EP0767290A1 - Process for operating a power plant - Google Patents

Abstract

A power plant comprises a gas turbine group followed by a waste heat steam generator and then a steam circuit. The gas turbine group has at least one compressor, burner, turbine and generator. In one of the heat-exchanger stages of the waste heat steam generator (15) that operates in the lower temperature regime, a water quantity of more than 100% circulates. The excess above 100% branches off at the end of this stage and is evaporated in at least one pressure stage (26). The steam (37) is passed to a turbine (17). The still hot water (36) from the pressure stage goes to a feed water reservoir and deaerator (22), whence further steam arising is passed to another turbine (18).

Description

Technical field

The present invention relates to a method for operating a power plant according to the preamble of claim 1.

State of the art

In a power plant, which consists of a gas turbine group, a downstream heat recovery steam generator and a subsequent steam circuit, it is advantageous to achieve a maximum efficiency to provide a supercritical steam process in the steam circuit.
Such a circuit has become known from CH-480 535. In this circuit, a mass flow of the gas turbine cycle medium is branched off and used recuperatively in the gas turbine for the purpose of optimal utilization of waste heat from the gas turbine group in the lower temperature range of the heat recovery steam generator. Both the gas turbine and steam processes have sequential combustion. However, in the case of modern, preferably single-shaft, gas turbines, this configuration leads to an undesirable complication in terms of design.

Presentation of the invention

The invention seeks to remedy this. The invention, as characterized in the claims, is based on the task of maximizing the steam circuit-side heat absorption in the lower temperature range of the heat recovery steam generator in a power plant of the type mentioned, in connection with a single-shaft gas turbine.

The main advantages of the invention can be seen in the fact that, despite the simplest design, a better use of the exhaust gases from the last turbine down to 100 ° C and lower is achieved by the steam circuit side heat absorption within a first heat exchange stage in the lower temperature range of the heat recovery steam generator, common known as the economizer.

Advantageous and expedient developments of the task solution according to the invention are characterized in the further claims.

An exemplary embodiment of the invention is explained in more detail below with reference to the drawings. All elements not necessary for the immediate understanding of the invention have been omitted. The direction of flow of the media is indicated by arrows. Identical elements are provided with the same reference symbols in the various figures.

Brief description of the drawings

• Fig. 1 eine Schaltung einer Kraftwerksanlage und
• Fig. 2 ein H/T-Diagramm dieser Schaltung gemäss Fig. 1.

It shows:

• Fig. 1 shows a circuit of a power plant and
• FIG. 2 shows an H / T diagram of this circuit according to FIG. 1.

Ways of carrying out the invention, commercial usability

Fig. 1 shows a power plant, which consists of a gas turbine group I., a downstream of the gas turbine group I. Heat recovery steam generator II., And a heat recovery steam generator II. Downstream. consists.

