modes, the combined capacity of the combined cycle in some cases increases to 51.75% and provides a decrease in the lowest heat transfer coefficient to 5942 BTU / kWh (BTU / kWh) (see Figures 10 and 21).

Thus, from the point of view of the total amount of energy, it is obvious that additional combustion is not detrimental to the overall performance of the combined cycle, IF adding additional energy to the lower cycle is accompanied by a corresponding increase in the productivity of the lower cycle.

Optimization of the preferred cycle option

As mentioned earlier, one of the main improvements for several variations of the present invention is the flexibility of the system. In the event of additional combustion, a new combined cycle power plant may include several different gas turbines along with a steam turbine manufactured in accordance with customer requirements to expand the operating range of the new combined cycle power plant.

Since productivity is defined as the volume of electricity generation divided by input energy, the study of the energy flow can be used to determine the steam cycle productivity required at a given power. Therefore, from the point of view of the combined capacity of the combined cycle, the energy generated is considered the combination of steam and gas turbine generation. Input energy is the total amount of input energy of the gas turbine in conjunction with the energy added to the burners due to additional fuel combustion. For this reason, with respect to FIG. 21, a formula for calculating the productivity (η) of a combined cycle in accordance with several preferred embodiments of the invention is presented as follows:

GTO - gas turbine production;

STO - steam cycle generation;

GTI - input energy of a gas turbine;

SFE - input energy PGRT due to additional combustion.

The parameters in the formula above are generally known. An unknown variable is the energy produced in the steam cycle. This indicator is a function of several other energy inputs, including the power of the steam turbine generator, exhaust gas loss of the HRSG, auxiliary load factor, and steam cycle productivity. Firstly, it is necessary to calculate the amount of energy that is transferred to the steam from the HRSG. This is defined as regeneration steam (2108) and is calculated using the following formula:

HRS = [HGI + SFE-HE] (1-HGL) (3)

HRS - PGRT energy transferred to the pair;

HGI - heat of exhaust gas turbine;

SFE - heat of additional combustion;

NOT - exhaust emissions PGRT (PV);

HGL - heat loss.

According to this formula, the heat transferred to the steam is calculated as the sum of the heat of the exhaust gas turbine and the heat added as a result of additional combustion, minus the loss of exhaust PGRT, adjusted for the loss of heat in the environment. At this stage, this determines the amount of energy involved in the steam cycle. To determine the generation of electricity from this energy, the generation of a steam cycle (2111), the input energy must be adjusted according to the performance of the steam cycle (2109), losses of the steam turbine generator (2110) and auxiliary load (2112). The formula for determining the power of a steam turbine generator takes the following form:

STO = HRS × SCE × AXF × (1-SGL) (4)

HRS - PGRT energy converted to steam;

SCE - basic steam cycle performance;

AXF - auxiliary load factor;

SGL - steam turbine generator losses;

= (1 - SGE (steam generator performance))

The value of the steam cycle output characterizes the conversion of steam energy into the energy of the shaft of a steam turbine, which is then corrected by the generation of electricity in the steam cycle, both by adjusting the capacity of the generator and by reducing the generated energy due to auxiliary load.

Knowing the results of these calculations, as well as on the basis of a given output for a given gas turbine (see Fig. 29 for the range of volumes of generated electricity for the preferred embodiments of the present invention), it is possible to determine the required steam cycle capacity, which will be the productivity of the preferred variant of a power plant with a combination cycles equal to the performance of a traditional power plant, known from the prior art and based on the use of the same gas tours in. FIG. 30 illustrates the steam cycle productivity levels that are needed when increasing the parameters of a combined cycle power plant described in several preferred examples of the present invention. Note that the parameter along the horizontal axis is the ratio of the energy generated by the steam turbine to the total energy produced by all gas turbines.

Using Figs. 29, 30 and the above formulas for calculating the steam cycle productivity and the total capacity of a power plant, the designer can determine which combination of gas turbines is most suitable in terms of energy productivity and economic benefits, as well as determine the relative complexity of the steam cycle (supercritical conditions steam, the amount of heating of the supply water, inlet temperature, etc.), which will constitute the desired capacity of the power plant. Figures 47-50 illustrate a design method for a power plant / structure providing for selection, design and financing of a power plant in accordance with a preferred embodiment of the present invention.

Flexibility of a Preferred Embodiment

As noted earlier, flexibility (transformability) is one of the advantages of the present invention. From the analysis of FIG. 30, it can be seen that, at a lower ratio of energy of steam / gas turbines, a steam cycle with a more moderate power can be used in order to provide the same capacity of the power plant in accordance with the present invention. However, with a low steam / gas turbine ratio, it would be possible to use ultra-supercritical steam modes to increase the power of the power plant by combining cycles known in the art. If in the implementation model shown in FIG. 26, at 725 MW, an ultra-supercritical lower cycle is used, the heat output will decrease from 6006 BTU / kW · h to 5912 BTU / kW · h.

However, unlike the preferred options with higher gas / steam turbine ratios, this configuration has less functional flexibility. With reduced ratios, the regulation of the preferred embodiment will be more similar to the prior art when it is necessary to modulate a gas turbine to control the power plant load at the highest operating point of the power plant. Depending on economic feasibility, high productivity, low productivity or capital costs determine which steam / gas turbine ratio the power plant designer ultimately chooses.

Possible ratio PT / GT and for the preferred option

On Fig shows the increase in the parameters of a steam turbine, achieved in accordance with several preferred variants of the present invention. A traditional combined cycle power plant may have a steam turbine capacity of nominally 55% of the total gas turbine capacity. Therefore, the total capacity of the power plant can be defined as 1.55 of the power of the gas turbine (1.00 for the gas turbine plus 0.55 for the steam turbine). According to this example, in accordance with the present invention, a steam turbine can be designed so that its power is 2.1 times higher than the power of the gas turbine and the total power of the power plant, thus, would be 3.1 times greater than the power of the gas turbine (1.0 for a gas turbine plus 2.1 for a steam turbine).

This example embodiment of the present invention provides an efficiency of 3.1 / 1.55 = 2.0 times more than is known from the prior art. However, in order to remain on par with the prior art in terms of productivity, the basic steam cycle productivity should be 48.3% (see FIG. 30). In supercritical modes of steam operation, with improved steam characteristics and when heating feed water, the basic levels of steam cycle productivity can reach the specified value. Therefore, some preferred variants of the present invention, provided certain gas turbines are used, are able to practically double the power of combined cycle power plants in comparison with the prior art, significantly reduce the amount of hardware that would be necessary in the prior art to achieve the same power, and at the same time achieve performance levels comparable to the prior art.

Since in accordance with the present invention it is assumed to use a HRSG with a single pressure level, and for effective use of a HRSG with a single pressure level, the flow of feed water in the low temperature section of the HRSG must be appropriate to ensure the absorption of energy of the exhaust gases of a gas turbine, the analysis shows that a steam / gas turbine needed to achieve such a goal should be approximately 0.75.

Based on the fact that the relative amount of energy generated by a gas turbine is 1.0, and the ratio of steam / gas turbine is 0.75, the total capacity of the power plant is 1.75 times the capacity of the gas turbine, which leads to the generation of gas turbines energy equal to (1.0 / 1.75) or approximately 0.57, i.e. 57% of the total energy generated by a power plant.

Design limitations

Although several preferred embodiments of the present invention suggest a wider range of combined cycle parameters for a given set of gas turbines than was available in the prior art, however, there are limitations to the design of combined cycle power plants using new technologies. The following are some of these limitations:

1. At a gas burning temperature in the inlet tubes above 1600 ° F, the HRSG will require a more expensive water screen design.

2. Based on the design of the water screen, the HRSG may be limited by the temperature of the gas burning in the inlet tubes at approximately 2400 ° F.

3. The exhaust gases of PGRT must contain a significant amount of oxygen to maintain the combustion of additional fuel.

4. Burners that provide additional heat to the HRSG should be able to maintain low NOX levels even with high combustion parameters.

5. The cycle should be designed so that it is possible to work within the design parameters for pressure and temperature of a steam turbine.

6. The cycle should be designed in such a way as to maintain appropriate productivity, cost, emissions, as well as any other restrictive parameters, the existence of which makes the project acceptable from an economic and environmental point of view.

In view of such limitations, FIG. 29 shows a range of rated power for combined cycle power plants described in some preferred embodiments of the present invention. Please note that these power plants are based on one or two gas turbines and cover a range of less than 150 to 1,050 MW. 29 is not intended to demonstrate all possible combinations that can be used in accordance with a preferred embodiment of the present invention, but is only examples of various gas turbines for illustration.

The impact of economic aspects on the design of the power plant

All design engineers who are specialists in this field study many power plant projects, in terms of their relative economic advantages, before choosing the final configuration of a power plant. This applies to prior art combined cycle power plants and combined cycle power plants using the system and method in accordance with the present invention. New power plants should be commercially viable if a decision is made on their construction.

A design engineer can study various options, such as an electrolytic type cooling tower with low cost and high auxiliary loads (electric fans with rotary electric drive) or a hyperbolic cooling tower with high cost, having a single small side load (natural draft air flow, fans not required). Thus, there is an economic assessment of energy savings depending on the capital costs required for such savings. Based on the current and expected energy costs, capital costs and other factors, the project developer selects the most economical option. The most effective option, in terms of energy conservation prospects, is not always the most cost-effective choice.

The same estimates will be required when considering some preferred variants of the present invention. Although ultra-supercritical steam modes entail higher steam productivity, increased fuel economy may not outweigh the added costs of more sophisticated hardware. At high interest rates, some preferred embodiments of the present invention will provide a significant increase in capacity with only a nominal increase in interest on price. Given low fuel costs, larger power plants may be acceptable without a corresponding increase in steam cycle productivity. Once again, it should be noted that some preferred variants of the present invention provide a specialist designer in this field and a power plant designer with a wide selection when searching for the most commercially viable option for a power plant.

When, in a number of embodiments, the present invention is considered to be primarily a steam rather than a gas turbine power plant, a combination of economic evaluations is usually key. Since these turbines will be large in size and with strong end exhaust streams, they usually use one, two or three exhaust pipe bodies, each of which has a dual-stream exhaust device. Fig. 51 is an illustration of a General Electric steam turbine according to the information in the brochure "Steam Turbines for Large Power Plants" by John K. Reinker and Paul B. Mason (GE informative document entitled "Steam Turbines for Large Power Applications" by John K. Reinker and Paul B. Mason (General Electric Reference GER-3646D, 1996)). The housing on the left side is a combined high / intermediate pressure section, and on the right are two larger exhaust sections (low pressure) for dual flows. Various sizes of exhaust sheaths are allowed, designed in accordance with the length of the blade at the last stage. There may also be significant differences in the cost of various skins.

The choice of the height of the blades of the last cascade of the steam turbine, the size of the exhaust casing and their number refers to very general economic estimates of large steam power plants. The steam cycle can become more productive as a result of an increase to the next larger exhaust casing or, possibly, as a result of the addition of another exhaust casing. However, the increase in steam cycle productivity must be weighed in comparison with the increase in the cost of additional hardware.

Another factor that is gaining importance is the size of the condenser and the heat dissipation equipment. Once again, a lower exhaust pressure results in a higher steam cycle performance, but the cost of equipment to achieve a lower exhaust pressure should not outweigh the fuel economy.

Considering the economic aspects, the developer must provide the designer with an operational scenario for the new power plant. Since the electrical load of the system is very dynamic and constantly changing, information is needed on the load profile, which presents examples of the load at a given power plant as a function of time. Fig. 31A is taken from the US Department of Energy report, DOE / EIA-0614, "Competitive Electricity Prices: Critical Pricing of Energy Supply Systems and the Financial Situation of Utility Energy Services" (US Department of Energy report numbered DOE / EIA-0614 entitled "Electricity Prices in a Competitive Environment: Marginal Cost Pricing of Generation Services and Financial Status of Electric Utilities"). 31A illustrates a typical hourly load profile for a system (power grid) during the day. Based on weekly information, such a profile indicates a lower load on weekends and holidays, and in accordance with the data for the year - the need for some adjustments based on seasonal changes. Since most power plants work part-time for most of their lives, an economically optimal option is obtained when designing a power plant with the expectation of working with the highest capacity at average load, as opposed to the rated capacity of the power plant.

This is noted by M. Boss in the General Electric information brochure, “Steam Turbines for STAG Combined Cycle Power Systems” (M. Boss, GE informative document GER-3582E (1996), entitled “Steam Turbines for STAG ™ Combined-Cycle Power Systems”). As explained in this paper, despite the fact that the steam cycle productivity can be increased to a maximum, if the circular velocity of the steam turbine exhausts at the blade of the last cascade is approximately 550 feet per second, the economically optimal value is usually achieved at a circular speed from 700 to 1000 feet per second at rated steam turbine output. James S. Wright in the General Electric information brochure, Aspects of Optimization, Evaluation, and Testing of a Steam Turbine Cycle (James S. Wright, GE informative document GER-3642E (1996), entitled "Steam Turbine Cycle Optimization, Evaluation, and Performance Testing Considerations" ) offers selection criteria for the exhaust turbine housing. In this example, the choice is made between three exhaust bodies of different sizes, taking into account the increase in efficiency of the exhaust body as its size increases. The largest housing is not chosen, since an increase in its effectiveness is associated with additional cost indicators. Guided by the same logic, they also do not choose the smallest case. Thus, usually choose a medium-sized case, which is optimal from an economic point of view.

PGRT with a single pressure level

In order to make PGRT effective in the single pressure mode, it is necessary to first consider the effectiveness of its project. 11 shows a vapor enthalpy curve versus temperature for a pressure of 1800 psi. inch. As you can see, the heat content (enthalpy) of the vapor is not a linear function of temperature. This phenomenon greatly complicates the heat transfer by exhaust gases having an almost linear characteristic (see FIG. 12). As can be seen from FIG. 11, at the boiling point at 621 ° F, the enthalpy of the water / steam mixture increases from 648 BTU / lb to 1154 BTU / lb without any increase in temperature. The amount of heat absorbed in this section of the PGRT (evaporator) will be significantly greater than in any other section for a given temperature change.

Between temperatures between 100 and 400 ° F, the average heat capacity of water reaches 1.014 BTU / lb / ° F. This value is usually linear and varies slightly with pressure. Therefore, in this section, the heat transfer between water and exhaust gases will be relatively constant.

To maximize the efficiency of heat recovery in PGRT, as well as to ensure that the maximum amount of steam enters the steam turbine, a system control method is required to optimize the flow of feed water / steam through each section of the PGRT. Such an optimization scheme will be incorporated into the instrumentation of the power plant control system.

Regulatory system

There are many possible control technologies for steam turbines, but flow control and sliding pressure control are popular. In the case of flow control, the steam turbine includes a set of valves configured to maintain the design inlet pressure. In the case of adjusting the sliding pressure, the pressure at the inlet to the steam turbine has the ability to "slip" or vary with the load (steam flow) of the steam turbine. For combined cycle power plants where a heat recovery option is used, it is often advantageous to use sliding pressure adjustment. This control method allows the use of high volumetric flow rates of steam turbine flows by using steam with a lower specific volume (low pressure) at partial load. This allows you to maintain the performance level of the steam turbine at or near the design level. In addition, lower pressure steam boils at a temperature lower than high pressure steam; therefore, lower temperature exhaust gases in PGRT associated with a lower load can provide greater vaporization.

Energy use

As previously shown, in order to produce high pressure steam in a HRSG, not only does it have to have a common energy component to produce steam (the total necessary BTUs), but the energy must also have the proper temperature to cause the necessary heat transfer. In addition, it is desirable to maximize the use of waste heat, as well as not to produce a large amount of hot water or not to significantly increase the temperature of the exhaust gases of the PGRT compared to the optimal values. The use of additional combustion becomes extremely useful to achieve such goals.

It will now be useful to consider the concept of PGRT for a single pressure level used in a gas turbine. As has been shown, such an option, being designed for exhaust gas temperature of PGRT 180 ° F, leads either to the formation of an excessive amount of hot water, or to a too high temperature of exhaust gases PGRT, known from the prior art. This is due to insufficient input heat - at higher temperatures, and also because of the excess heat available - at lower temperatures. To illustrate, imagine PGRT that added heat to exhaust gases without increasing temperature (not a very likely option for a gas turbine / PGRT assembly). Imagine that to add heat PGRT was designed in such a way that it absorbs more fuel and air, but without increasing the input temperature. This scenario involves the formation of more exhaust gas turbine and, consequently, more hot water. The effectiveness of PGRT will not change, only its power will change in proportion to the added heat.

This concept is important because it relates not only to PGRT, but also to the traditional practice of combining cycles when additional combustion is used. In the current level of technology, additional combustion increased the flow of steam, but did not lead to an increase in the efficiency of the steam cycle.

Due to the presence of an excessive amount of oxygen in the exhaust gas of a gas turbine (the oxygen level decreases from 21% in ambient air to approximately 12-15% in the exhaust of a gas turbine at full load) the fuel can burn directly in the HRS without the need for additional air. This practice allows the use of additional combustion to increase the temperature of the exhaust gases. Combustion and the heat recovery process for additional combustion is 99% effective, since only 1% of the heat received in the HRSG is lost in the environment. This is a significant improvement over the Rankine cycle boilers, which can have an efficiency of only 80 to 90%. The main reason for this difference between the traditional Renkine cycle and PGRT boilers is that the traditional Renkine cycle boilers absorb the cold ambient air for combustion and can then emit exhausts at temperatures from 350 to 400 ° F, and the preheated exhaust gas from the gas turbine temperatures from 800 to 1200 ° F, and then exhausts at temperatures from 160 to 200 ° F.

An increase in the energy level of exhaust gases as a result of additional combustion significantly increases their ability (heat transfer ability) to form high pressure and temperature steam. In addition, the high energy of the high end of the PGRT is offset or serves to balance the excessive energy of the low located end of the PGRT, typical in the combined cycle of the prior art.

In other words, the additional heat entering the PGRT inlet, which increases the temperature of the exhaust gases, can be transferred to the feed water stream, which did not have enough energy to turn into steam. Not only does the overall steam flow increase, but also the efficiency of the steam cycle due to the higher proportional ratio of high pressure steam.

Thus, adding fuel to the lower cycle, as well as providing additional heat, can be used to increase the overall efficiency of the lower cycle.

System Overview

On Fig presents a conceptual diagram of a combination of cycles with the addition of heat to the lower cycle. 13, the upper cycle fluid (1301) enters the upper cycle engine (1302), where fuel and / or heat (1303) is added to increase its temperature. The liquid carries out work, which is converted by the engine of the upper cycle into the mechanical energy of the shaft. This energy drives the load of the upper cycle (1304). A load is an electric generator, pump, compressor, or any other device that requires mechanical energy to start. The liquid discharged from the upper cycle engine is directed through the exhaust pipe (1305) to a heat recovery device (1306). In addition, fuel and / or heat (1314) is added to the upper cycle fluid at the point where it enters the heat recovery device. After passing through the device for heat recovery, the upper cycle fluid is discharged into the open reservoir (1307).

For this example, the top loop is an open loop. In other words, the upper cycle fluid is taken from a large reservoir and discharged into the same reservoir. The device for heat recovery (1306) receives part of the energy of the exhaust of the upper cycle and transfers it to the liquid of the lower cycle (1308). In this example, the lower cycle fluid is heated at a single pressure level on the high pressure line (1309). This liquid then enters the lower cycle engine (1310), where it drives the lower cycle load (1311). Again, the load is an electric generator, pump, compressor, or any other device that requires mechanical energy to start.

From the lower cycle engine, the lower cycle liquid enters the heat exchanger (1312), from where heat is removed to the outside. Then the lower cycle fluid enters the fluid transfer device (1313). For this example, the lower cycle is closed, which means that the liquid of the lower cycle is constantly circulating in a closed cycle.

The embodiment of the present invention shown in FIG. 13 differs from FIG. 5 in two main respects:

1. Fuel and / or heat added (1314) to a heat recovery device 1306 that is not added in accordance with FIG. 5; and

2. According to FIG. 13, there is only one liquid — high pressure liquid (1309) entering the lower cycle engine (1312), while in accordance with FIG. 5, high pressure liquid (509) is supplied to the lower cycle engine (512) ), intermediate pressure (510) and low pressure (511).

Using the method of adding fuel and / or heat to the lower cycle, not only the energy of the lower cycle is increased, but also the efficiency of the cycle, since then intermediate and low pressure liquids are transferred to the high pressure liquid level. A high-pressure fluid per unit mass of the flow can do more work than an intermediate and low-pressure fluid.

There are several different types of fluids that can be used in the concept of combining cycles, including water, air, steam, ammonia, coolers, mixtures, etc. The demonstration of the preferred model of this concept is not intended to limit the number of cycles used in the combined cycle, to limit the types liquids in a combined cycle with one specific type, limiting the pressure of a liquid that can be used in a cycle, or restricting any cycle to its open or closed type, but only shows that process of increasing the thermal efficiency when combined, cycles can often be accomplished through the strategic use of additional fuel and / or heat input.

Heat Transfer Analysis

PGRT Low Pressure Economizer Section

As mentioned earlier, the problem of producing high pressure steam in conventional power plants with combined cycles is the distribution of energy between exhaust gases and steam generated. In addition, to optimize heat recovery, it is desirable that the temperature of the exhaust gases upon exit from the HRSG be optimal. Therefore, it is necessary to study the heat recovery process more deeply.

To optimize heat recovery in the lower temperature zones of the PGRT (approximately 470 ° F is the temperature of the inlet exhaust gases, up to 180 ° F is the temperature of the exhaust gases), a significant amount of heat must be removed using water supplied under pressure. The average heat capacity of the exhaust gases in this range (from 470 to 180 ° F) is 0.257 BTU / lb / ° F (this value may vary slightly with changes in the oxygen content in the exhaust gases / amount of additional combustion). At temperatures between 100 and 400 ° F, the average heat capacity of water is 1.014 BTU / lb / ° F.

Therefore, to obtain an increment in feed water temperature corresponding to a decrease in exhaust gas temperature, the flow ratio should be (1.014 / 0.257) or 3.95 Ibs (lbf) of exhaust gas per 1b (lbf) of feed water in a given temperature range PGRT. The ratio of the flows at this level or close to it will provide optimization of heat recovery for this section of the HRSG. Changes in parameters, such as the oxygen content in the exhaust gases, the temperature of the feed water at the inlet, and other factors can be monitored using the hardware of the power plant control system, and it is also possible to calculate and adjust the optimal amount of feed water through each section of the PGRT.

Experience has shown that maintaining the temperature of cold water at the inlet of the low-pressure economizer section (feed water directly from the condenser) can have a detrimental effect on the durability of the component parts of this section. This is due to corrosion in the economizer as a result of the fact that the tubes and fins of the economizer are cooler than the dew point of the exhaust gases of the HRSG. Since these components are usually constructed of carbon or low alloy steel, moisture condensation on the surfaces of the tubes and plates corrodes them, reducing the efficiency of the heat exchanger. Two common methods are used to combat this problem. One uses a device for pre-heating water so that warmer water enters the economizer. Another way is to make parts of the low-pressure economizer from a material that is not subject to corrosion, for example, stainless steel. Any of these methods is acceptable, and usually one that is determined to be economically optimal is selected.

Economizer Section High Pressure PGRT

The economizer section of the high-pressure PGRT heats the feed water from approximately 400 ° F (leaving the economizer section of the low pressure) ideally to the saturation temperature of the pressure in the evaporator section. For the 1800 psia intermediate pressure used in this example, the saturation temperature is 621 ° F. In this range, the average heat capacity of the feed water is 1,230 BTU / lb / ° F. To heat this water, the exhaust gas of a gas turbine must enter the section at a temperature approximately 50 ° F above the temperature of the supply water at the outlet, or 671 ° F. The average heat capacity of the exhaust gases in this range (from 671 to 470 ° F) is 0.264 BTU / lb / ° F (this value may vary slightly with changes in the oxygen content in the exhaust gases / amount of additional combustion). Therefore, for this PGRT section, the flow ratio should be (1,230 / 0,264) or 4.66 Ibs of exhaust gas per lb of feed water. Since this flow ratio is not consistent with the optimal flow ratio of the low pressure economizer, adjustments will be required to compensate for this discrepancy (different optimal flows through each section).

PGRT Evaporator Section

The section of the evaporator (subcritical mode of use) is unique compared to other sections of the PGRT, in terms of the fact that its inlet and outlet temperatures are constant (for the operation mode at constant pressure). This increases the stability of the heat transfer process, and the logarithm of the average temperature difference (LMTD) fluctuates less with differences in flow than the same value in other sections, since the temperature of the outlet is essentially constant. LMTD is a non-linear variable heat transfer that is used to determine the heat transfer ability of a heat exchanger.

Due to this constant temperature parameter, the sections located under the evaporator - economizers of high and low pressure, have relatively constant (with some changes in pressure / load) inlet temperatures.

However, in a number of preferred embodiments, additional combustion significantly changes the inlet temperature of the evaporator section, as well as the superheater and intermediate heater sections. Such increasing and decreasing temperatures determine the steam flow through the HRSG, and therefore, the production of a steam turbine. Therefore, in contrast to the economizer sections, the optimized flow ratio is not applicable for higher sections of the PGRT.

Since the PGRT evaporator section absorbs a large fraction of the available heat and also actually forms steam, the evaporator power is modulated mainly by the input temperature of the exhaust gases, which is generally a function of the input temperature of the exhaust gases of the PGRT. Therefore, the control of this section is carried out mainly through the incoming fuel.

Sections of superheater and intermediate superheater PGRT

Both of these sections are similar, as they both heat the steam to a higher temperature. The superheater section receives saturated steam from the evaporator section and heats it to the inlet temperature of the high pressure turbine. A desuperheater is used to adjust the temperature to the desired value at the outlet of this section.

The intermediate superheater section receives steam from the turbine low pressure section and overheats it to the inlet temperature of the turbine intermediate pressure section. At the exit from this section, to control the temperature of the superheat, a device for removing heat from the superheat (desuperheater) can be used, which occurs due to the efficiency of the cycle. As noted by authors Eugene A. Avallone and Theodore Baumeister III in Marks' Standard Handbook For Mechanical Engineers (Ninth Edition) (ISBN 0-07-004127 -X, 1987)) in sections 9-24 to 9-25:

“Thermoregulation of superheated steam by direct spraying of water ... leads to an equivalent increase in the production of high-pressure steam without thermal losses ... Typically, thermostatic sprinklers are not used to control the temperature of superheating of steam, since their use reduces the overall efficiency of the thermal cycle processes. However, they are often installed in case of controlling the temperature of the superheat of steam in emergency situations. "

On Fig shows curves reflecting the thermal conditions of the sections of the superheater and the intermediate superheater as a function of flow. The curves do not show a small effect from the desuperheater, the released flows, the heat loss in the pipes or other small corrections. It is necessary to pay attention to the fact that both sections require a proportional change in the amount of heat with a change in flow (load of a steam turbine). Therefore, it may be advantageous, although not necessary, to manufacture these sections as one section in a HRSG so that each has its own corresponding surface area for heat transfer.

PGRT Surface Area

To obtain the necessary heat transfer from the exhaust gases of a gas turbine to water / steam, it is required that in each section there is sufficient surface area for heat transfer. The working equation that describes the general heat transfer process is:

Q = U × A × LMTD (5)

Q - heat transferred, in BTU / hr

U - total heat transfer coefficient, in BTU / hr / ft ² / ° F

A is the total surface area in sq. feet LMTD is the logarithm of the average temperature difference with the logarithm of the average temperature difference, defined as

Where

GTTD - large output temperature difference;

LTTD is a smaller difference in output temperature.

The difference in outlet temperature is:

1) the temperature of the exhaust gases entering the PGRT section, minus the temperature of the water or steam outside the section, and

2) the temperature of the exhaust gases leaving the PGRT section minus the temperature of the water or steam inside the section.

Obviously, the larger is the large difference in temperature at the outlet and the smaller is the smaller difference in temperature at the outlet. If they are equal, any of them is equal to the logarithm of the average temperature difference. If the large or smaller temperature difference at the outlet becomes too small, the surface area (A) must be very large in order to compensate for the lack of value. Since the surface area is essentially the entire effective surface area of all the tubes and plates in the PGRT section, increasing the area increases the size, weight and cost of the PGRT.

Another parameter in the heat transfer equation (U) is determined by the surface coefficient of heat transfer between water / steam and the inner wall of the pipe, the thermal conductivity of the pipe material and its thickness, as well as the surface coefficient of heat transfer between the exhaust gases and the outer wall of the pipe.

For general purposes, the determining parameter in this equation is the surface coefficient between the exhaust gases and the outer wall of the pipe. This is because it is here that there is the greatest resistance to heat transfer, and just as it is typical for a group of series-connected resistors in an electric circuit, the largest resistance controls the flow. Consequently, the factors most influencing the change in the heat transfer coefficient represent the greatest problem for engineers designing HRSGs and choosing surface areas for each section.

From a management point of view, the selection of space in each section is important, since after the construction of the HRSG these areas can no longer be changed, i.e. they become a fixed value. Factors that affect the change in the value of U are parameters that change the speed of exhaust gases on the surface of the pipe. The prevailing deviation is the change in exhaust flow. Since a gas turbine is a constant volume unit, such a change takes place with changes in ambient temperature. In addition, it can occur with a change in the load of the gas turbine. If it were possible to minimize the influence of these factors, PGRT would be easier to design and optimize for work in a narrow range.

As will be demonstrated in an example of a preferred embodiment of the present invention, the proposed system and method allows a gas turbine to operate at full load (temperature control) over a wide range of the combined load of a power plant with combined cycles. This is markedly different from the prior art, which uses changes in the load of a gas turbine to modulate the total load of a power plant with combined cycles. Therefore, at some operational points, the only significant changes in the HRSG flow will be attributed to changes in the ambient temperature (additional combustion fuel adds less than 1% to the exhaust gas flow). At ambient temperatures of -20 to 100 ° F, the exhaust gas flow will vary by approximately ± 13%. And in the case of a gas turbine, known from the prior art, only one load change could lead to significant changes in the flow of exhaust gases. The General Electric PG7241 (FA) gas turbine model, at 55% load, produces only 70% of the total exhaust gas load. With changes in the environment, such a change in the total flow can reach only 61% of the designed PGRT flows. This difference between the flow and the design leads to the inefficiency of the PGRT and requires design compromises to coordinate with a wide range of operating conditions.

Due to the high temperatures in PGRT resulting from constantly maintained additional combustion, the logarithms of the average temperature difference observed in a number of embodiments of the present invention are more significant than in the prior art. Therefore, the required surface areas are reduced, and the total size of the HRSG can be smaller. This leads to significant cost savings in terms of both the construction and the occupied area of the object.

PGRT Management

In the prior art, PGRT control capabilities for balancing heat transfer have been limited. The control of the heat removal of the superheat in the superheater and the intermediate superheater was common. Additional combustion to control the production of steam is usually not used due to its negative impact on productivity and the associated additional costs. In the prior art, bypass pipes were sometimes used around certain sections of the economizer and feed water.

In some preferred embodiments of the invention, steam production is essentially controlled by the fuel consumption for additional combustion. More energy input means more steam is generated. For better control of heat transfer, sections of multi-tube burners arranged in rows can be used. This arrangement allows you to position the fuel (heat) supply at more than one point along the exhaust gas flow from the HRSG and, taking into account the limited heating and subsequent cooling of the section, in several places along the HRSG, serves to lower the overall temperatures of the HRSG (possibly avoiding the construction of a more expensive water screen).

As in the prior art, superheater methods will be used in the superheater to remove the heat of superheated steam, while the heat of superheat in the intermediate steam superheater should be limited by emergency control of superheated steam temperatures. The temperatures of superheated steam can be kept in equilibrium due to the careful selection of heat transfer areas in the HRSG and by reconciling the flow balancing in a multi-threaded superheater design. The flow of feed water through the high-pressure economizer is regulated by the optimal ratio of exhaust gas / feed water flows, just like the flow in the low-pressure economizer. According to this embodiment, the HRSG having six sections and operating at the same pressure level is a much simpler object to control and adjust than the 12-section boiler operating with three pressure levels, according to FIG. 6, illustrating the prior art.

Comparison of PGRT - Preferred Embodiment of the Present Invention and Prior Art

PGRT in a number of examples in accordance with the invention can be largely similar to PGRT of the prior art, in particular in that they have a large number of pipes for transporting feed water, recovering heat from the exhaust gases of a gas turbine and transferring it to water / steam in pipes through convective heat transfer. This device is very large. Both options PGRT, i.e. a variant of the present invention and known from the prior art, are placed in a large casing directing exhaust gases from the exhaust outlet of a gas turbine into the exhaust riser of the HRSG. PGRT can be a horizontal or vertical structure, depending on the requirements and mechanical limitations of the structure.

Some preferred embodiments of the invention, however, will use only one level of pressure. This does not exclude the use of additional pressure levels, just a single pressure level most clearly reflects the model of the preferred mode of the embodiment of the invention. This option differs from the prior art in which multi-level pressure HRSG was used for maximum heat recovery.

In the case of using a single pressure level and design involving constantly maintained additional combustion, in the preferred embodiment of the HRSG, a smaller area for heat exchange may be required than in the prior art. This will reduce the size, distance in the service area, weight and cost. However, some of the savings will be spent on the purchase of more heat-resistant materials and / or the possible construction of a water screen in this version of PGRT.

With a smaller total surface area in the preferred HRSG, it is likely that the back pressure of the exhausts experienced by the gas turbine in connection with the HRSG will be reduced. This will serve to increase the production of a gas turbine and its efficiency. However, additional combustion tends to increase this back pressure, which would reduce some of the performance benefits achieved as a result of reduced exhaust restrictions.

Due to the flexibility that the use of the present invention provides for the steam cycle, gas turbines will operate at full load over a wide range of gross power plants with combined cycles. This circumstance will allow to ensure a more constant flow in PGRT, provide for a more optimal design and eliminate losses in operational efficiency under partial load conditions.

With only one pressure level, the control and monitoring conditions of the HRSG in accordance with the invention are greatly simplified. Even with small changes in flow and / or temperature in the HRSG according to the invention, it is possible to make small corrections to the flows in the feed water / steam sections to compensate for these changes. With additional sections, more significant changes in the exhaust gas flow, and with a less complete control system, PGRTs known in the art are more susceptible to externally continuously varying system parameters, in contrast to the preferred embodiment of the present invention, which is a proactive control system.

New fully combined cycle power plant

The new power plant, fully operating in a combination cycle mode, according to the invention is similar to the power station of the prior art, but has both small and significant differences. Further, in comparison with the prior art, basic equipment samples, their functioning, and also the impact of costs will be considered.

Gas turbines

Gas turbines used in a number of preferred embodiments of the present invention may be standard gas turbines known in the art. The only difference relates to operating conditions related to the degree of differential pressure through PGRT according to a preferred embodiment of the invention. The basic engine, controls, layout, and overall organization may remain unchanged compared to the prior art. Consequently, in connection with the main details of the equipment, no engineering or design costs will follow. This will allow the use of proven technology and maintain a high level of reliability of the power plant. Obviously, to increase the efficiency gas turbine can be applied operational improvements of gas turbines, such as inlet cooling, evaporator cooling and other similar methods.

Heat Recovery Steam Generators (PGRT)

PGRT according to a preferred embodiment of the present invention can be smaller, more compact, operate at the same pressure level, provide controlled heat transfer and are optimally suited for continuously maintained additional combustion. In the presence of a single pressure level, in contrast to the multi-level pressure mode, the operation of some PGRT in accordance with the invention is easier to monitor and control. To optimize heat recovery and cycle performance for each given operational parameters, control mechanisms that control the combustion rate, flows in the sections, and / or temperature at the outlet of the section can be used.

With the operational flexibility embedded in the steam cycle, gas turbines will be able to operate at full load in a wide range of power plant loads, ensuring a constant flow of exhaust gases in the HRSG and, thus, ensuring more efficient operation. Fewer pressure levels, higher cycle throughputs, more consistent operation — all result in increased reliability and lower operating and maintenance costs.

The need for more heat-resistant materials or, possibly, for the construction of a water screen for a number of PGRT implementation cases in accordance with the present invention will lead to an increase in the initial cost, as well as an increase in operating and maintenance costs. It is doubtful that such an increase in costs will be higher than the amount of money saved as a result of eliminating unnecessary pressure levels, associated control devices and additional heat transfer area.

Currently, no PGRT has been described similar to those that demonstrate the advantages of the present invention. However, traditional boilers of steam power plants have been built for decades, and, of course, this technology can be applied in some versions of PGRT according to the invention. In addition, many HRSGs were built to operate in multilevel pressure or a single pressure level, and many of them were designed for some degree of additional combustion (including the construction of a water screen from high temperatures). Of all the main components in the implementation according to the invention, this will require the greatest efforts of engineers and designers. However, as mentioned earlier, PGRT with constantly maintained combustion and a uniform pressure level is a new concept, but does not go beyond technological practice and is available for development by specialists in this field.

Steam turbines

In the prior art, steam turbines were designed primarily for the heat generated by the HRSG. For large power plants with combined cycles, an empirical rule applies, according to which the power of a steam turbine reaches approximately 50% of the output of a combined gas turbine. With the use of additional combustion, this percentage can increase, but due to the negative effect on productivity that was observed when using hardware, systems and methods of the prior art, such an increase was usually small. General Electric's informative document GER-3574F (General Electric informative document GER-3574F (General Electric informative document GER-3574F, General Electric informative document, General Electric informative document GER-3574F, General Electric informative document, General Electric informative document GER-3574F 1996), entitled "General Electric Combined-Cycle Product Line and Performance" by David L. Chase, Leroy O. Tomlinson, Thomas L. Davidson, Raub W. Smith, and Chris E. Maslak), according to Table 14, that additional combustion in PGRT can increase the output of a power plant by combining cycles in the prior art by 28%, but only if the heat is increased by 9% oic total power of the combined cycle (when a certain fuel consumption).

In the prior art, the main focus was on PGRT with multiple pressure levels and steam turbines using such steam. Therefore, these steam turbines were characterized by relatively small high-pressure flows, medium-sized intermediate-pressure flows due to the addition of intermediate-pressure steam from PGRT and relatively large high-pressure flows due to the further addition of low-pressure steam from PGRT. This leads to lower volumetric flow rates for the high and intermediate pressure sections of the steam turbine, as well as potential problems with the load on the end exhaust of the low pressure sections. In addition, the steam cycles themselves were somewhat inefficient, since the intermediate and low pressure steam generated by the HRSG has a smaller potential energy reserve for the work that high pressure steam performs. As a result, intermediate and low pressure steam flows negatively change the requirements for the load on the end exhaust of the steam turbine, as well as the heat removal.