The present gas turbine group I is based on sequential combustion. The provision of the fuel necessary for operating the various combustion chambers, which cannot be seen in FIG. 1, can be accomplished, for example, by coal gasification cooperating with the gas turbine group. Of course, it is also possible to obtain the fuel used from a primary network. If the supply of a gaseous fuel for operating the gas turbine group is provided via a pipeline, the potential from the pressure and / or temperature difference between the primary network and the consumer network can be recuperated for the needs of the gas turbine group, or in general the circuit. The present gas turbine group, which can also act as an autonomous unit, consists of a compressor 1, a first combustion chamber 2 connected downstream of the compressor, a first turbine 3 connected downstream of this combustion chamber 2, a second combustion chamber 4 connected downstream of this turbine 3 and a second combustion chamber 4 connected downstream of this Turbine 5. The flow machines 1, 3, 5 mentioned have a uniform rotor shaft 39. This rotor shaft 39 itself is preferably mounted on two bearings which are not visible in the figure and which are placed on the head side of the compressor 1 and downstream of the second turbine 5. Depending on the design, for example to increase the specific output, the compressor 1 can be divided into two or more partial compressors, not shown. In such a constellation, an intercooler is then connected downstream of the first partial compressor and upstream of the second partial compressor, in which the partially compressed air is intercooled becomes. The heat generated in this intercooler, which is also not shown, is returned to the process in an optimal, useful manner. The sucked-in air 6 flows as compressed air 7 into a housing (not shown in any more detail) which includes the compressor outlet and the first turbine 3. The first combustion chamber 2, which is preferably designed as a coherent annular combustion chamber, is also accommodated in this housing. Of course, the compressed air 7 can be supplied to the first combustion chamber 2 from an air storage system, not shown. The annular combustion chamber 2 has, on the head side, distributed over the circumference, a number of burners, not shown, which are preferably designed as premix burners. Diffusion burners can also be used here. In order to reduce the pollutant emissions from this combustion, particularly as far as the NOx emissions are concerned, it is, however, advantageous to provide an arrangement of premix burners in accordance with EP-PS-0 321 809, the subject matter of the invention from the cited publication being an integral part of this description is also the way of supplying a fuel 12 described there. As far as the arrangement of the premix burners in the circumferential direction of the annular combustion chamber 2 is concerned, such a one can deviate from the usual configuration of the same burners if necessary, and premix burners of different sizes can be used instead . This is preferably done in such a way that a small premix burner of the same configuration is arranged between two large premix burners. The large premix burners, which have the function of main burners, are related to the small premix burners, which are the pilot burners of this combustion chamber, in terms of the size of the burner air flowing through them, i.e. the compressed air from the compressor 1, which is determined on a case-by-case basis . The pilot burners work as independent premix burners in the entire load range of the combustion chamber, whereby the air ratio remains almost constant. The Zuoder The main burner is switched off according to certain system-specific specifications. Because the pilot burners can be operated with the ideal mixture in the entire load range, the NOx emissions are very low even at partial load. With such a constellation, the circulating streamlines in the front area of the annular combustion chamber 2 come very close to the vortex centers of the pilot burners, so that ignition is only possible with these pilot burners. When starting up, the amount of fuel that is supplied via the pilot burner is increased until it is controlled, ie until the full amount of fuel is available. The configuration is chosen so that this point corresponds to the respective load shedding conditions of the gas turbine group. The further increase in output then takes place via the main burner. At the peak load of the gas turbine group, the main burners are also fully controlled. Because the configuration of "small" hot vortex centers between the "big" cooler vortex centers originating from the main burners, which is initiated by the pilot burners, turns out to be extremely unstable, even with main burners operated lean in the part-load range, a very good burnout is achieved with low CO and in addition to NOx emissions UHC emissions reached, ie the hot swirls of the pilot burner immediately penetrate the small swirls of the main burner. Of course, the annular combustion chamber 2 can consist of a number of individual tubular combustion chambers, which are also arranged in an inclined ring, sometimes also helically, around the rotor axis. This annular combustion chamber 2, regardless of its design, is and can be arranged geometrically so that it has practically no influence on the rotor length. The hot gases 8 from this annular combustion chamber 2 act on the immediately downstream first turbine 3, whose caloric relaxing effect on the hot gases is deliberately kept to a minimum, ie this turbine 3 will therefore consist of no more than two rows of moving blades. In such a turbine 3 it will be necessary to equalize the pressure on the end faces for the purpose of stabilization of the axial thrust. The hot gases 9, which are partially relaxed in the turbine 3 and flow directly into the second combustion chamber 4, have a very high temperature for the reasons explained, preferably it should be designed for the specific operation so that it is still around 1000 ° C. This second combustion chamber 4 has essentially the shape of a coherent annular axial or quasi-axial ring cylinder. This combustion chamber 4 can of course also consist of a number of axially, quasi-axially or helically arranged and self-contained combustion chambers. As far as the configuration of the annular combustion chamber 4 consisting of a single combustion chamber is concerned, a plurality of fuel lances not shown in the figure are arranged in the circumferential direction and radially of this annular cylinder. This combustion chamber 4 has no burner: The combustion of a fuel 13 injected into the partially released hot gases 9 coming from the turbine 3 takes place here by self-ignition, provided the temperature level permits such an operating mode. Assuming that the combustion chamber 4 is operated with a gaseous fuel, for example natural gas, the outlet temperature of the partially released hot gases 9 from the turbine 3 must still be very high, as set out above around 1000 ° C., and of course also under part-load operation , which plays a causal role in the design of this turbine 2. In order to ensure operational safety and a high degree of efficiency in a combustion chamber designed for self-ignition, it is extremely important that the flame front remains locally stable. For this purpose, a number of elements (not shown in detail) are provided in this combustion chamber 4, preferably arranged in the circumferential direction on the inner and outer walls, which are preferably placed upstream of the fuel lances in the axial direction. The task of these elements is to create vortices which induce a backflow zone, analogous to that in the premixing burners already mentioned. Since this is Combustion chamber 4, due to the axial arrangement and the overall length, is a high-speed combustion chamber in which the average speed of the working gases is greater than approximately 60 m / s, the vortex-generating elements must be designed to conform to the flow. On the inflow side, these should preferably consist of a tetrahedral shape with inclined surfaces. The vortex generating elements can either be placed on the outer surface and / or on the inner surface. Of course, the vortex-generating elements can also be axially displaced from one another. The outflow-side surface of the vortex-generating elements is essentially radial, so that a backflow zone is established from there. However, the auto-ignition in the combustion chamber 4 must also be ensured in the transient load areas and in the partial-load area of the gas turbine group, i.e. auxiliary measures must be taken to ensure auto-ignition in the combustion chamber 4 even if the temperature of the gases in the area flexes the injection of the fuel should stop. In order to ensure reliable self-ignition of the gaseous fuel injected into the combustion chamber 4, a small amount of another fuel with a lower ignition temperature is added to it. For example, fuel oil is very suitable as an "auxiliary fuel". The liquid auxiliary fuel, appropriately injected, fulfills the task of acting as a fuse, so to speak, and enables self-ignition in the combustion chamber 4 even when the partially released hot gases 9 from the first turbine 3 have a temperature below the desired optimal level of 1000 ° C should. This precautionary measure, to provide fuel oil to ensure auto-ignition, is of course always particularly appropriate when the gas turbine group is operated with a greatly reduced load. This precaution also makes a decisive contribution to the combustion chamber 4 being able to have a minimal axial length. The short overall length of the combustion chamber 4, the effect of Swirl-generating elements for flame stabilization and the continuous guarantee of auto-ignition are therefore responsible for ensuring that the combustion takes place very quickly and that the fuel stays in the area of the hot flame front to a minimum. An effect that can be measured directly in terms of combustion relates to NOx emissions, which are minimized in such a way that they are no longer an issue. This starting position also enables the location of the combustion to be clearly defined, which is reflected in an optimized cooling of the structures of this combustion chamber 4. The hot gases 10 processed in the combustion chamber 4 then act on a downstream second turbine 5. The thermodynamic characteristics of the gas turbine group can be designed so that the exhaust gases 11 from the second turbine 5 still have so much caloric potential so that they are shown here using a heat recovery steam generator 15 Steam generation stage II. And steam cycle III. to operate. As was already pointed out in the description of the annular combustion chamber 2, it is arranged geometrically in such a way that it has practically no influence on the rotor length of the gas turbine group. It can also be determined that the second combustion chamber 4, which runs between the outflow plane of the first turbine 3 and the inflow plane of the second turbine 5, has a minimal length. Furthermore, since the expansion of the hot gases in the first turbine 3 takes place over a few rows of moving blades for the reasons set out, a gas turbine group can be provided whose rotor shaft 39 can be supported on two bearings in a technically perfect manner due to its minimized length. The power output of the turbomachines takes place via a generator 15 coupled to the compressor, which can also serve as a starter motor. After expansion in the turbine 5, the exhaust gases 11, which still have a high caloric potential, flow through a waste heat steam generator 15, in which steam is generated in various ways in heat exchange processes, which steam then forms the working medium of the steam circuit connected downstream. The calorificly used exhaust gases then flow into the open as flue gases 38.