Due to problems with volumetric flow rates and cost / benefit ratios, inlet pressure ratings for steam turbines of combined cycle power plants were limited to approximately 1800 psi. inch. When using PGRT with several pressure levels, there was no need to use traditional heating of feed water, since PGRT always had excess heat to perform this function. Thus, increasing the productivity of the steam cycle through the improvement of these particular characteristics is usually not considered.

A steam turbine used according to a preferred embodiment of the invention can be large and more efficient than the prior art. The steam turbine used in accordance with the invention can provide an efficiency of about 0.75-2.25 (or more) times greater than the total power of the gas turbine. With an equivalent number of gas turbines and PGRT, providing combustion at temperatures up to 1400 ° F, the total capacity of the power plant with combined cycles can increase by 2.00 times or more, compared with the prior art. This is equivalent to the use of a steam turbine in a number of cases corresponding to preferred embodiments of the invention, which can be designed for nominal values 4.50 times greater than the same values of a steam turbine of the prior art (for a steam turbine that was connected to the same gas turbines) .

The steam turbine in accordance with the invention may be similar to the steam turbine of the prior art, however, most likely, it will have higher characteristics of the input pressure. In addition, the steam turbine in accordance with a preferred embodiment of the invention may involve the use of steam extracted by heating the feed water, which will increase the productivity of the steam cycle. Without the use of intermediate and low pressure steam from PGRT, the steam flow into the high pressure section of a steam turbine under standard conditions will be maximum through any section. This will increase the volumetric efficiency in the high pressure section. From this point of view, the steam will be removed from the steam turbine to be directed to various feed water heaters, fuel heaters, a smaller steam turbine operating from the boiler feed pump and / or for other types of work within the power plant. This operation reduces the flow at the end exhaust, reducing the possibility of overloading the end exhaust of the steam turbine. All of these features are typical of a steam turbine that would be used in a traditional steam power station.

Due to the large increase in the nominal values (from about 50% of the total power of the gas turbine to 100-200% of the total power of the gas turbine), the steam turbine may require larger blades of the last cascade and / or more low-pressure sections. This will entail a relatively small increase in power costs compared to the cost of attracting additional gas turbines, PGRT, switches, transformers, foundations, etc., which would be required to increase power in the framework of the prior art.

With the exception of greater throughput, higher nominal operating parameters, improved efficiency and a larger set of blades and / or additional sections of low pressure, a steam turbine may look like a steam turbine in the prior art. Typically, it is designed to extract a steam stream from a turbine for traditional heating of the feed water, and not to draw the stream into the turbine from the intermediate and low pressure sections of the HRSG. However, a steam turbine is extremely similar to a steam turbine with the same operating parameters and inlet conditions as in modern traditional steam power plants. Therefore, the new cycle combining method allows the use of more traditional and more efficient steam turbine hardware and more efficient steam turbine cycles. This maximizes the productivity of the lower cycle, significantly increases power, reduces the overall size of the power plant with combined cycles and the cost of installing equipment without any loss in reliability.

Mode of operation

With a high level of additional combustion (large flow ratios of a steam turbine / gas turbine) and the ability to vary these combustion parameters, the options of a power plant in accordance with the invention represent a model in which the lower cycle is much more independent than in the prior art. Due to this phenomenon, as well as the fact that, from the point of view of emissions and performance, it is better to control gas turbines at full load, most load options of a power plant with combined cycles in accordance with the invention are achieved by changing the level of additional combustion and, therefore, the load of a steam turbine, while gas turbines continue to operate at full load or close to full load. This differs from the prior art when additional combustion was used to obtain only a small increase in power generation by the power plant during peak operation, and the total load of the power plant was controlled mainly by changing the load of the gas turbine.

In some preferred embodiments of the invention, at a combined full load of a power plant, gas turbines will operate at full load, with either the HRSG reaching its limit on the combustion temperature or the steam turbine reaching its limit on the inlet pressure. From this point of view, with a decrease in the power plant load, additional combustion is also reduced, steam generation is reduced, and therefore the steam turbine load is also reduced. The process of reducing the load continues until the moment when the corresponding flows can no longer be supported in PGRT.

After the respective flows become impossible to maintain in the HRSG, the steam turbine and / or HRSG have reached the operational limit. At this stage, it will be necessary to reduce the load of a gas turbine or several gas turbines. After reducing the entire load on gas turbines, the load on the steam turbine can be increased to bring the system in line with the load requirements. An illustration of the intended operating mode with several gas turbines is given in FIG. 43. This control method can be used to reduce the load of the total full load of the power plant to the lower limit of the HRSG and / or steam turbine only by changing the level of additional combustion and due to the functioning of the gas turbine with full load. After reaching this lower limit, one gas turbine can be unloaded and then the corresponding PGRT will begin to produce less steam. Accordingly, another gas turbine may remain at full load or close to this level, and the corresponding PGRT can increase the level of additional combustion. This results in more steam entering the steam turbine. The overall result is a transitional zone of operation when the load on one gas turbine decreases, while the steam turbine compensates for most of this decrease in load. After reducing the total load of the power plant to a level sufficient to bypass this transitional zone of operation, one gas turbine is taken out of operation (preserved), while the remaining HRSGs will continue to carry out additional combustion at a high level, and the steam turbine will operate at a significantly higher load than the upper limit of the transition zone. Such a functioning scheme allows gas turbines to remain in full load mode or in a mode close to this level, in a wide range of the total expected generation of the power plant (from about 50 to 100% of capacity) and only with a small fraction attributable to operation in the transition zone, when one gas turbine is transferred from full load to idle state. In accordance with FIG. 43, such a transition zone is between 70 and 80% of the power plant load.

An exemplary model of the control structure that performs the above procedures is conceptually shown in the graphs of Figs. 16-19. A more detailed description of this model is given later in this document.

Exploitation

Since the level of additional combustion is comparable with the prior art, the performance of a steam turbine is significantly increased. When using PGRT capable of withstanding temperatures reaching 2400 ° F at the inlet, the steam turbine can be designed according to the design (for example) on the basis of achieving rated values approximately 2.25 times higher than the power of all gas turbines combined. This is significantly higher than the level achieved by steam turbines from the prior art, where in such variants the nominal power of a steam turbine usually reached a level of 0.4-0.6 times higher than the power of all gas turbines. This significantly increases the power of this power plant, since a steam turbine can now have indicators 4.50 times higher than the same parameters of a steam turbine of the prior art. Also, as mentioned earlier, the functional flexibility achievable in this case allows gas turbines to operate at full load over a wide range of total plant capacities. This increases the capacity of the power plant during partial load operation and reduces the NOX level for gas turbines, which are typically characterized by an increase in such emissions during partial load operation.

With this increase in power compared with the prior art, increased functional flexibility and reduced cost per kilowatt of energy, this model of a power plant with combined cycles, from a functional and economic point of view, is more suited to meet modern energy requirements at seasonal peak loads. In addition, insignificant functional variables (such as insulation of feed water heaters or functioning under pressure at the inlet of the high-pressure section by 5% higher than the nominal) will allow in this embodiment to achieve even greater power than the nominal, with only minor damage to the production rate . Since seasonal peak periods last only a few days a year, the proposed method is inexpensive to increase energy production during these periods (which can be sold at a very high price at the lowest cost). Increased income more than compensates for the shortcomings and increased cost of fuel, which also occurs during seasonal peak loads. This creates a cost-effective alternative for designers of power plants and utilities. As stated in POWER MAGAZINE (ISSN 0032-5929, March / April 1999, p. 14):

"Reserve limits are omitted throughout the country ... According to the Midwest exchange rate, the following figures are reported in the press: above $ 7000 / MW ($ 7.00 / kW · h), and the rate of $ 6000 / MW ($ 6.00 / kW · h) reported in Alberta ... ”

Other industry newsletters also cite price data at the level of $ 10,000 / MW ($ 10.00 / kWh).

Although according to the prior art, additional combustion reduces the overall thermal performance of the combined cycle, the above example of the invention showed that such a statement is incorrect. Due to the use of fuel additives in the additional combustion mode with the aim of not only adding heat, but also to increase the resources of the lower cycle, it becomes possible to achieve or exceed the overall efficiency of the combined cycles of the prior art. This is due to the use of higher inlet steam pressure, larger and more productive steam turbines, the conversion of the low pressure steam used in the prior art to high pressure steam, and the use of conventional heating feed water. Partial load operation is also improved because, according to the invention, gas turbines will operate at full load (at which they are most efficient) for most of their operating time (except when they are taken out of service).

Partial load operation

When reducing the system load, it is necessary to reduce the load of the power plant with a combination of cycles to bring into line with the requirements of the electrical system. In the prior art, this was achieved by reducing the load of gas turbines. This mode of load control leads to a decrease in the performance of gas turbines, as well as the overall performance of power plants with combined cycles. In some embodiments of the present invention, however, the load is controlled by changing the level of additional combustion. Thus, gas turbines remain at full speed or close to this level, remaining the most efficient and having the lowest level of emissions of harmful substances. The main part of the load change in this case is due to a decrease in the amount of steam generated and a subsequent decrease in the power of the steam turbine. This mode of operation will provide an increase in the level of productivity when working with a partial load for the entire power plant with combined cycles, as well as a reduction in the maintenance of gas turbines, as a result of a reduction in the cyclic repetition of thermal operations (the natural temperatures of a gas turbine usually change when its load changes).

On Fig at some levels of partial load reflects the levels of expected performance of conventional power plants with combined cycles, known from the prior art, along with the performance that can be achieved using some of the preferred options in accordance with the present invention. As can be seen from these graphs, when reducing the load from 100%, the operational performance of combined cycles in the prior art is constantly reduced from the optimal level. However, for some preferred embodiments of the present invention, there is an increase in productivity with an initial decrease in load from 100% until it begins to drop below about 80%. Such partial load performance information indicates that when implementing the present invention, significant fuel economy is achieved compared to the traditional combination of cycles in the prior art.

Operation at maximum load

The present invention is particularly suitable for providing energy during peak periods. During these periods, power generation in a combined cycle power plant in accordance with the invention may temporarily exceed its rated load. As mentioned earlier, this temporary excess of load compared to the nominal can provide a huge economic advantage, since energy at peak season can be sold at a price hundreds of times higher than the usual price for energy produced in normal time. Therefore, there is a great incentive for power plant owners to produce such energy. As mentioned earlier, in the prior art this problem was tried to be solved by using additional combustion in HRSG. This not only reduces the efficiency of the traditional combination of the cycle at peak loads, known from the prior art, but, due to the need for additional power of the steam turbine, its basic performance also decreases when operating under normal load outside the peak season (in the absence of additional combustion, the steam turbine already working at partial load). Thus, the ability to increase the energy production of traditional power plants at the peak of the season with combined cycles under different loads is accompanied by losses in the total capacity of the power plant.

The power of a steam turbine can be increased by increasing the mass of the stream. Accordingly, a technology that increases the flow of steam through a steam turbine usually increases its total power. Since the present invention primarily speaks of the Rankine cycle, the combined cycle as such increases the effect of the steam turbine on more serious changes in the total power of the power plant with combined cycles. Therefore, this effect on the production of a steam turbine is much more effective than in combined cycle gas turbine power plants, as follows from the prior art.

According to the table below, as the pressure increases, the steam flow also increases.

Examples of increasing peak energy Inlet Pressure, psia Inlet temperature ° F Specific Volume, ft ³ / Ib Mass / vapor flow, Ib / hr Volumetric flow, ft ³ / hr Steam flow increase, (new / old) 2400 1050 0.338245 2,000,000 676490 basis 2520 1050 0,320349 2111730 676490 1,055865 2520 1000 0.304021 2225143 676490 1,112571 2520 950 0.286872 2358162 676490 1,179081 2640 1000 0.288236 2346998 676490 1,173499 2640 950 0.271554 2491183 676490 1,245591

If such an increase in pressure occurs together with a corresponding decrease in the inlet temperature of the steam, a further increase in the mass of the flow is possible. At the same time, isolating the feed water heaters will ensure the direction of a larger steam flow to the exhaust of the steam turbine, further increasing its production. Unlike the prior art, this method provides the opportunity to increase the nominal peak power of the power plant with combined cycles and implement the present invention without compromising the overall performance of the power plant during operation under normal off-peak loads.

Cost

A significant advantage of the implementation in accordance with this preferred option is cost savings. As noted earlier, a PGRT power plant with design inlet temperatures up to 2400 ° F due to additional combustion can easily use a steam turbine with a capacity of 2.0 times higher than the total capacity of gas turbines. Therefore, the total capacity of the power plant is 3.0 times greater (2.0 steam turbines + 1.0 gas turbine) of gas turbine power. The combined cycle power plant in the prior art includes a steam turbine with a rated power of approximately 0.5 times the total capacity of gas turbines. Therefore, the power ratio is essentially 2.0 (3.0 / 1.5). In other words, a combined cycle power plant according to the invention has a power 100% higher than in the prior art. An example of such an increase is shown in FIG. 39, which shows the heat balance for a 1040 MW version using two industrial standard gas turbines of the GE PG7241FA model and a larger steam turbine. On Fig shows a combined cycle from the prior art, using the same composition and model of gas turbines and a standard smaller steam turbine, designed for a nominal power of 520 MW.

It follows that to ensure power equal to the power of the above example, a power plant with combined cycles of the prior art needs to increase units of equipment by 100%. This means the acquisition of additional gas turbines, another steam turbine, additional PGRT, switches, transformers and other necessary systems and real estate needed to maintain this equipment. This will increase the cost of installation and commissioning of the power plant by 100%.

At 1999 prices, the cost of installing a modern high-performance power plant with combined cycles can be estimated at about $ 450 per kilowatt of energy. Therefore, the construction of a power plant with a capacity of 720 MW (720,000 kW) will cost 324 million US dollars. If you increase the capacity of the power plant to 1,050 MW, the cost of the installation will reach $ 472 million. In contrast, in accordance with the present invention, it is possible to use fewer pieces of equipment to eliminate this undesirable rise in price by reducing the cost of a kilowatt of energy at the rated power of the power plant.

Retooling

Another important characteristic of the present invention is the possibility of its use for the modernization of existing power plants. Many existing steam power plants today produce costly electricity compared to the highly efficient combined cycle power plants discussed in this document. As the energy supply market becomes privatized, it will be especially important for electricity producers to maintain competitive prices. For this, technologies are needed that would allow existing power plants to be competitive compared to new combined cycle power plants.

Since, in accordance with a preferred embodiment of the present invention, it is assumed to use a single pressure level (mainly), a higher vapor pressure typical of steam turbines of traditional steam power plants, a higher power ratio of a steam turbine / gas turbine, and also small design dimensions are provided, this option Ideal for retrofitting existing steam plants. When using the present invention, large steam power plants may well do without their boilers and use steam directly from PGRT. This increases cycle productivity and (in many cases) significantly reduces emissions. This result can be achieved by using the existing steam turbine, condenser and other infrastructure already available at the power plant. Thus, owners can get a power plant with a combination of higher efficiency cycles with lower capital costs and minimal requirements for real estate.

Recommended Implementation Model — Typical Configuration

Overview

The configuration of some embodiments in accordance with the invention is similar to the prior art in that gas turbines and PGRT are used to generate energy and convert exhaust heat to steam. However, in embodiments of the invention, HRSG with constant combustion maintenance is used, producing a significantly larger amount of steam, moreover, in the mode of a single pressure level (or basically a single pressure level). This increased volume of steam (and usually of a higher pressure) drives a steam turbine, and is much larger in comparison with the prior art steam turbine associated with the same gas turbines.

Due to the large flows of feed water, it will be heated in PGRT, as well as in a separate closed heating circuit using traditional feed water heaters by extracting steam heat. To increase the efficiency of the cycle, gas fuel heaters are also used.

Development

On Fig presents a diagram of a model implementation of a preferred embodiment of a power plant in accordance with the present invention. The exhausts of each gas turbine (1520) are directed to the corresponding PGRT (1509) and drive the corresponding generator (1521). These exhaust gases produce steam in the HRSG, which then generates energy in a steam turbine and eventually undergoes condensation in a condenser (1595).

Feed water heating - PGRT feed water heating circuit

Condensate from the condenser (1595) enters the low-pressure boiler feed pump (1530), where it is pumped to the outlet pressure level. From this point, the low-pressure (low pressure) feed water control valve (1560) maintains an optimal flow through the low-pressure economizer (1501), while directing excess feed water flows to the first of a number of traditional feed water heaters (1533). The flow leaving the low-pressure economizer continues to move toward the low-pressure boiler feed pump (1531) and its pressure rises. From here it passes through a high-pressure economizer (1502). However, after exiting the low pressure economizer, part of the feed water flow is directed through the feed water balancing valve (1561) to maintain optimal flow through the high pressure economizer section (1502). In addition, some of the flow is directed to the gas fuel heater (1575) through the pipeline (1571). After preheating the gas fuel, this flow returns to the inlet of the low-pressure feed pump of the boiler (1530) through the pipeline (1572). The remaining feed water continues to flow into the high pressure economizer, and the stream leaving the high pressure economizer is mixed with the feed water leaving the last feed water heater (1537). This flow is located at the desuperheater valves (1510) and (1511), while the bulk of the flow continues to move to the evaporator (1504).

Feed water heating - traditional heating circuit

In a parallel feed water heating circuit, the last water passes through the first feed water heater (1533), where it is heated. This stream then moves through the second and third feed water heaters (1534) and (1535), respectively. At the outlet of the feed water heater (1535), the flow deviates from the parallel PGRT circuit through the feed water balancing valve (1561), mixes with this feed water and continues to move into the high-pressure feed pump of the boiler (1532), where its pressure rises. From this point, the flow passes through the fourth and fifth feed water heaters (1536) and (1537), respectively. Water from this heating circuit is now mixed with feed water from a parallel PGRT circuit and fed to the PGRT evaporator (1504) (except for the flow required for desuperheater valves (1510) and (1511)).

Evaporator

In the evaporator section, boiling water is converted to steam and enters the superheater section (1505). If the superheated steam is too hot, the desuperheater valve is moved to the position to spray condensate from the superheat heat removal pipe (1550) to the steam superheater supply pipe and control the input temperature of the turbine high pressure section (1590). In the high-pressure section of the turbine, the steam expands until the first extraction of steam occurs, when a small portion of the steam is removed from the turbine through a non-return valve (1568) into the pipeline (1558). This steam is fed into the fifth feed water heater (1537), preheating the water passing through it. Condensed steam from the fifth feed water heater flows down into the fourth feed water heater (1536). The steam in the high pressure section of the steam turbine (1590) remaining after the selection continues to move to the exit point from this section, becoming the so-called cold superheated steam. Cold superheated steam continues to move to the intermediate superheat section of the PGRT (1506).

Intermediate superheater

On the way to the intermediate superheater section, a certain amount of steam (second steam extraction) passes through a non-return valve (1564) to the pipeline (1554). This steam enters the fourth feed water heater (1536), which pre-heats the feed water passing through it. Condensed steam from the fourth feed water heater flows down to the inlet of the high-pressure pump of the feed water boiler (1532).

Cold superheated steam now passes through the PGRT intermediate superheater section to return to the intermediate pressure section of the steam turbine. If its temperature is too high, the desuperheater valve (1511) is moved to the position to spray condensate from the superheat heat removal pipe (1551) into the superheater supply pipe and control the input temperature of the turbine intermediate pressure section (1591). In the intermediate pressure section of the turbine, the steam expands until a third steam extraction takes place, when a small portion of the steam is removed from the turbine through a check valve (1567) into the pipeline (1557). This steam enters the third feed water heater (1535), preheating the feed water passing through it. Condensed steam from the third feed water heater flows down into the second feed water heater. The steam in the high pressure section of the steam turbine (1591) remaining after the selection continues to move to the exit point from this section, becoming the so-called transitional steam.

Transitional steam

Transitional steam continues to move to the low pressure sections (1592) and (1593) of the steam turbine. On its way to the low-pressure section, a certain amount of steam (fourth steam withdrawal) through the non-return valve (1565) is deflected into the pipeline (1555). This steam enters the second feed water heater (1534), which pre-heats the feed water passing through it. Condensed steam from the second feed water heater flows down into the first feed water heater (1533).

In the low-pressure sections of the turbine, the steam expands until the fifth steam take-off occurs, when a small portion of the steam is removed from the turbine through a check valve (1569) into the pipeline (1559). This steam enters the first feed water heater (1533), preheating the feed water passing through it. Condensed steam from the first feed water heater is returned through a pipe (1512) to the inlet of the low-pressure boiler feed pump (1530).

The steam in the high pressure sections of the steam turbine (1592), (1593), which was not removed, continues to move through the section to the exit point from it at the condenser (1595). The mechanical energy generated by a steam turbine drives a generator (1594), which produces electricity.

Low load operation

To operate at low loads, there is an insignificant flow of high pressure steam (therefore condensate flows from the condenser are insignificant) to maintain optimal levels of supply water passing through the HRSG. In this mode of operation, the valves (1560) and (1561) are closed. In the absence of a flow of feed water to remove heat, all steam extraction lines (1558, 1554, 1557, 1555, 1559) do not pass the stream. The entire flow of feed water passes through the HRSG, since the parallel circuit for the flow of water is closed.

When reducing the load from this point by reducing the steam flow (reducing additional combustion), the flow of feed water through the HRSG is no longer sufficient to absorb the heat of the exhaust gases and at the same time maintain the optimum temperature of the exhaust gases. Therefore, work below this point will increase the temperature of the exhaust gases and lower the performance of the combined cycle. At this stage, the design engineer must evaluate the operation parameters and determine whether it is more economically advantageous to reduce the load of gas turbines at this stage of operation or to continue modulating the intensity of additional combustion, allowing the temperature of the exhaust gas of the PGRT to rise. At some stage of work with a reduced load, however, it becomes economically viable to reduce the load of gas turbines.

Model of an embodiment of the invention - Power plant with a capacity of 725 MW

Overview

Consider, by way of example, yet another preferred embodiment of the present invention, a combined cycle power plant with a rated power of 725 MW. This sample power plant uses two General Electric gas turbines, model PG 7241 (FA). These gas turbines will emit exhaust gases, each in its own PGRT, with a common nominal pressure level of 2,400 psi. inch. An intermediate superheat steam turbine with a rated power of 400 MW will exhaust directly through a condenser operating at a pressure of 1.2 inches of mercury. Due to the large flows of feed water, the feed water will be heated in the HRSG. As well as in a separate feed water heating circuit using conventional feed water heaters by supplying steam. Gas fuel heaters will also be used to increase cycle performance.

Development

The General Electric gas turbine project has a design capacity of 170,770 kW according to the ISO standard, with a differential pressure of 3.0 inches in water in a PGRT and an exhaust pressure difference of 10 inches in water. The total capacity of gas turbines is thus 341,540 kW. On Fig schematically shows such a power plant. The numbers indicated at the various stages of the work process correspond to the “points” listed in the tables of FIGS. 36, 37 and 38. The data corresponding to these “points” in the tables of FIG. 36, FIG. 37 and FIG. 38 determine the pressure, temperature, enthalpy and flow at the corresponding “point”. In general, the information contained in FIGS. 35-38 represents the so-called “heat balance”, which is the total balance of mass and energy for a given cycle. It should be noted that in this example, deaeration (air exhaust) is carried out in a condenser.

Layout

On Fig presents a topological drawing of a power plant in accordance with some preferred variants of the present invention. It is very similar to the prior art General Cycle S207 FA power plant shown in FIG. 22. The most noticeable difference between the two schemes is the configuration of the steam turbine. In the prior art, a steam turbine has a relatively underloaded high / intermediate pressure section and one low pressure section. In some embodiments of the invention, the high / intermediate pressure section is comparable to the prior art, but has a significantly increased volumetric flow rate. To illustrate the effective use of higher steam flows at lower pressure, a second low pressure section is shown. However, depending on the economic assessment, this section may not be required.

Comparison with Prior Art

On Fig and 24 shows the layout of power plants S 207 FA company General Electric and 2X1 501G company Westinghouse, respectively. A General Electric power plant requires approximately 2.3 acres to house, and a Westinghouse power plant requires approximately 3.3 acres. The energy density in these two versions is approximately the same - 220 MW per acre. At the same time, some options in accordance with the present invention may have a design capacity of 725 MW, as shown in the example in FIG. 26, and an energy density of 315 MW per acre, or a design capacity of up to 1050 MW (see FIG. 29) and density energy 455 MB per acre. That is, a power plant implemented in accordance with the present invention allows to produce significantly more electricity in this fleet of equipment. This factor is not only an advantage for new facilities, but is also of great importance from the point of view of the possibility of modifying existing power plants, where the equipment fleet begins to operate above par.

In addition, combining cycles in the prior art is also more expensive in terms of fuel consumption, capital costs and technical support. 23 and 25 represent conditional economic data for the balances of power plants of General Electric Model S 207 FA and Westinghouse Model 2X1 501G, respectively. The numbers in the tables include annual fuel costs, capital costs and technical support for each power plant. Fig.27 also represents the conditional economic balance sheet data for a power plant in accordance with one embodiment of the present invention. Please note that in this case the figures for annual fuel costs, capital costs and technical support are less than the same figures for power plants with combined cycles from the prior art. Therefore, when using the present invention, the cost of generating electricity is reduced across all cost categories.

Power Plant Implementation Model - Supercritical Steam Mode

Overview

As another example, we consider an embodiment of a power plant with a combination of 1040 MW rated power cycles using an ultra-supercritical high-temperature steam mode. In this embodiment, the power plant uses two gas turbines manufactured by General Electric, model PG 7241 FA. Each of these gas turbines emits exhaust into its single pressure PGRT, projected at 3,860 psi. inch. In this embodiment, a double intermediate superheat steam turbine with a nominal power of 730 MW will also be used, which will exhaust directly through a condenser under a pressure of 1.2 inches of mercury. Due to the large flows of feed water, this water will be heated in PGRT, as well as in a separate feed water heating circuit using traditional feed water heaters through incoming steam extraction. Gaseous fuel heaters will also be used to increase cycle performance.

Development

The development of gas turbines at General Electric has a design capacity of 170,770 kW according to the ISO standard, with a differential pressure of 3.0 inches in water and a differential pressure of 10 inches in water. The total capacity of gas turbines is thus 341,540 kW. On Fig given a schematic representation of such a power plant. The numbers indicated at the various stages of the work process correspond to the “points” listed in the tables of FIGS. 40, 41 and 42. The data corresponding to these “points” in the tables of FIGS. 40, 41 and 42 determine pressure, temperature, enthalpy and flow at the corresponding “point”. This “heat balance” is the total balance of mass and energy for a given cycle. It should be noted that in this example, deaeration (air exhaust) is carried out in a condenser.

Comparison with Prior Art

Elevated vapor temperature (1112 ° F) and pressure (3860 psi) are consistent with data used in advanced steam cycles, sometimes called ultra-supercritical. For information on ultra-supercritical steam turbines and related cycles (see the information booklet “Steam turbines for ultra-supercritical power plants” by Klaus M. Redzlaff and V. Anthony Roigger (General Electric (GE) Reference GER-3945, 1996) (Steam Turbines for Ultrasupercritical Power Plants "by Klaus M. Retzlaffand W. Anthony Ruegger (General Electric Reference GER-3945, 1996). It should be noted that at an exhaust temperature of 1123 ° F, the standard industrial gas turbine manufactured by General Electric model PG 7241 FA does not have a sufficiently high-temperature exhaust energy for imagery anija such temperatures in steam at the required flow. Thus, these ultrasupercritical conditions even not considered available in the prior art. The supercritical steam power plant is comparable in accordance with the present invention, with subcritical steam power plant in accordance with the present invention, and the main difference is an improved performance. Higher vapor pressure, higher vapor temperatures, as well as the use of a second intermediate superheat provide additional power for this option. It should be noted that under such steam conditions and even with a large increase in power, the combined cycle capacity in the embodiment in accordance with the present invention approaches the performance of the prior art using the same gas turbine technology (6229 BTU / kW as opposed to 6040 BTU / kW) .

However, productivity is just one part of the economic formula. Other core costs, capital expenditures and technical support will be higher for the supercritical mode option compared to the subcritical mode. Therefore, as mentioned earlier, in order to determine the best option in an individual embodiment of a power plant with combined cycles in accordance with the invention, a complete economic analysis must be completed. In general, when fuel costs are high, supercritical options become economically optimal, and in the case of low fuel costs, subcritical options are preferred.

Power Station Load Profile

Power plants with energy distribution

As mentioned earlier, in order to maintain a constant frequency of energy (60 Hz for the United States), the electricity from all power plants integrated into the power supply network should be equal to the energy consumed by users of this network. For this reason, power plants “distribute” their capacity or are controlled by the Association of Energy Companies to meet the needs of the system.

As a result of distribution, most power plants will spend quite a bit of their working time at nominal output. Instead of operating at full rated power, many power plants will operate at medium load and will share the system load with all other power plants integrated into the power supply system. These statistics can be visually confirmed by examining the duration of the load on the curve of FIG. 31B, which presents data on a typical long-term distribution of the used load of a power plant in relation to the percentage of time. It should be noted that, taking into account such long-term observations, most power plants will operate at peak load for less than 10% of the time, and at average load - for 70% of their time.

On Figa shows information about a typical hourly load for the South Atlantic region of the United States over a 24-hour period. As can be seen from these data, work at a maximum load of 62,000 MW during the day significantly exceeds the time at a low load point of 40,000 MW. In addition, the total system capacity is probably higher than 62,000 MW, possibly at 70,000 MW (70 gigawatts, GW). This means that, with the exception of seasonal peak loads (hot summer days), even in the most active hours outside the peak season, many power plants do not operate at rated power. For this reason, power distribution plants may encounter significant changes in load and potentially spend only a few hours a year when operating at rated power.

To determine a typical steady-state load profile, the data in FIG. 31A is segmented. The time periods when the load was above 60 GW were determined as the peak of work. Periods with a load between 50 to 60 GW were considered the operation of the power plant at an average load, and periods with a load below 50 GW were considered work at night. Such a profile was considered as an average daily profile of the operation of a power plant. For weekends, eight hours a day was considered average workload, and the rest was considered night work (based on average weekdays for average and night workloads on weekends). On Fig shows the detailed data of these calculations. Using the calculation data of FIG. 32, a typical load profile used for comparison is as follows:

Period Average power of a power plant (%) Hours per week Night period 60 77 Average load 80 71 Peak loads 100 twenty

Note that although the power in accordance with FIG. 32 for peak loads is only 87.86%, this number has been aligned to 100.00% for the purposes of this discussion. The corrections for night and average power were less than 1% of the values in Fig. 32 and corrected for reduction to compensate for the correction for increase for operation at peak loads.

Power Station Load Profile Model

Using the data of the above table, you can use the calculated load profile to determine the annual power factor and the amount of fuel consumed at a given power plant with a combination of cycles based on the data of Fig. 33 at partial load operation. It is very important to emphasize on the basis of the table above and FIG. 31B that the efficiency of a power plant using prior art technology will rarely (if ever) reach optimal, from an economic point of view, operational performance. In contrast, embodiments of the present invention, as follows from FIG. 33, will always be more optimal than prior art models.

ECONOMIC ISSUES OF THE PRESENT INVENTION

Economic considerations

The costs of operating a power plant with combined cycles are diverse. However, the biggest costs for power plant manufacturers are usually fuel costs, capital costs (debt obligations), and maintenance. These three types of costs make up the bulk of the cost (expressed in cash per kilowatt hour) of electricity production in large power plants with combined cycles. Some of the items of expenditures are salaries, taxes, insurance, licensing and other types of expenses. In order to make an economic comparison of some of the implementation options in accordance with the present invention and the prior art, we will focus on three main types of expenses: fuel, debt and maintenance.

Fuel costs

The biggest cost element of a large modern power plant with combined cycles is the cost of fuel. Regardless of whether natural gas, oil, or other combustible fuels are used, a combined cycle power plant must consume large amounts of fuel to produce large amounts of electricity. In short, a power plant actually converts energy in one form (raw materials) into energy in another form (electricity). For this reason, since the function of the power plant is to carry out this conversion, the efficiency of the conversion process is the key to the economic success of the power plant.

Prior art power plants have an average power of about 48% (LHV) for older combined cycle power plants, such as General Electric S106B, and up to 60% (LHV) for the proposed advanced General Electric S107H model that has not yet been implemented in a commercial form. The indicated power levels are based on the lowest heat transfer coefficient of the fuel (LHV). However, such power levels occur when operating at full load, and, as indicated in Figs. 31A and 31B, most power plants operate at very low load for a very short time.

When operating with a partial load, according to FIG. 8, there is a power loss that would be expected with a reduced load for power plants with a combination of prior art cycles. Using this data, FIG. 33 shows a significant increase in the partial load performance level implemented in some preferred embodiments of the present invention compared to the combined cycle of the prior art (lower thermal power here indicates a more optimal functioning). This increase in productivity in partial load mode, together with an increase in efficiency in full load mode, allows in some embodiments in accordance with the present invention to achieve better economic indicators in terms of fuel consumption compared with the prior art.

Based on the load profile in Fig. 32 and using the level of thermal power (productivity) according to Fig. 33 in Fig. 34, the table summarizes the annual fuel costs for a combined cycle power plant in accordance with the invention compared to modern combined cycle power plants, known from the prior art. In any case, many of the embodiments of the power plant in accordance with the invention use less fuel per year than any of the prior art power plants.

Capital expenditures

Following the cost of fuel, the most significant type of cost for a new combined cycle power plant is the cost of fixed capital. This is the amount of cash needed to service a debt (loan payments). Although plant performance is an important factor, the overall cost of a power plant is also an important economic aspect. As mentioned earlier, it is necessary to evaluate the economic aspect of small parts of a power plant with combined cycles (i.e., large-sized exhaust sections of a steam turbine) in the same way as the economic aspect of an entire power plant with combined cycles. Small drops in the thermal capacity of a power plant (or a slight increase in efficiency) should not exceed compensation by increasing fixed capital. For this reason, power plant developers and engineers are striving to create the best economic alternative to existing power plants.

Due to their higher energy density, the use of fewer pieces of equipment and reduced construction costs, some embodiments of the present invention have significantly lower capital costs (reduced to 30%) than combined cycles in the prior art. On Fig the table shows the capital costs for the model of power plants with combined cycles in accordance with the invention in comparison with modern power plants of the prior art. In any case, a combined cycle power plant in accordance with the present invention requires significantly less financial investment than any of the prior art power plants.

Technical costs

Another large type of cost for owners of a power plant is the average cost of annual technical support, especially the cost of technical support for large equipment items. For a large power plant with a capacity of 725 MW of the prior art, as shown in the example, these costs can exceed 10 million US dollars per year. Therefore, power plants with a reduced size of the cost of technical support are economically viable.

Using a high energy density design that reduces the number of units of basic equipment, and using low-maintenance steam turbines as the main energy production machines instead of gas turbines that require frequent maintenance, the present invention will significantly reduce maintenance costs compared to combined cycles of the prior art. On Fig in a table presents data on maintenance costs for power plants, implemented in accordance with the present invention, in comparison with existing modern power plants with combined cycles in the prior art. In any case, the combined cycle power plant of the present invention requires less intense maintenance effort than any of the prior art power plants.

Comparison of total cost

Fig. 34 shows an economic comparison of an example of a combined cycle power plant implemented in accordance with the present invention and modern combined cycle power plants of the prior art. As can be seen from these data, an example of a power plant with combined cycles in accordance with the invention is less expensive to operate than power plants with combined cycles of the prior art in all three categories of costs: fuel, capital costs and technical support.

In addition, compared with Westinghouse's 2X1 501G power plant, NOX emissions are more than tripled, or about 180 tons per year. Over the 20 years of its existence, the combined cycle power plant of the invention has saved $ 469 million compared to the Westinghouse 501G combined cycle plant of the prior art. Such savings exceed the initial costs of $ 340 million required to construct a 2X1 501G power plant with a combination of Westinghouse production cycles and, from an economic point of view, represent a significant advantage for power plant manufacturers in a competitive market.

OPERATION OF THE PRESENT INVENTION

Example PGRT Control Method

Given the unique arrangement of equipment, the use of predominantly PGRT of a single pressure level and the need to optimize heat recovery, Fig. 16 shows an example of a control system to meet all these needs. The control system is an example of a combination of cycles described in the preferred embodiments of the invention shown in Fig.9 and Fig.15, although this system can be widely used in other embodiments of the present invention. In this example, there is one PGRT for each gas turbine. Note that the above example reflects the PGRT control system for this application and is a demonstration of the principles of flow control, optimal heat transfer, and combining PGRT and feed water heating circuits. In other applications, such a scheme may be modified taking into account particular circumstances. However, many principles from this management scheme will also apply.

According to FIG. 16, control starts at (1601) and continues with a block process (1602), where control over the control loop begins. Monitoring then proceeds to block operation (1603). At this stage, the control device examines the input from the operating unit (1611), which includes the ambient temperature and the load of the gas turbine (in particular, the exhaust of the gas turbine in the HRSG in this control loop). In accordance with the characteristic curve embedded in this software, the control device determines the exhaust stream of the gas turbine. Using the direct feed input for the steam flow required by the steam turbine and the steam flow already produced by other PGRTs, on the operation unit (1604), the control unit calculates the predetermined steam flow as the steam flow required by the steam turbine minus the flow from other PGRT. Control continues toward the decision block (1605) and compares the steam flow required by the HRSG with the optimal flow for the high pressure economizer.

If the power plant operates with reduced load, control passes to the operating unit (1606). At this stage of the work, the flow of the high-pressure economizer is below the optimum required from the HRSG. For this reason, the exhaust gases of the gas turbine will contain more heat than can be regenerated in the HRSG. As a first step in reducing the load, the control starts modulating the valves (960) and (967) in the closing direction to reduce the flow of feed water through the parallel heating circuit. After this circuit is completely disconnected, the second control phase reduces the energy production of the gas turbine. Now control returns to the initial operating unit (1602).

From the decision block (1605), if the flow necessary for PGRT is higher than the optimal flow of the high-pressure economizer, control passes to the operation block (1620). If the gas turbine operates at a lower load than the full load, the first control phase is to increase the load of the gas turbine. As soon as the gas turbine starts operating at full load, the valve (967) opens to start heating the supply water in a parallel circuit. Using the input data of direct supply to the operating unit (1610) about the pressure in the evaporator section and the temperature at the outlet of the high-pressure economizer, the valve (967) is modulated to ensure the optimum water temperature at the outlet of the high-pressure economizer. The pump (932) begins to operate as soon as the flow begins to pass through the valve (967).

Now control passes to the decision block (1621). If the steam flow required for the HRSG is less than the optimal flow of the low-pressure economizer, control passes to the operation unit (1622). At this power plant load, there is still no need for heating low-pressure feed water, since more than enough heat to heat the feed water is contained in the exhaust gases of the low-pressure economizer. For this reason, the valve (960) is closed. Management returns to the initial operating unit (1602).

From the decision block (1621), if the steam flow necessary for the HRSG is greater than the optimal steam stream of the low-pressure economizer, control then passes to the operation block (1623). At a given load of a traditional power plant, low-pressure feed water heating is necessary, since there is not enough heat in the exhaust gases to heat the feed water in the low-pressure economizer. For this reason, the valve (960) is changed to control the flow through the low-pressure economizer to its optimal value. Control returns to the initial operating unit (1602).

An example of a method for monitoring the entire station

While ensuring the logical organization of the management of the entire plant, some of the primary tasks include achieving improved efficiency and consistent low emissions. These goals are most achievable when operating gas turbines at full load or close to full load. The control logic for the combined control of the combined cycle in this example will focus on achieving precisely these goals. Obviously, the specialist understands that in order to achieve other goals, this management scheme can be easily reoriented to achieve other priorities.

Main control loop

In FIG. 17, control starts in (1701) and proceeds to the operation unit (1702), where control along the control loop begins. Management then transfers to the operation unit (1704). At this stage, the control device examines the input data of the operating unit (1703), including the current total station load and a link for the load (desired load level). In accordance with these data, the control device determines the requirements for changing the load. At the decision block (1705), the control device studies the need to change the load. If there is no such need to change the load, control returns to the initial operating unit (1702).

If a change in load is required, control passes to the decision block (1706), where it is necessary to determine whether a decrease or increase in the entire load of the power plant is required. If it is necessary to increase it, operational management should be transferred to the subroutine for increasing output (1708). An example of a variant of this routine is illustrated in the flowchart of FIG. If the load needs to be reduced, operational management follows the procedure for the output reduction subroutine (1707). An example of a variant of this procedure is illustrated in the flowchart of FIG.

Production Increase Subprogram

According to FIG. 18, the output increase subroutine begins at step (1801) and proceeds to the decision block (1802). If the power plant is not in the transition zone of operation (the stage where one gas turbine is in the process of commissioning or decommissioning), then control passes to the decision block (1804). Note that, according to Fig. 43, the transition zone of operation is in the range from 70 to 80% of the load of the power plant. To achieve various goals, the specified range can be changed by specialists.

If the power plant is in the transition zone of operation, the process proceeds to the transition control subroutine (1805). An example of a variant of this routine is illustrated in the flowchart of FIG. Then control returns to the end point of the subroutine of the operating unit (1803). All process returns to this block (1803) are returned to the subroutine block in accordance with FIG. 17 (1708), and finally to the initial operational unit for general control of the entire station (FIG. 17, 1702).

From the decision block (1804), subject to the operation of all gas turbines of the station, the process proceeds to the decision block (1820). At this junction, the control device determines whether all gas turbines are operating at full load. Since the best method to achieve the objectives is to operate all gas turbines at full load, then if all gas turbines do not work at full load, control passes to the operation unit (1821), where the load on one or more gas turbines increases. Now control returns to the endpoint of the subroutine of the operating unit (1803).

From the decision block (1820), if all gas turbines operate at full load, control passes to the decision block (1822). This block determines if the steam turbine or PGRT is not operating at its upper limit of operation. For PGRT, this is usually the temperature of the additional combustion, for a steam turbine - the value of the inlet pressure. It may also be a functional limit on performance and another parameter. If any of these limits is reached, control passes to the operating unit (1823), which includes a light in the control room, signaling to the operators that the power plant is operating at full capacity. Now control returns to the endpoint of the subroutine of the operating unit (1803).

From the decision block (1822), if the steam turbine or PGRT did not reach the upper limit of operation, control passes to the operation block (1824), where the fuel flow to the PGRT increases. Now control returns to the endpoint of the subroutine of the operating unit (1803).

From the decision block (1802), if all gas turbines do not take part in the operation of the station, the process proceeds to the decision block (1810). At this junction, the control device determines whether all the gas turbines participating in the operation are operating at full power. Since the best way to achieve these goals is to operate gas turbines at full load, if all gas turbines have not reached full load, control passes to the operation unit (1811), where the load on one or more gas turbines increases. Now control returns to the endpoint of the subroutine of the operating unit (1803).

From the decision block (1810), if all gas turbines are operating at full load, control passes to the decision block (1812). This block determines whether the steam turbine or PGRT has reached the upper limit of operation. In addition to the temperature or pressure limit, this may also be an operational limitation on performance or other system requirements. If one of these restrictions is reached, control passes to the transition control subroutine, the operation unit (1813). An example of this procedure is presented in the flowchart of FIG. Then, control returns to the end point of the operating unit subroutine (1803).

From the decision block (1812), if the steam turbine or PGRT did not reach the upper limit of operation, control passes to the operation block (1814), where the fuel flow to the PGRT increases. Now control returns to the endpoint of the subroutine of the operating unit (1803).

Output reduction subroutine

In accordance with Fig. 19, the production reduction procedure begins at step (1901) and proceeds to the decision block (1902). If the station does not operate in the transition zone (the stage where one gas turbine is in the process of commissioning or decommissioning), control passes to the decision block (1904). Note that, according to FIG. 43, the transition zone of operation is in the range from 70 to 80% of the power plant load.

If the power plant is in the transition zone of operation, the process proceeds to the transition control subroutine (1905). An example of a variant of this routine is illustrated in the flowchart of FIG. Then, control returns to the endpoint of the operating unit subroutine (1903). All process returns to this block (1903) are carried out in accordance with FIG. 17 (1707), and in the end, a transition is made to the initial operational unit for general control of the entire station (FIG. 17,1702).

From the decision block (1904), subject to the operation of all gas turbines, the process goes to the decision block (1920). At this junction, the control device determines whether the steam turbine or PGRT are operating at the lower limit of operation. For PGRT and a steam turbine, these limits are set by the engineers of the station, which determine the optimal point for starting a shutdown of the gas turbine. If none of these limits is reached, control passes to the operating unit (1921), where the supply of fuel to the HRSG decreases. Now control returns to the end point of the operating unit subroutine (1903).

From the decision block (1920), if the gas turbine or PGRT are at the lower limit of operation, control passes to the decision block (1922). If the generation of the power plant exceeds the upper limit of the transition zone of operation, control passes to the operating unit (1924), where the load on one or more gas turbines decreases. Now control returns to the end point of the operating unit subroutine (1903).

From the decision block (1922), if the station’s output reaches the upper limit of the transition zone of operation, control passes to (1923), the transition control subroutine. An example of this procedure is presented in the flowchart of FIG. Now control returns to the end point of the operating unit subroutine (1903).

From the decision-making block (1904), if all gas turbines do not take part in the operation of the station, the process proceeds to the decision-making block (1910)). At this junction, the control device determines whether the steam turbine or PGRT is operating at the lower limit. For PGRT and a steam turbine, these limits are set by the engineers of the station, which determine the optimal point for starting a shutdown of the gas turbine. If none of these limits is reached, control passes to the operation unit (1911), where the supply of fuel to the HRSG decreases. Now control returns to the end point of the operating unit subroutine (1903).

From the decision block (1910), if the gas turbine or PGRT are at the lower limit of operation, control passes to the decision block (1912). If the generation of the power plant exceeds the upper limit of the transition zone of operation, control passes to the operating unit (1914), where the load on one or more gas turbines decreases. Now control returns to the end point of the operating unit subroutine (1903).

From the decision block (1912), if the station’s output reaches the upper limit of the transition zone of operation, control passes to (1913), the transition control subroutine. An example of this procedure is presented in the flowchart of FIG. Now control returns to the end point of the operating unit subroutine (1903).

Transitional Operation

In accordance with FIG. 20, the transition control routine begins in (2001) and proceeds to the decision block (2002). If it is necessary to increase the load, operational control passes to the decision-making block (2010).

If the power plant operates at the lower limit of the transition zone, the operational control proceeds to (2011) and an additional gas turbine begins to operate. Then, control returns to the operation unit (2012). At this stage, the load of the power plant changes in accordance with the prescribed program for the production of gas turbines and steam turbines, in particular during the transition period. Now control returns to the end point of the subroutine of the operating unit (2030).

If a reduction in load is required, operational control proceeds to the decision block (2020).

If the power plant operates at the lower limit of the transition zone, operational control passes to (2021), the gas turbine is taken out of operation and shuts down. Now control returns to the end point of the subroutine of the operating unit (2030).

If the power plant is not at the lower limit of the transition zone, operational control proceeds to (2022), where the load is modulated in accordance with a predetermined production program for gas and steam turbines for the transition zone of operation. Now control returns to the end point of the subroutine of the operating unit (2030).

conclusions

The described method of controlling PGRT and a power plant demonstrated how positively the advantages of the present invention can be applied in the operation of a power plant. It should be noted that in the examples of control systems in accordance with the flowcharts shown in Fig-20, some particular steps may be added or omitted that do not affect the universal approach in accordance with the present invention or the scope of the invention.

The essence of the advantages of the present invention boils down to the fact that while a large number of control circuits can be involved in solving problems of reducing overall costs and saving the environment, the basic use of HRSG with a single pressure level (or close to a single one) in combination with additional combustion is capable of improve overall economic performance and reduce environmental protection in comparison with the production technologies of existing power plants. Moreover, the proposed new method of operating a power plant and generating maximum energy with a wide range of loads while maintaining the full load mode of a gas turbine (as opposed to the prior art) puts the proposed control technology at the highest level in the field of design of power plant control systems.

INDUSTRIAL APPLICATION

Numerous innovations of the present invention will be described with certain references to a preferred embodiment, where these innovations are particularly advantageous for solving the specific problems of power plants operating in a combination mode of cycles with a high energy density. However, it must be understood that such an option would be just an example of the beneficial use of the innovations of the present invention. In general, the proposals contained in the description of the present invention do not necessarily limit any of the various claimed patent claims. Moreover, some claims may relate to some features of the invention, but not relate to others.

Upgraded Systems

Today, many power plants operating on nuclear, coal or fuel oil are still being used. With the increasing influence of competition on the electricity market, which leads to efficient operation, environmental concerns from the point of view of the greenhouse effect and emissions of pollutants, the modernization of existing steam turbine power plants and their transformation into power plants with combined cycles becomes an increasingly suitable way out. However, traditional combined cycle power plants produce steam at three pressure levels, while existing steam turbines in traditional steam plants are designed to use only high pressure steam.

In the General Electric information brochure, “Steam Turbines for STAG ™ Systems with Combination Cycles for Power Generation,” M. Boss (GER-3582E, 1996) (GE informative document GER-3582E (1996), entitled "Steam Turbines for STAG ™ Combined Cycle Power Systems ", by M. Boss) the author describes the main difference between a steam turbine of a traditional steam power plant and a steam turbine of a traditional power plant with combined cycles:

“The mass flow in the exhausts of a power plant with combined cycles in a three-level pressure system can exceed the flow at the throttle by 30%. This is directly opposite to the conditions of many power plants with fuel boilers, where the exhaust flow is 25-30% less than the flow at the throttle, "steam extraction from a heat turbine during multi-stage heating of feed water."

This phenomenon significantly complicates the modernization of traditional steam power plants with their transformation into traditional power plants with combined cycles based on the prior art. Since conventional steam power plants allow the input steam only at high pressure, they are not suitable for the supply of intermediate and low pressure steam generated PGRT traditional system with combined cycles. In order to be effective, as already discussed, conventional combined cycle power plants known in the art have a steam / gas turbine energy ratio of about 0.5: 1. For this reason, to modernize a traditional 400 MW steam power plant with transformation into a traditional power plant with combined cycles and increase the total power of the power plant to 1,200 MW, a 800 MW gas turbine would be required. The existing infrastructure, fuel lines, real estate (area) and, most importantly, high-voltage lines may not have the appropriate size or characteristics for such an increase in operating values (an increase of 200%).

In addition, to obtain a high level of efficiency in a system combining prior art cycles, it would be necessary to modify the steam turbine so that it can receive intermediate and low pressure steam, as well as modify the entire steam path (internal components, including rotating or stationary blades) since the ratio of the exhaust steam to steam at the throttle would change from 0.75 for a traditional steam power plant to 1.30 for a traditional power plant with a combination of cycles, and this change is 1.73 times (1.3 / 0.75). This very significant change in the steam path of a steam turbine is an expensive procedure and may entail some limitations since the existing turbine housing may not be suitable for redesign. Complicating the situation is the fact that most of the equipment available at a real steam power plant (capacitors, pumps, pipes, etc.) is unsuitable for the traditional system design with combined cycles. And objects such as feed water heaters are not used at all in the combined cycle of the prior art.

Numerous variations of the present invention, however, are an ideal solution to the problem of upgrading traditional steam power plants to transform them into combined cycle power plants. Since in some cases, in accordance with the present invention, the generation of mainly high pressure steam is set, the present invention is ideal for such an upgrade. Modern cycle combining technology produces steam at pressures up to 1800 psia, and a common utility standard for steam power plants is 2400 psia, which is just the preferred inlet in some preferred embodiments of the present invention. In addition, since the present invention allows the use of a higher energy ratio of steam / gas turbines (for example, approximately 1.2: 1.0), only 330 MW of gas turbine power is needed to modify a traditional 400 MW steam power plant into an efficient combined cycle power plant as described in some embodiments of the present invention. Also, a lot of equipment of a traditional steam power plant, including a steam turbine, feed water heaters, condensers, pumps and other auxiliary equipment items, can be used practically without any changes.

Comparison of modernization - an embodiment of the invention and the prior art

U.S. Patent Nos. 5,375,410 and 5,442,908 to Briesch and Costanzo suggest a hybrid-style power plant suitable for use in retrofit applications, but it also proposes the use of three-level HRSGs. However, they do not provide for the use of additional combustion, as well as the cooling of exhaust gases of the PGRT with feed water. The power plants modernized in this way operate as traditional ones with combined cycles in case of non-use of boiler fuel. For comparison, the present invention uses boiler fuel and / or additional combustion of HRSG, providing the best ratio between fuels, saving fuel costs, partial load requirements and / or levels of emissions of harmful substances.

An example to compare the results of the modernization of existing steam power plants is illustrated in Fig. 44. In this example, a steam power plant is being upgraded, with rated design conditions of 2400 psia inlet pressure (psi, gauge) at a single reheat temperature and an inlet / reheat temperature of 1050 ° F. Such steam characteristics are typically associated with the use of fossil fuels, such as coal or oil. Although the steam turbine of the power plant is in good condition, difficulties may arise due to environmental standards, expensive boiler repairs, and economic factors in the electricity market. Any of these factors or their combination is an incentive for owners of a power plant to consider the possibility of re-equipping an existing power plant by transforming it into a more environmentally friendly and efficient power plant with combined cycles.

A 400 MW traditional steam power plant has a heating capacity of 7620 BTU / kWh. If the fuel is expensive, it will be cost-effective to upgrade it to a power plant with combined cycles. However, a power plant may (partly due to reduced heat output) have a low estimated cost. In this example, it is assumed that the power plant is valued at $ 50 million, which corresponds to only $ 125 per 1 kW. At a low cost of fuel, modernization can be economically unjustified.

When planning the economic indicators of modernization, it is necessary to choose the best combinations of equipment that will maximize the efficiency and power of the power plant. The design of a large steam turbine, as in this example, its design will be similar to that shown in Fig. 51. As you can see from this illustration, the rotating and stationary blades in the high / intermediate pressure housing on the left side of the image are much smaller in size than those intended for low pressure housings on the right side. Although the set of blades in the low-pressure housing can be changed, this will require a change in the housing itself, which will affect the foundation supporting the design and capacitors. The foundation, supporting structure and capacitors related to low pressure housings are large, heavy components that are time consuming and expensive to replace. For this reason, it is desirable to use low-pressure casings of a steam turbine with small or no modifications at all and to make changes in the steam path only in high / intermediate pressure sections.

To maximize the low-pressure section of an existing steam turbine, it is advisable to bring the exhaust level of the new power plant with combined cycles to the level of the former steam power plant, approximately in the amount of 1,587,000 Ib / hr. Using the General Electric PG 7241 (FA) model as the gas turbine engine in this embodiment and the three-level PGRT, the total steam production in this gas turbine will be only 528,000 Ib / hr. For this reason, for such an upgrade in the prior art, three gas turbines of the indicated model would be needed. Such a power plant with combined cycles of the prior art will have a capacity of approximately 800 MW and a thermal capacity of 6040 BTU / kW · h after modernization. However, having significantly reduced the flow in the high / intermediate pressure sections of an existing steam turbine, it is necessary to change the set of blades in these sections. Due to the low volume flow rate, the inlet pressure of the steam turbine will be reduced to 1800 psia. The capacity of the upgraded steam turbine will be 300 MW. It should be noted that since the combined cycle of the prior art does not use feed water heaters, the latter will not participate in the work. In general, the modernization of the plant, including changes in the high / intermediate pressure section of the steam turbine, is extensive and expensive, and $ 10 million has been allocated to carry out such modernization of the steam turbine.

In accordance with the present invention, there are at least two modernization options for transformability. The first option uses only one standard industrial model of a General Electric gas turbine and PGRT. This option requires a high intensity of additional combustion, but also provides the production of large quantities of steam. With the same exhaust stream as a conventional steam power plant, the flows into the inlet of a steam turbine account for approximately 93% of the level of a traditional steam power plant design. Therefore, this steam turbine can be used without change, with a decrease in inlet pressure of only 7% in standard mode. In addition, this project uses existing feed water heaters. The capacity of the modified steam turbine will be approximately 375 MW at the thermal power level of the entire power plant with a combined cycle of 6235 BTU / kWh.

The second option provides for two standard industrial gas turbines of the General Electric model PG7241 (FA) and heat recovery steam generators. In this embodiment, the steam turbine exhaust stream exceeds the design level by 15%. Approximately the same ratio will increase exhaust pressure, and the overall level of efficiency will be less. In this embodiment, the inlet steam flows make up 87% of the design value of the steam power plant, so the steam turbine can be used unchanged, but with reduced inlet pressure. Such a project would also be able to use existing feedwater heaters. The capacity of the modified steam turbine for the second option will be approximately 395 MW, and the thermal capacity of the entire power plant with a combination of 6060 BTU / kWh cycles.

The table in FIG. 44 shows data for various upgrade options. As described earlier, the ultimate determining factor is economic valuation. If the cost of fuel or the utilization rate of the power plant is small, the upgrade may not be justified. An increase in the cost of fuel can dictate the conditions for increasing the efficiency of a power plant, but at a reasonable cost level. Restrictions on fuel supplies, power lines or floor space can create restrictions on energy production or the number of pieces of equipment. Ultimately, the implementation in accordance with the present invention offers a greater number of options, better operating conditions for existing steam turbines, smaller changes in infrastructure and lower costs compared to the costs of upgrading a power plant with combined prior art cycles.

Combined Cycle Power Plants

The present invention is particularly suitable for use in combined cycle power plants where there is currently a tendency to use gas combined cycle turbo systems. The characteristics of the present invention are attractive for such systems, due to the reduction in the number of pieces of equipment, the reduction of floor space and capital costs. For example, when using the present invention, it is advisable to design power plants with a high energy density with combined cycles if the initial capital cost is 25% lower than that of the same prior art system.

Thus, a conventional station with combined cycles costing $ 340 million (see FIG. 25) using the methods of the present invention can be built for $ 240 million in capital costs (see FIG. 27). Primary savings of $ 100 million. This savings is equivalent to saving 10 million dollars a year if the interest rate is 8% and depreciation is 20 years. Assuming that the cost of fuel for a prior art power plant with a capacity of 725 MW is $ 93.4 million per year (see FIG. 25), an annual savings of $ 10 million corresponds to 10.7% of the total annual fuel costs for such a power plant. This means that, in terms of the amount of fuel used, the present invention may be less efficient (10.7%) compared to modern systems with combined cycles and at the same time more cost-effective. Obviously, the aim of the present invention is to achieve fuel efficiency and environmental efficiency. Thus, significant savings can be achieved over the entire life of the station.

In many new designs of power plants or in situations where the power plant is modernized or improved on the basis of the existing one, the amount of space for the construction of a new station is fixed. Thus, the possibility, when using the present invention, to provide an equal amount of generated energy with less cost per area for a power plant makes the invention attractive, especially when the overall efficiency of the plant can be maintained at the same or higher level.

In addition, the provision of the ability to operate efficiently over a wide range of partial loads is a great advantage of the present invention compared to the prior art, both from the point of view of fuel efficiency and from the point of view of harmful emissions. In conclusion, it must be said that the invention allows to reduce the total heat emission of a high-power plant, which is a very attractive point in terms of reducing the negative impact of the consumed heat on the environment, especially taking into account recent studies of global warming, etc.

Energy transfer fluids

As is known to those skilled in the art, it is preferable for energy transportation to use liquids consisting mainly of air in the upper steam cycle and / or hot water in the lower cycle. At the same time, the present invention allows the use of a wide range of other types of liquids, including ammonia, chlorinated fluorinated hydrocarbon, petroleum products, etc.

These are just some examples of the types of liquids for energy transport, and in the context of the present invention, with any mention of liquids for energy transport, this concept in accordance with the present invention should be considered as the broadest for appropriate use.

Combustible fuel and / or fuel / heat sources

As is known to those skilled in the art, in preferred embodiments, combustible fuels are used, consisting mainly of natural gas, and the present invention allows the use of a wide range of other liquid fuels, including hydrocarbon based fossil fuels, heating oil, diesel fuel and kerosene. Of course, different types of fuel can be used separately or in combination, as in systems using mixed types of fuel, which are also included in the scope of the present invention. In the context of the present invention, only some types of combustible fuel are mentioned, but any mention of combustible fuel must take into account the widest application of this concept for the corresponding application in accordance with the present invention.

Similarly, any reference to the term “fuel / heat source”, even a specific reference to heat from burning natural gas, may also include heat generated from burning any fuel as defined above, but may also include all or part of the heat generated by a geothermal source, atomic reactor as a result of fission of the atomic nucleus, indirect combustion and / or other energy sources.

Gas turbine engine

After the refusal to regulate the electricity market, there was a surge in activity among developers trying to be the first to bring new capacity to the market. The developers strategy was that after the construction of a sufficient number of power plants in a particular region, banks and other financial institutions would be reluctant to respond to proposals for financing the construction of additional power plants in this region. Therefore, the general point of view is such that the one who builds the station first wins the economic race.

This market situation has affected gas turbine manufacturers. At present (the second quarter of 1999) there is a queue of approximately three years for gas turbines manufactured by General Electric. In recent years, the wait for an order for one such gas turbine has been less than 10 months. This is also emphasized in POWER MAGAZINE (ISSN 0032-5929. March / April 1999, h / 13):

“Gas turbines that have been selling at a small profit in the last few years are suddenly in great demand, like stocks with a dot-com address, when utilities and independent energy producers are in a hurry to absorb capacity across North America. Some companies order dozens of turbines thus closing the production capacities of the main producers for many years to come. "

Such an increase in demand for gas turbines not only significantly increased their selling price, but made it difficult to acquire some models of gas turbines without waiting in line for 2-4 years. Therefore, in the framework of the present invention, an attempt has been made to circumvent this problem by producing large amounts of energy in a steam turbine. This reduces the need for high power gas turbines, and in some embodiments of the present invention can be achieved double power through the use of gas turbines, which were used in the combined cycle of the prior art.

Westinghouse Gas Turbine Engine Model 501G

The 501G model gas turbine engine is the next step in Class "F" technology. Class “G” technology engines are characterized by high pressure ratios, more advanced materials from which the blades are made, and also a combustion temperature of up to 2600 ° F. In order to avoid serious thermal deformations or other damage due to high temperature, it is necessary to supply steam to the gas turbine for cooling in the combustion chamber / turbine section of these gas turbines. Thus, in this new technology, the proper functioning of a gas turbine depends on the steam cycle. This option of organizing equipment provides higher efficiency of the combined cycle at full load, however, this technology has some disadvantages. Some of them are described below:

1. The technology has not yet been tested.

2. The cycle is not very adaptable, since additional combustion is limited to less than ten percent of the increase in energy. In addition, additional combustion reduces the overall efficiency of the power plant.

3. At higher combustion temperatures, NOX forms faster and the expected level of NOX is 42 PPM for natural gas versus 9 PPM for a gas turbine of this model.

4. If there is an integrated steam cooling system in the combustion section of the 501G model, there is a need for ultra-pure steam. Since the passages of the cooling steam in the components of the gas turbine are small in size, deposits and growths that can form with unclean steam are not allowed. Therefore, for the production of high-purity steam, special condensate fine purification systems are needed.

5. A study of the heat balance for the 501G model shows that some of this chilled steam is consumed by a gas turbine (possibly when passing through a turbine section). For a power plant with combined cycles of the 2X 501G model, this part of the steam is from 35,000 to 45,000 Ib / hr. This increases the requirements for filling water, increases the load on the condensate fine purification systems, and in addition, this consumption can increase over time, since the small ducts from where the leak occurs increase in size as thermal deformations, erosion, or other factors develop, which detrimental to effectiveness.

6. Most combined cycles, to improve efficiency, operate with variable pressure in the steam cycle. However, the cooled steam flowing out of the intermediate pressure boiler in the HRSG must be maintained at an almost constant pressure for adequate cooling. This reduces productivity under partial load conditions compared to even traditional combined cycle power plants.

7. Given the higher pressure ratio during operation of the 501G model, a fuel gas pressure level of from 600 to 650 psia is required, while for the FA model of General Electric, from 350 to 370 psia. Pipeline production generally does not meet these requirements, therefore, fuel gas compressors will be required in cases where Class F engines do not need them.

8. Such gas turbines require more than three hours to reach full load, which limits their use during peak periods, while Class "F" engines reach full load in just thirty minutes.

Thus, it is obvious that more technologically advanced engines (including the proposed General Electric class “H” engines) have many possibilities for increasing system performance by combining cycles by several percent, however, to achieve this relatively small increase, they require the use of a limiting, expensive and sophisticated technology. In accordance with the present invention, some promising engines can be used, for example, the 501G model (it would only require some changes for cooling steam), however, the use of PFA models and other industrial systems manufactured by General Electric is mainly considered, due to the experience of their use, simplicity low emissions and improved efficiency.

Comparison of combined cycles: Gas turbine technology “G” / “H” and technology “F”

In light of the liberalization of the electricity production market and, consequently, the emergence of competitive conditions caused by such liberalization, the electricity industry now prefers more advanced and complex means of energy production. Especially among gas turbine-based systems, gas turbine technologies of class “G” and “H” have become preferred.

However, the use of such technology is not without drawbacks, both from an economic and environmental point of view. In particular, gas turbine technologies “G” and “H” do not provide the necessary flexibility in operation compared to their predecessor - technology “F”. These new technologies require the mandatory integration of steam and gas turbine cycles, as new gas turbines require the cooling of internal components with steam. Without this requirement, gas turbines will not work. In this case, the steam cycle will not be able to respond to load changes as possible even in the prior art, since the requirements of the cooling steam will dictate the conditions of some part of the produced steam.

To control the load requirements, new technologies, as in the existing level of technology, are focused on changing the functioning of a gas turbine. However, due to the essence of integrable cycles, these technologies allow the use of additional combustion in a small volume or do not allow its use at all. This circumstance, together with the requirements for the station under partial load conditions, leads to a significant reduction in thermal power at partial load even in comparison with older stations using the “F” technology, where there is a direct connection between the cooling of a gas turbine and the operation of a steam turbine is absent. Thus, newer gas turbines are usually suitable for use in power plants operating at standard load. This contradicts the conditions of work with a load that varies according to demand during the day and from season to season, typical of many new power plants.

In addition, stations using new technologies have higher combustion temperatures, which leads to the need to use more exotic materials in their construction. Higher temperatures thus lead to significantly higher maintenance costs, as well as higher NOX emissions.

In addition, new gas turbines use higher degrees of pressure increase in the engine compressor to achieve higher efficiency. As a result, there is a need for a higher inlet pressure of natural gas, which in many situations requires the introduction of fuel gas compressors. Fuel gas compressors consume a lot of energy, which leads to lower efficiency, higher costs and lower reliability of power plants with combined cycles.

In light of the limitations of operational flexibility, partial load performance, increased NOX levels, requirements for possible compression of fuel gas, and while the designs in the G and H technologies were not tested even during short-term operation, the present invention draws attention to the use of older gas turbine technologies, such as class “F” technology. Moreover, the present invention allows to separate the gas turbine and steam turbine cycles while ensuring the possibility of the gas turbine at maximum fuel and emissions efficiency. The present invention, using the “F” technology, provides for the creation of a power plant that has significantly improved partial load performance, flexibility, reduced emissions and the total cost of the installation.

With a heat level of 6006 BTU / kW · h in accordance with the present invention, versus 5830 BTU / kW · h for Westinghouse “G” technology and 5690 BTU / kW · h for General Electric technology “H”, an increase in efficiency with the standard load for these more advanced (nevertheless limited from an operational point of view) power plants with combined cycles from the current level of technology will be only 3 and 5.5%. Given the decrease in productivity at part load, the additional maintenance costs, the increase in capital costs and the lack of flexibility in operation, it is not very likely that the G and H technologies (even with their significantly higher performance when working at full load) will provide such economic advantages as are achievable with the implementation of the present invention as applied to combined cycle power plants.

Although the conclusions of the present invention are based on the use of the “F” gas turbine technology, they also comply with the “G” and “H” technologies, but subject to a skillful approach to the production of gas turbines. It should be noted, however, that the present invention is not limited to the involvement of a particular production, a specific model of a gas turbine, but is applicable to the entire range of known industrial gas turbines.

Preferred Power Plant Development Model

Since the preferred option is a more flexible system with a combination of cycles, it provides high efficiency (at full and partial load) and significant cost advantages associated with a higher energy density in accordance with this project. A method for selecting a power plant that is optimal in terms of operation and financing is described below.

The choice

According to the flowchart of FIG. 47, the process starts at the start block (4701) and proceeds to the decision block (4702), where it is decided whether the construction of a new power plant or the re-equipment of an existing one will be considered. If the station is rebuilt, operational control passes to the decision block (4704). If the power plant is a mixed fuel project, operational control proceeds to the mixed fuel project subprogram (4705). Otherwise, the operational control proceeds to block (4706), where the station designer, using information in (4707) and other information about the location for the proposed station, for example, transmission line capacity, area, and commercial cost of electricity, selects the desired power station capacity combined cycle (CCR).

Knowing the performance of a combined cycle power plant, the developer proceeds to (4708) and, comparing the data obtained from (4709), selects a gas turbine for the preferred embodiment of the combined cycle from the list of samples, such as, for example, in Fig. 29 (note - Fig. 29 provides only a partial list for demonstration purposes). After selecting gas turbines, their total power (GTP) can be determined. Turning to (4710), steam turbine energy (STP) can be calculated as CCR -GTP.

Based on the energy of the steam turbines and the total capacity of the gas turbines, the operational control proceeds to (4711), where the ratio of the energy of the steam / gas turbines (STP / GTP) is calculated. Operational control now proceeds to (4712), where the desired steam efficiency and characteristics are determined based on a typical curve similar to that shown in FIG. 30. Operational control now moves to 4801 to conduct an economic assessment of the selected combined cycle.

According to Fig. 48, an economic assessment starts at (4801) and goes to block (4802), which compares information about the load, fuel types, fuel cost, and other factors listed in (4803) to determine fuel costs and average annual costs for a specific type of fuel (US $ / kWh).

The process then proceeds to (4804), which compares information on equipment costs, equipment installation, financing, and other factors listed in (4807) to determine capital costs and average annual capital costs (US $ / kWh).

The process now moves on to (4806), which compares information on inventory costs, maintenance, tools, and other factors listed in (4807) to determine maintenance costs and average annual maintenance costs (US $ / kWh).

The process goes to (4808), which compares information on employee wage costs, taxes, insurance, and other factors listed in (4809) to determine the remaining costs and their average annual cost (US $ / kWh).

Based on the data on fuel, capital costs, maintenance and data on the costs of other needs, together with the factors listed in (4811), a complete economic model for the proposed plant with combined cycles in accordance with the present invention is determined.

The process proceeds to decision block (4812) to determine if the selected option is acceptable. If so, the process proceeds to (4813), where this option is compared with other acceptable options. Operational control is sent to the identification unit (4814). If the calculated option proves to be advantageous over other acceptable options, it becomes preferable and is saved as such (4815). Operational control continues to move toward the decision block (4816). If the new option is not preferred, operational control passes to the decision block (4816), bypassing (4815).

From the decision block (4816), if new options are necessary, operational control returns to the design / financing process (4701) in FIG. 47. Otherwise, the process proceeds to (4817), where the preferred option is selected as the business plan for the combined cycle project and the process ends in (4818).

Upgraded Power Plants

According to Fig. 49, the process of re-equipping the station begins at (4901) and proceeds to the decision block (4902), where it is determined whether the station is being upgraded to operate on mixed fuel or not. If the station is designed for hybrid fuels, the process moves on to the mixed fuel design routine (4903). After returning from this subroutine, the process proceeds to (4904) to determine economic issues (see Fig. 48).

If the station is not designed to operate on mixed fuel, control passes to the decision block (4905). At this junction, it is necessary to determine whether the existing steam turbine will be modified or used “as is”. If changes are planned, the process proceeds to (4906), where the new parameters of the steam turbine are determined in accordance with the data of block (4907). From this point, the process proceeds to (4908). From the decision block (4905), if the steam turbine will be used “as is”, the operational control proceeds to (4908).

Comparing the data (4909), the parameters of the proposed combined cycle are determined, and the process goes to (4910). If there is data on fuel, capital and other factors of importance listed in (4911), the ratio of the energy of steam / gas turbines is chosen. When moving to block (4912), using data similar to the data of FIG. 29, gas turbines can be selected. Now the process goes to (4801), where economic issues are determined (see Fig. 48).

Mixed Fuel Power Plants

Mixed fuel power plants can use several types of combustible fuel to generate energy, as well as atomic, geothermal or other heat sources. By combining the combined cycle described in the preferred embodiment of the present invention with a mixed fuel cycle, an increase in overall performance and economic performance is possible.

In accordance with the flowchart of FIG. 50, the mixed fuel design routine starts at block (5001). Control passes to the decision block (5002), where it is decided whether combustible fuel or a heat source such as nuclear or geothermal will be used. If combustible fuel is to be used, the process proceeds to (5005), where gas turbines are selected to operate the mixed fuel power plant - in accordance with the cost of the fuel, the size of the steam turbine, the desired station parameters and other factors, as indicated in (5006). From this point, the routine returns to the call point.

From the decision block (5002), if an atomic or geothermal heat source is used, the process proceeds to (5003), where gas turbines are selected for the mixed fuel station in accordance with the cost of the fuel, the size of the steam turbine, the desired station parameters and other factors as indicated in (5004). From this point, the routine returns to the call point.

OPTIONS

General Provisions

As noted above, one of the most important advantages of the present invention is its operational flexibility. This is evident not only from the possibilities of choosing the operating parameters of a power plant with combined cycles, but also from the ability to implement other solutions, such as upgrading existing steam power plants or arranging mixed fuel cycles. The following is a description of other options that can be successfully used in the implementation of the present invention.

Equipment layout

In US Pat. No. 5,649,416, James H. Moore describes various equipment arrangements that include gas and steam turbines connected and driven by a common generator. Although the embodiment of FIG. 26 shows gas turbines and a steam turbine, each with its own generator, there is no reason to insist on such an option for using equipment layouts. The present invention involves the implementation of a new system and method, and the layout of the equipment may be as described by Moore in his patent, or if desired, be used in a different form. Thus, the present invention contemplates any combination of single-shaft system configurations.

Other upper / lower cycles

The present invention has been discussed mainly from the point of view of using traditional Brighton / Renkine cycles for combining cycles. However, it should be noted that the advantages of the present invention are also applicable to other cycles. There are no real restrictions on the use of other cycles in the present invention, however, it is contemplated that, in particular, the Kalina General Electric cycle (lower cycle) may be particularly suitable for use in implementing the present invention.

Thus, in this document, the terms “lower” and “upper” cycles should be interpreted in the broadest sense with respect to the possibility of using the Breiton, Renkine, Kalina, etc. cycles known to specialists. In addition, it should be noted that in this system with a combination of cycles in accordance with the present invention, it is especially intended to use multiple repeating cycles.

Boiler feed pumps powered by small steam turbines

For demonstration purposes, it was assumed that the boiler feed pumps referred to in this document are powered by electric motors. However, in more powerful steam power plants, these pumps often run on small steam turbines called boiler pump turbines. The turbines of the boiler pumps have several advantages over engines, the most important of which are the response to the load and the reduction of the load on the blades at the exhaust outlet.

Since boiler pump turbines use low pressure inlet steam (usually less than 200 psia), they usually consume a sufficient amount of steam. This flow of steam used by the boiler pump turbines is equal to the decrease in steam in the low pressure section of the main steam turbine. This reduces the load on the vanes of the last cascade and often leads to an increase in cycle productivity.

Advanced Steam Features

U.S. Patent 5,628,183 Rice examines higher temperature and pressure studies in Europe as well as in the United States with the assistance of the Department of Energy (DOE) and the Electric Power Research Institute (EPRI) )). Research has included the operation of solar turbines as part of an installation project designed for increased cycle efficiency using an inlet steam temperature of 1,500 ° F in a steam turbine. Although long-term observation has not been carried out since the reliability of the elevated levels of pressure and / or steam temperatures is confirmed, this technology can be easily implemented based on the present invention.

Progress in gas turbine technology

The development of gas turbine technology is focused on the use of more efficient compressors, new metallurgy, higher combustion temperatures, higher pressure ratios, as well as other ways of influencing the increase in productivity. As these advances develop and become available, gas turbines will be suitable for introduction into the duty cycle in accordance with the present invention.

PGRT non-corrosive low pressure section

The problem of the harmful effects of the condensate of the exhaust gases of gas turbines and its ability to cause corrosion of pipes and ribs in the low-pressure section of the PGRT has already been considered. A generally accessible way to avoid this problem is to supply feed gas to the HRSG preheated to a temperature well above the dew point of the exhaust gas turbine, which prevents the formation of moisture on the heat exchange surfaces in the HRSG. This method has been demonstrated in the description of embodiments of the invention in accordance with Figs. 35 and 39.

Another suitable way to avoid this problem is the use of non-corrosive material for the manufacture of pipes and ribs in PGRT, usually stainless steel. In this case, the need for preheating of the supply water is eliminated and the possibility of continuing the process of cooling the exhaust gases of a gas turbine is ensured, which makes it possible to realize the energy recovery of the above gases even at a higher level. Undesirable, however, is the additional cost of stainless steel material. In some cases, this cost may exceed the amount of energy saved. But if fuel prices are high and the cost of the material is relatively small, this option can be cost-effective.

Combined high / low pressure pump

To create the required pressure level in the steam cycle, a pump is usually used to pump feed water to the proper pressure. In some embodiments, including the embodiments of FIGS. 9, 15, 35, and 39, dual pumps are indicated to provide low and high pressures. It can be pumps for multiple pumping, as shown in the drawings, or one pump. In all cases, they consist of several impellers that sequentially pump feed water. A pump compartment with openings at the points of proper pressure can provide an intermediate pressure for the supply water, while the remainder of the water continues to flow towards the high pressure outlet. Other options for organizing pumps are also possible. The preferred embodiment of the present invention does not in any way limit the size or model of the pump, but allows any pump or combination of pumps that provide appropriate maintenance to be used.

Waste Heat Recovery

When discussing power plants with combined cycles of the prior art and the characterization of the present invention, losses due to the inefficiency of the equipment in the system as a whole were mentioned. For example, in the form of generator losses due to non-ideal (not zero) resistance in its winding. In general, most of the losses of a system in any power plant with combined cycles can be calculated by waste heat, that is, heat generated but not converted into mechanical or electrical energy. Losses in the generator, boiler feed pump, lubricating oils, losses due to heat radiation from the gas and steam turbines into the environment are just some of the types of heat losses in a traditional system with combined cycles. In traditional systems with combined cycles (known from the prior art), it is assumed that such sources of waste heat are usually present, but not compensated, since in such configurations of a power plant, the cost of heat recovery is unprofitable, and any incentives for using low-energy waste heat in a useful form are absent.

Due to the excessive amount of low-energy heat in the exhaust gases of a gas turbine, in the prior art, multilevel pressure PGRT is used to ensure maximum heat recovery. Despite the use of constantly maintained additional combustion, the energy level in the high temperature section of the HRSG is equal to or higher than the energy level in the lower temperature section, when setting the task of replenishing feed water heaters by taking steam from a steam turbine, conventional devices operating according to the Rankine cycle, traditional systems with combined cycles of the prior art are not used.

If the demand for low energy heat is increased in accordance with the present invention, other heat sources may be used. According to FIG. 21, they include losses in a gas turbine (2102), losses in a steam turbine generator (2110), and other losses. In this case, low temperature heat, for example, heat from engine lubricating oil, heat loss in the generator, and air cooling of the gas turbine chamber can be used to pre-heat the feed water and can replace the steam recovered for use in low temperature feed water heaters. The use of this heat not only increases the thermal power of the power plant, but also reduces the heat removal requirements of this station.

In these circumstances, the present invention is in a sense unique since waste heat sources can be used together with feed water heaters (as shown in FIG. 15) to add heat to the water, which is then overheated inside the HRSG. Such a practical application of feed water heaters in the prior art was not possible, since PGRT was used to perform this function, and heating feed water did not imply any advantages for the system with combining the prior art cycles. Thus, with the reasonable use of heating of the feed water with additional combustion in accordance with the present invention, there is provided a method for efficiently recovering heat, which in the prior art remained waste heat not subject to regeneration.

It should be noted that the possibility of using waste heat in practice can be considered a significant improvement in terms of increasing the efficiency of the entire system with combined cycles. Consider, for example, the case in which 1-2% of the waste heat generated in the system is regenerated and sent for useful use. Considering that a combined cycle power plant with a capacity of 1000 MW will spend about $ 175 million a year on fuel, which means that an increase in the total cycle capacity of even 1% will result in significant fuel savings ($ 1.75 million in year). Over the 20 years of the power plant’s existence, total fuel economy is estimated at US $ 35 million. Thus, the recovery of waste heat using the present invention represents a new potential for improving the overall economic efficiency of power plants with combined cycles, which was practically unavailable when using the prior art.

It should be noted that the regeneration of waste heat at a power plant with a combination of cycles is a direct increase in the overall efficiency of thermal conversion, providing a direct decrease in atmospheric heating. In connection with the growing concern about global warming, the issue of regeneration of lost heat should be on the same level as developers to reduce emissions of NOX and other pollutants. Since it is estimated that more than 100,000 MW of additional capacity of the power plants will be commissioned in the next ten years, issues of waste heat deserve close attention of people interested in preserving the environment. In addition, since a significant part of the waste heat of power plants with combined cycles is released into the environment, this can seriously affect plant and animal life.

Growth in the number of geothermal stations

The present invention is largely applicable where existing or proposed power plants with a low level of productivity will be supplemented by a gas turbine (1) to supplement the production of geothermal energy in order to meet the requirements of given loads; or (2) to replace losses or decline in geothermal energy production in the case of an existing geothermal power plant. Since the number of units of equipment for geothermal installation is usually fixed, a loss of efficiency or a decline in the energy production of an existing geothermal power plant can lead to unprofitable operation. In some cases, a decline in the flow of geothermal energy can lead to the closure of the power plant, as the amount of electricity generated may fall below a critical point, after which the operation of the power plant is considered impractical.

The present invention can be successfully applied to these situations, generally in the same way as it is applied in the case of waste heat recovery of a conventional power plant with combined cycles. The only difference is that the "waste heat" used in the present invention is regenerated from a geothermal source. The result of using geothermal heat in conjunction with a gas turbine at an optimal level of combustion is the production of a stable amount of energy (regardless of the quality or stability of the geothermal energy source itself). Since the advantages of the present invention are largely associated with additional combustion in the HRSG, the geothermal energy source in this application can be used through heat exchangers to supply such additional combustion in the HRSG, thereby replacing the fuel and / or heat that is typically supplied for this purpose. Thus, a decrease in the production and / or efficiency of a geothermal energy source will only lead to an adequate increase in additional combustion from other fuel and / or heat sources. Power plant performance parameters remain constant, and may even increase for the upgraded power plants discussed in other sections of this document.

Application for heat power systems

As mentioned earlier, the present invention is applicable to heat power systems and options for the associated production of heat and energy, in which both the drive shaft and the heat are used under the same operating conditions, as, for example, in commercial public power systems and industrial power plants. In such systems, a certain amount of heat from a power plant with combined cycles can be used for spatial heating, chemical processing of raw materials, pulp processing, paper drying, combined production of heat and electric energy and / or for other industrial processes, etc.

It is expected that the invention will find particularly widespread application in all forms of heat power systems and systems for the associated production of heat and electricity. However, the above examples illustrate only the range of application of the present invention. Specialists will undoubtedly expand the range of its application without prejudice to the general thrust of the present invention.

COMPARISON OF FUNCTIONING EFFICIENCY

Choosing the optimal steam / gas turbine performance ratio

To overcome the problems of working with a partial load associated with a change in the load of the power supply system, in accordance with the present invention, a steam turbine is used as the primary engine. A steam turbine can easily reduce the load by closing the intake valves or changing the intake pressure in the engine (by changing the level of additional combustion). This has a weak effect on the efficiency at part load compared to the attenuation of the combustion temperature observed in the case of a gas turbine. In addition, a steam turbine can be designed for optimum efficiency at a design level of partial load, while a gas turbine is almost always the most efficient at full load.

Understanding the differences between gas and steam turbines determines the advantages available to a steam turbine in terms of its operational flexibility. Gas turbines consist of a compressor section that compresses the incoming air (usually under environmental conditions), from three to thirty times the level of atmospheric pressure. This air should then enter the combustion zone, where it heats up when the fuel is burned from 1600 to 2600 ° F at full load, depending on the design parameters of the gas turbine. These hot and compressed gases are then expanded through a turbine section into a gas turbine to produce energy that drives not only the compressor, but also the generator. About 2/3 of the energy generated in this section of the turbine is necessary for the operation of the air compressor, and the remaining 1/3 is used for the operation of the electric generator. Due to the complexity of development associated with the need to coordinate a compressor, a combustion system, and a turbine section to ensure operation in an integrated system mode, gas turbines are very complex machines in their structure. Manufacturers usually have various gas turbine models. However, they are designed for non-matching output or power values. Creating a model in accordance with the specifications of the customer is unlikely, and not economically viable.

Steam turbines, in turn, have very flexible design capabilities. They are calculated according to certain parameters of the pumps of the boiler of the power plant for the formation of water under pressure and the parameters of the boiler of the power plant to ensure the formation of heat to turn this compressed water into steam. For this reason, a steam turbine can be easily adapted to energy requirements at the design stage by simply changing the configuration to increase the turbine's throughput to allow for a larger steam flow. This is easily achieved through the use of large stationary or rotating blades in a steam turbine. Typically, a steam turbine designer has the option of selecting from a family of vanes in the high pressure section of a steam turbine that can grow by no more than 0.25 inches. The blades in the low pressure sections of the turbine usually have higher increment values. With this design, two different steam turbines can have power ranging from 100 to 300 MW, and yet both can be suitable for the same casing (the appearance of the two turbines is identical). The key difference is the blades inside the steam turbine and the throughput of each turbine.

In addition, by appropriate selection of the low pressure vanes, it is possible to “overload”, in terms of efficiency, the vanes of the last stage at full load. For this reason, at full load, these blades become less efficient than at partial load. Then, as the load decreases, the efficiency of the low-pressure compartment increases. Such a project is preferable for power plants, most of the time working with a partial load, but associated with the need to achieve peak load in a short period of time. It is precisely their operational flexibility coupled with a low maintenance requirement and proven reliability that explains the preference for using steam turbines as the primary engine in a combined cycle power plant.

Some preferred embodiments of the present invention are a system in which exhaust gases are included in a HRSG, as in the state of the art. However, these exhaust gases usually contain a large amount of oxygen. In particular, the oxygen content in the air usually decreases from 21% in ambient air to 12-15% in the exhaust of a conventional gas turbine at full load. As a result, a large amount of oxygen remains in the exhaust gas of the gas turbine to burn additional fuel. If enough fuel were burned, all of the steam that would be produced as low pressure steam in accordance with the prior art could be enriched to the level of high pressure steam with a corresponding modernization of the system described in the preferred embodiment of the present invention. In this case, the high-pressure steam flow will increase significantly, and the capacity of the steam turbine relative to the gas turbine or gas turbines (steam / gas turbine energy ratio) will change from a nominal value of 0.5 for the existing level of technology to a value usually exceeding 1.0 . For this reason, instead of primarily following the idea of implementing a gas turbine cycle with a steam turbine regeneration cycle, the present invention is rather a traditional steam power plant with additional gas turbine power generation and a piping system for directing exhaust gases from a gas turbine to boilers of a steam power plant for preheating air and increase boiler efficiency. In order to maximize productivity, in accordance with some embodiments of the present invention, complete integration of the cycles is required, including the use of waste heat, heating of the supply water and the introduction of a control system to optimize heat transfer.

Comparison of the prior art and embodiments of the invention

As described in detail, the technology of combining cycles in the current level of technology is rooted in the technology of small thermal power plants. In a modern power plant with combined cycles in the prior art, the exhaust gases from a gas turbine enter a HRSG, typically having two or three pressure levels. The steam from each of these pressure levels is then directed to the appropriate point on the steam turbine, corresponding to the pressure level of the PGRT section. Additional combustion is used as a means to achieve higher power, but this is resorted to intermittently to meet peak load requirements, and higher power is achieved only with a decrease in thermal efficiency. The main load control of a power plant with combined cycles in the prior art is still carried out by modulating the load of a gas turbine. PGRT single pressure can be used with a corresponding decrease in thermal efficiency. Higher pressure inlets to the steam turbine are not justified, since the low volumetric flow rates to the steam turbine block any advantages of the process achieved due to the high pressure due to a decrease in turbine productivity in the high pressure section.

Using the combined cycle technology of the prior art, the power plant is basically a power plant based on the operation of a gas turbine with the addition of HRSG to regenerate waste heat. A steam turbine is also designed taking into account the best use of such regenerated heat (which is converted into steam in HRSG under pressure of various levels). A typical steam / gas turbine power ratio for such combined cycle power plants is between 40 and 60%, and for a GE Model S207FA power plant it typically reaches approximately 0.57. If it is necessary to use high, intermediate and low pressure steam, the steam turbine has sufficiently low flows in the high pressure section and high flows in the low pressure section. This reduces the efficiency in volumetric flow of high pressure and increases the relative size and cost of the exhaust compartments. The feed water is heated in PGRT and the traditional feed water heating is not used by steam extraction of a steam turbine. Preheating of the supply water from the condenser may be used, but the purpose of this process is not to increase efficiency, but rather the need to avoid condensation of water vapor in the exhaust gases. Since water is contained in the air entering the interior and is also formed as a combustion product of hydrocarbon fuels, an increased concentration of water vapor in the exhaust gases lowers the dew point. Cold water coming directly from the condenser can cause condensation on the tubes and plates of the economizer. Such condensation showed the ability to lead to corrosion of the plates, reducing the efficiency of heat transfer, and also have a detrimental effect on the operation of HRSG. Thus, the use of a device for preheating water is possible in some applications.

In conclusion, I would like to say that the combined cycle power plant of the prior art is primarily a gas turbine power plant using a steam cycle designed to achieve a compromise between the best cycle efficiency and the optimal level of heat recovery of exhaust gases. There are opportunities such as additional combustion to increase the capacity of the power plant by a nominal amount (usually less than 25%), but such an increase becomes available to the detriment of the thermal power of the power plant. Due to the rigidity of the design parameters of the gas turbine, the level of operational flexibility in terms of power or the actual design of the power plant with a combination of prior art cycles is quite low. Essentially, prior art is a rigid power plant design based on the use of a gas turbine engine or a set of engines using PGRT and a steam turbine with a rated power of 50% of the gas turbine power. The steam turbine operates in dependent mode and follows the load of the gas turbine.

In some embodiments, in accordance with the present invention, the gas turbine exhausts the HRSG at a single pressure level (or basically a single pressure level) designed to maintain continuous additional combustion. Such additional combustion significantly increases steam production and, therefore, increases the flow of feed water in such a way that additional pressure levels in the HRSG are not required to cool the exhaust gases to the optimum temperature (approximately 180 ° F). Feed water flows that exceed optimal flows through the HRSG are routed to conventional feed water heaters by taking steam from a steam turbine to improve the efficiency of the steam cycle. Due to operational flexibility, the combined cycle power plant design described in several preferred embodiments of the present invention has a steam / gas turbine power ratio that can vary from about 0.75 to 2.25. Of course, specialists in this industry understand that other ratios are also possible within the scope of the present invention. For many options for changing the load of a power plant, gas turbines remain basically at or close to the level of the most effective load (100%), and for changing the load of a steam turbine, the intensity of additional combustion changes.

In other words, embodiments of the system in accordance with the present invention inherently correspond to a large central steam power plant, similar to steam power plants known in the art, with a boiler that has been replaced by heat recovery steam generators (PGRT) that constantly burn fuel, just like and boiler in a traditional steam power station. However, now gas turbines are introduced into the working cycle, which supply hot gases enriched with oxygen (12-15%) to the boiler (PGRT), increasing the efficiency of the cycle and providing the possibility of burning additional fuel. The feed water is heated in the low-temperature sections of the PGRT and in traditional feed water heaters due to the extraction of steam from the steam turbine. A steam turbine has a larger size and greater throughput of steam flows through the high and intermediate pressure compartments and smaller flows - through the low pressure compartment (the steam extracted to heat the supply water reduces the end flow of the exhaust), increasing the efficiency of volumetric flow rates and reducing the relative volume of exhausts. In some embodiments, the present invention is a combined cycle power plant with great operational flexibility, full and partial load performance, and less costly to manufacture, operate, and maintain.

Major hardware costs

In addition to fuel and capital costs, another cost item for combined cycle power plants is the cost of technical support, especially the maintenance of key facilities such as a steam turbine and gas turbine. The cost of technical support varies for different models of equipment, its complexity, as well as the level of service (high or low temperature, continuous cycle operation, etc.). Typical maintenance costs are investigated based on the price / kilowatt-hour of energy (mills / kWh, where “mill” is $ 0.001 or 0.1 cents). The following is a list of estimated maintenance costs for some of the most important equipment items along with annual maintenance costs based on a power rating of 200 MW at a capacity of 70% (1,500,000,000 kWh per year):

Equipment (model) Maintenance Intervals (mills / kWh) Annual costs (US $) 2400 psi Steam Turbine inch (psia) 0.5 750,000 Gas turbine GE Frame & FA 2,5 3750000 Gas turbine Westinghouse 501 G 4,5 6750000

As can be seen from the table, maintenance during the operation of a gas turbine is much more expensive than a steam turbine. In addition, a gas turbine of progressive technology (Model 501G) - with an increased combustion temperature, single-crystal blades and a steam-cooled combustion section - is also an expensive object from the point of view of maintenance.

In the prior art, gas turbines produce approximately 67% of energy, and a steam turbine - 33% (steam / gas turbine power ratio 0.5: 0.1). Since gas turbines are modulated to change the load, and the steam turbine follows these changes, this ratio is fairly constant over the entire load range. For this reason, for a year of operation in a power plant with a combination of prior art cycles, the gas turbine maintenance factor would relate to 67% of the generated kilowatt hours of energy, and the steam turbine maintenance factor to 33%.

In some embodiments in accordance with the present invention, the ratios are not so simple, since when changing the entire load of a power plant, the power of a steam turbine is modulated to the highest degree possible, and the power of gas turbines is maintained at or close to full load.

Partial Load Performance Comparison

The conclusions of the present invention in the best way become apparent when comparing the operational indicators shown graphically in Fig.6, 15, 33 and 22-28.

33 illustrates the differences in partial load operation between two modern conventional combined cycle power plants and two power plant embodiments in accordance with the present invention. The graph shows that the performance of the options in accordance with the present invention is significantly higher than the same performance of conventional power plants with combined cycles in partial load mode. As can be seen from the comparative data in the tables in Figs. 25 and 27, in contrast to the typical performance profile in accordance with the present invention, less fuel is used and 100 million US dollars less is required for the construction, in addition, the NOX level for these options is less than 1/3 of the power plant’s emissions with a combined Westinghouse model cycle. Thus, the embodiment of the present invention shown in FIG. 26 leads to cost savings and a simultaneous saving on environmental pollution due to the reduced level of harmful NOX emissions. Such properties are characteristic of the implementation options in accordance with the present invention and essentially represent the best opportunities in two arrays of tasks - achieving economic efficiency while reducing environmental pollution.

For use in accordance with the present invention, PGRT, in comparison with the prior art, must have a more robust construction to withstand elevated pressures and temperatures. This is achieved in several ways. Firstly, the use of a water screen (vertical tubes filled with feed water), which may be required to enclose the PGRT fuel space to protect it from high combustion temperatures. Alternatively, the exhaust gases may first be cooled by passing through a superheater section (approximately 800 ° F), then they may be reheated to 1600 ° F before further passing through the HRSG. 1600 ° F is currently the upper temperature limit specified by manufacturers for standard PGRT design. Another alternative is to use double burner gratings in PGRT. After heating the exhaust gases of the gas turbine to 1600 ° F, they are cooled through the initial compartments of the HRSG, then, through combustion, more fuel (heat) is added to the lower point of the stream. Compared to a single burner, this leads to a practical doubling of the amount of heat without exceeding the limiting temperatures of the PGRT.

CONCLUSION

The present invention involves a large number of applications, but it should be noted that in the light of modern trends in the energy industry, its use in the field of creating power plants with combined cycles is especially advantageous. In the prior art, additional combustion in PGRT was not used as a means of increasing the productivity of the power plant as a whole, and the present invention uses this concept.

In the context of a general improvement in the efficiency of the system, the present invention puts forward the idea of using additional combustion not to produce more steam, as was the case in the known options for generating electricity, but to produce better quality steam (high mass vapor content). It follows that when spending additional fuel on additional combustion in PGRT, it is possible to produce higher energy and, thus, more highly efficient steam, if used in conjunction with the corresponding lower cycle engine.

Using the present invention to generate electricity will require the production of heat recovery steam generators (PGRT) capable of withstanding higher temperatures than current standards, while materials capable of achieving these goals are available, and prior art steam power plants and PGRT have demonstrated the ability to withstand such increased temperature levels. In addition, the data presented indicate that in many circumstances such PGRT are smaller in comparison with existing models, that is, the costs of their construction and maintenance can be comparable or even less than for existing models. In addition, the ability of PGRT data to use a single pressure level, in some cases, can lead to savings in their development and construction.

It must be emphasized that from the point of view of energy density in accordance with the present invention there are potential opportunities for a significant expansion of the number of hardware required for the implementation of the power plant. To achieve a satisfactory level of overall performance, traditional power plants use several gas and steam turbines, and when such facilities operate at partial load, the entire system can be operated at an acceptable level of efficiency. This is mainly due to the fact that the operation of a gas turbine at partial load is usually very inefficient. The present invention solves the problem of such detrimental effects by operating all gas turbines with optimum efficiency (economic and environmental), thus allowing, with fewer gas turbines to be operated, to achieve the same overall efficiency and environmental impact, while consuming less auxiliary resources.

These opportunities for improving efficiency in accordance with the present invention are complemented by a mechanism by which the power of a power plant can temporarily exceed its nominal level, albeit with a lower level of efficiency. Such an excess of power of the power plant to ensure operation under increased loads can be an important factor from the point of view of the economic side of the construction of the power plant due to obstacles, since environmental protection and logistics problems that must be overcome for the construction of new power plants priority issues to determine the feasibility of building new power plants. Moreover, the present invention allows the beneficial use of existing power plants, expanding their own capabilities, due to the dynamic adaptation of the characteristics of the power plant to meet temporary overloads. This ability can lead to significant cost savings and more environmentally friendly use of the power plant in accordance with the present invention, by reducing the additional attraction of financial costs and costs associated with environmental protection, in light of the requirements for a temporary increase in power plant output.

In conclusion, it should be emphasized that if in the past efforts to develop power plants were aimed at optimizing the functioning based on only fuel costs, power plants of the future should incorporate and optimize capital costs, costs associated with environmental impacts, the cost of real estate, legal costs regulation, as well as ever-increasing costs of technology and auxiliary equipment. The objective of the present invention is to consider all of these parameters together and to create a project for a power plant that is economical, efficient in terms of energy production and an environmentally friendly method of generating electricity.

Although a preferred embodiment of the present invention has been demonstrated using the accompanying drawings and described in detail above; it is clear that the invention is not limited to the described options, and its configuration can be changed, subject to modification, replacement of components, without departing from the essence and scope of the invention defined by the following claims.

Claims (33)

1. Power plant with a combination of cycles containing a) at least one gas turbine; b) at least one steam turbine; c) at least one heat recovery steam generator in which substantially continuous combustion of the fuel is supported; d) at least one control system adapted to maintain a substantially optimal temperature of the exhaust gases of the heat recovery steam generator, having inputs for controlling the temperature of the exhaust gases of the heat recovery steam generator, and outputs for providing operational control of the flow rate of the supply water through the heat recovery steam generator, wherein said power plant primarily functions as a Rankine cycle power generation system using one or more heat recovery steam generators that support said combustion of fuel to supply high energy steam through a lower cycle to said one or more steam turbines, and the Braith’s upper cycle is used to produce the initial energy consumed by said one or more gas turbines; and in said power plant for producing high-energy steam for operation of said steam turbines, a lower cycle with a substantially uniform pressure level is used. 2. A power plant with combined cycles according to claim 1, in which, to increase the efficiency of the cycles, heating of the feed water with the help of steam generated by a steam turbine is used. 3. The power plant with combined cycles according to claim 1, in which the high-pressure steam at the rated power of the power plant is mainly in a supercritical pressure state. 4. Power plant with a combination of cycles containing a) one or more gas turbines, b) one or more steam turbines and c) one or more steam generators of heat recovery with additional combustion of fuel, in which the coefficient of operation of the rated power as the ratio of the total production of all steam turbines to the total production of all gas turbines is more than 0.75; said additional combustion of the fuel of said one or more steam generators is substantially constant; in the named power plant for the production of high-energy steam for the operation of the named one or more steam turbines uses the lower cycle with a substantially uniform pressure level. 5. A power plant with a combination of cycles according to claim 4, in which to increase the efficiency of the cycles, heating of the feed water is used with the help of steam emitted by a steam turbine. 6. A power plant with combined cycles according to claim 4, in which steam and / or exhaust energy of a gas turbine is used to generate associated heat and / or joint production of heat and electric energy. 7. Power plant with a combination of cycles containing a) at least one gas turbine; b) at least one steam turbine; c) at least one heat recovery steam generator with additional fuel combustion, in which a boiler with a substantially uniform pressure level is used in said heat recovery steam generators; the said total input energy of additional fuel combustion at the rated power of the power plant with combined cycles is added at a level of 30% or higher than the input energy to all of the named gas turbines at their standard power in accordance with ISO; the exhaust gases on the way from the outlet of the evaporator section of the named heat recovery steam generators to the exhaust section of the named heat recovery steam generators are cooled primarily by the feed water; the exhaust steam of said steam turbines feeding feed water heaters preheats excess feed water that has not undergone pre-heating in said heat recovery steam generators; feed water passing through the heating compartments of the feed water in said heat recovery steam generators is controlled for optimal heat recovery; and the vapor pressure inside said steam turbines is increased in proportion to the increase in volumetric capacity of said steam turbine. 8. The power plant with a combination of cycles according to claim 7, in which to optimize the efficiency of the steam cycle of the heat conversion, single or multiple reheating, elevated steam temperatures, supercritical, ultracritical and / or increased parameters of the state of steam are used. 9. Power plant with a combination of cycles containing a) at least one gas turbine; b) at least one steam turbine; c) at least one steam generator for heat recovery with additional fuel combustion; d) at least one control device for controlling the temperature of the exhaust gases of the heat recovery steam generator, having inputs to which information about the temperature of the exhaust gases of the heat recovery steam generator is received, and outputs of which at least one is designed to perform one of the group tasks, including the operational control of the intensity of fuel combustion in the heat recovery steam generator and the operational management of the flow of feed water through the heat recovery steam generator, in which the said additional combustion of fuel in the said heat recovery steam generators is of a constant nature and the named additional combustion of fuel in the said heat recovery steam generators increases the inlet temperatures of the steam turbines to the values allowed under the nominal operation mode. 10. The power plant with a combination of cycles according to claim 9, in which to increase the productivity of the cycle is used, heating the feed water using steam of a steam turbine. 11. A power plant with combined cycles according to claim 9, in which said additional combustion of fuel and / or steam generated by a steam turbine is partially replaced by energy from the losses of the generator and / or other auxiliary devices. 12. Power plant with a combination of cycles according to claim 9, in which the high-pressure steam is mainly in a supercritical state at the rated output of the power plant. 13. The combined cycle power plant of claim 9, wherein steam and / or energy from the exhaust gas turbine is used to generate associated heat and / or co-production of heat and electric energy. 14. Power plant with a combination of cycles containing a) at least one fluid medium for energy transfer of the upper cycle; b) at least one upper cycle engine, including one or more fluid inlets for transferring upper cycle energy, fuel and / or heat inlets, one or more fluid outlets for transferring upper cycle energy, and one or multiple outputs of the driven load; c) at least one device for heat recovery, essentially maintaining fuel combustion continuously, having one or more inlets of a fluid outlet for transferring energy of the upper cycle, one or more inlets for supplying fuel and / or heat, one or more fluid inlet for transferring energy of the lower cycle, one or more fluid outlet for transferring energy of the lower cycle at one of the pressure levels selected from the group including low, intermediate and high pressure, as well as one or several ko output to reset the fluid for transfer of the energy of the upper loop; d) at least one lower cycle fluid; e) at least one lower cycle engine having one or more low and / or intermediate and / or high pressure inlets for receiving fluid for transferring energy of the lower cycle, one or more outlets of fluid for transferring energy of the lower cycle, and one or more outputs of the driven load; e) at least one heat exchanger having one or more inlets that receive fluid for transferring energy of the lower cycle, and one or more exits for fluid for transferring energy of the lower cycle; g) at least one fluid transport device having one or more inlets that receive fluid for transferring lower cycle energy and one or more fluid outlets for transferring lower cycle energy; and h) at least one control device that maintains a substantially optimal temperature of the exhaust gases of the heat recovery device, having inputs for data on the temperature of the exhaust gases of the heat recovery device and outputs for providing operational control of the fuel and / or heat supply, wherein said upper cycle energy transfer fluid feeds said one or more inlets of the upper cycle energy transfer fluid of the upper cycle engine; combustible fuel and / or heat supplies said one or more inputs of combustible fuel and / or heat of the upper cycle engine; said one or more exits of the upper cycle engine power said one or more exits of the exhaust gases of the heat recovery device; one or more sources of fuel and / or heat supply said one or more inputs of fuel and / or heat of the heat recovery device; said one or more fluid outlets of a low and / or intermediate and / or high pressure heat recovery device feed said fluid inlets of a low and / or intermediate and / or high pressure lower cycle engine; the named one or more outputs of the discharge of fluid into the environment of the device for heat recovery is vented to the atmosphere (directly or indirectly) or by means of a device for transporting the fluid can be directed to the named inputs of the upper cycle engine; the named one or more fluid outlets for transferring energy of the lower cycle of the lower cycle engine are sources to the named one or more inputs to the heat exchanger; named one or more exits of the heat exchanger connected to the named one or more inputs of the device for transporting fluid; the named one or more outputs of the device for transporting a fluid connected to the named one or more inputs of the device for heat recovery; said lower-cycle engine productivity is improved by adjusting the intensity of additional fuel combustion or adding said fuel and / or heat so that the fluid at low and / or intermediate and / or high pressure has a higher energy state than without adding such fuels and / or heat. 15. The combined cycle power plant of claim 14, wherein said top cycle engine comprises one or more gas turbines. 16. A power plant with combined cycles according to claim 14, in which a single upper cycle is used, and the said lower cycle engine contains one or more steam turbines. 17. The combined cycle power plant of claim 14, wherein said high pressure fluid contains predominantly a single high pressure steam. 18. The combined cycle power plant of claim 14, wherein said high pressure fluid contains predominantly high pressure steam in a supercritical state at a rated load of the power plant. 19. The combined cycle power plant of claim 14, wherein said heat recovery apparatus comprises predominantly a heat recovery steam generator with a single pressure level. 20. The combined cycle power plant of claim 14, wherein said upper cycle engine operates substantially at a peak in productivity and / or power generation level, and said lower cycle engine is used to modulate the total amount of energy generated by the combined cycle power station. 21. Power plant with a combination of cycles containing a) at least one fluid medium for energy transfer of the upper cycle; b) at least one upper cycle engine containing one or more fluid inlets for transferring energy of the upper cycle, one or more inlets for supplying fuel and / or heat, one or more outlets for discharging fluid for transferring energy of the upper cycle, and one or more outputs of the driven load; c) at least one device for heat recovery with constant maintenance of fuel combustion, operating mainly at a single pressure level and having one or more inlets of a fluid outlet for transferring energy of the upper cycle, one or more inlets for supplying fuel and / or heat, one or more fluid inlets for transferring energy of the lower cycle, one or more fluid outlets for transferring energy of the lower cycle, as well as one or more fluid outlets for transferring energy of the upper its cycle; d) at least one fluid medium for transferring energy of the lower cycle; e) at least one lower cycle engine having one or more inlets for receiving fluid for transferring energy of the lower cycle, one or more outlets of fluid for transferring energy of the lower cycle, and one or more outputs of the driven load; e) at least one heat exchanger having one or more inlets of a fluid for transferring energy of the lower cycle and one or more outlets of a fluid for transferring energy of the lower cycle; g) at least one fluid transport device having one or more fluid inlets for transferring lower cycle energy and one or more fluid outlets for transferring lower cycle energy; h) at least one control device adapted to maintain a substantially optimal temperature of the emissions of the heat recovery device and having inputs of data on the temperature of the emissions of the heat recovery device and outputs for the operational control of the combustion intensity in the boiler using combustible fuel , wherein said upper cycle energy transfer fluid feeds said one or more fluid inlets for transferring upper cycle energy of the upper cycle engine; combustible fuel feeds said one or more inputs of combustible fuel and / or heat of the upper cycle engine; named one or more outputs of the engine of the upper cycle feed exhaust gases named one or more inputs of the device for heat recovery; one or more sources of fuel and / or heat supply said one or more inputs of fuel and / or heat of the heat recovery device; said one or more fluid outlets under pressure from a heat recovery device supply said fluid inlets under pressure from a lower cycle engine; the named one or more outputs of the discharge of fluid into the environment of the device for heat recovery is vented to the atmosphere (directly or indirectly) or by means of a device for transporting the fluid can be directed to the named inputs of the upper cycle engine; the one or more fluid exits for transferring energy of the lower cycle of the lower cycle engine are sources to the one or more exits of the heat exchanger, the one or more exits of the heat exchanger are connected to the one or more inlets of the fluid transport device and the one or more exits devices for transporting a fluid connected to the named one or more inputs of the device for heat recovery. 22. The combined cycle power plant of claim 21, wherein said top cycle engine comprises one or more gas turbines. 23. Power plant with a combination of cycles according to item 21, which uses a single upper cycle, and the named engine of the lower cycle contains one or more steam turbines. 24. The combined cycle power plant of claim 21, wherein said high pressure fluid contains predominantly a single high pressure steam. 25. The combined cycle power plant of Claim 21, wherein said high pressure fluid contains predominantly supercritical high pressure steam at a rated load of the power plant. 26. Power plant with a combination of cycles according to item 21, in which the aforementioned device for heat recovery contains a heat recovery steam generator mainly with a single pressure level. 27. The combined cycle power plant of claim 21, wherein said upper cycle engine operates substantially at a peak in productivity and / or a level of power generation, and said lower cycle engine is used to modulate the total amount of energy generated by the combined cycle power station. 28. Power plant with combined cycles containing a) at least one upper cycle engine; b) at least one heat recovery device, essentially maintaining fuel combustion continuously, operating primarily at a single pressure, coupled to the said upper cycle engine so that the upper cycle fluid is ejected from the said upper cycle engine into the named device heat recovery; c) at least one lower cycle engine coupled to said heat recovery device such that the lower cycle fluid enters from said heat recovery device to said lower cycle engine and returns to said heat recovery device; d) at least one control device adapted to maintain the optimum temperature of the exhausts of the heat recovery device, having inputs of information about the temperature of the exhausts of the heat recovery device and outputs for the operational control of the intensity of fuel combustion in the said heat recovery device. 29. Power plant with combined cycles, containing a) at least one upper cycle engine; b) upper cycle fluid with a nominal heat transfer provided by the upper cycle engine; c) at least one heat recovery device, operating essentially with constant maintenance of fuel combustion, mainly at a single pressure, moreover, said heat recovery device has an exhaust compartment; d) at least one lower cycle engine; e) a lower cycle fluid with a nominal heat transfer provided by the lower cycle engine; e) at least one means of constantly adding heat to said heat recovery device in such a way that the ratio by weight of the upper cycle fluid flow to the lower cycle fluid flow through the exhaust compartment of the heat recovery device can be maintained substantially equal to the ratio of the nominal heat transfer of the fluid of the lower cycle to the nominal heat transfer of the fluid of the upper cycle due to the modulation of the means for constant maintenance of fuel combustion in a heat recovery device. 30. Power plant with combined cycles containing a) at least one upper cycle engine; b) upper cycle fluid with a nominal heat transfer provided by the upper cycle engine; c) at least one heat recovery device, operating essentially with constant maintenance of fuel combustion, mainly at a single pressure; d) at least one lower cycle engine; e) a lower cycle fluid with a nominal heat transfer provided by the lower cycle engine; e) at least one means of constantly adding heat to said heat recovery device such that the ratio by weight of the upper cycle fluid stream to the lower cycle fluid stream through the exhaust compartment of the heat recovery device at various levels of power generation is maintained equal to the ratio of the nominal heat transfer of the lower cycle fluid to the nominal heat transfer of the upper cycle fluid by modulating the operation of means for constantly maintaining fuel ignition in a heat recovery device. 31. Power plant with combined cycles containing a) at least one upper cycle engine; b) at least one heat recovery device, operating essentially with constant maintenance of fuel combustion, mainly at a single pressure; c) at least one lower cycle engine, in which the functional coefficient of the ratio of the nominal productivity of the sum of all lower cycle engines to the sum of all upper cycle engines exceeds 0.75; d) at least one control device adapted to maintain the optimum temperature of the exhausts of the heat recovery device, having inputs of information about the temperature of the exhausts of the heat recovery device and outputs for operatively controlling the intensity of fuel combustion in the said heat recovery device. 32. A method of operating a power plant with combined cycles, including a) the start of at least one upper cycle engine for the operation of the shaft with exhaust gases, b) essentially continuously supported combustion of fuel in at least one heat recovery device operating primarily at normal pressure, which, together with the additional heat of the exhaust gases of the upper cycle engine, drives at least one lower cycle engine for create shaft work; c) monitoring fuel combustion in the heat recovery device to ensure the optimal level of heat transfer from the power plant with combined cycles by using the temperature at the exit from the heat recovery device as input. 33. A method of operating a power plant with a combination of cycles, including: a) starting at least one upper cycle engine for shaft operation and upper cycle fluid exhaust with a nominal heat transfer provided by the upper cycle engine; b) essentially continuously supported combustion of fuel in at least one heat recovery device operating primarily at normal pressure, which, together with the additional heat of the exhaust gases of the upper cycle engine, drives at least one lower cycle engine for the operation of the shaft and the exhaust of the lower cycle fluid with a nominal heat transfer provided by the lower cycle engine; c) controlling the coefficient of the ratio of the mass flow of the upper cycle fluid to the lower cycle fluid flow through the exhaust compartment of the heat recovery device to maintain substantially equal to the ratio of the nominal heat transfer of the lower cycle fluid to the nominal heat transfer of the upper cycle fluid by controlling combustion fuel in the heat recovery device.

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