Assuming that the exhaust gases 11, which reach the waste heat steam generator 15 at G, the mode of operation of which is described below, the path of the feed water 34 flowing into the waste heat steam generator 15 and being conveyed by a pump 23 being tracked, a temperature of approximately 620 ° C, and under the condition of a minimal temperature jump of 20 ° C for heat transfer, these exhaust gases could only be cooled to 200 ° C in a useful way. In order to remedy this disadvantage here, the amount of feed water 34 is increased so far between points A, namely entry of feed water 34 into heat recovery steam generator 15, and B, branching off at the end of treatment within an economizer stage 15 a, in the example to 180%, that the cooling line (see FIG. 2, item 11/38) of the exhaust gases at point H, namely immediately before the branch B, experiences a kink as a result (see FIG. 2, item 41), which is up to 100 ° C is enough. In connection with the percentage amount of feed water, the relation applies that 100% fixes the nominal amount of water that is dependent on the energy offered by the exhaust gases 11.

The feed water 34, which has a temperature of approximately 60 ° C. and a pressure of approximately 300 bar, is introduced into the waste heat steam generator 15 in A and is there to be thermally upgraded to steam of approximately 540 ° C. The feed water heated to approx. 300 ° C in the economizer 15a is divided into two partial flows in point B. The one, here larger, partial water flow of 100% is thermally processed in the subsequent tube bundle 15b to supercritical high-pressure steam 27. As a result, the main part of the thermal energy is extracted from the exhaust gases 11 between the points G and H, which symbolize the effective distance of the tube bundle 15b mentioned. After a first expansion in a high pressure steam turbine 16 this steam 28 with the remaining energy between points D and E, which symbolizes the effective distance of a further tube bundle 15c in the waste heat steam generator 15, reheated and fed as medium pressure steam 29 to a medium pressure steam turbine 17. The residual expansion of the exhaust steam 30 from the medium-pressure steam turbine 17 then takes place in a low-pressure steam turbine 18, which is coupled to a further generator 19. It is also possible to transmit the power to the generator 14 by coupling to the shaft 39.

A smaller partial water flow 35 is branched off in the area from point B and fed via a throttle element 25 to an evaporation bottle 26, the pressure level of which corresponds to the saturated steam pressure of 150-200 ° C. The resulting steam 37 is fed to the medium-pressure steam turbine 17 at a suitable point. The still hot residual water 36, which is only used as a heat carrier for the evaporation, is passed via a further control element 24 into a feed water tank and degasser 22, in which, in addition to the preheating of the condensate, another steam 33 is developed, which the low-pressure steam turbine 18 at a suitable point is fed.

The ultimately expanded steam 31a, 31b from this low-pressure steam turbine 18 is condensed in a water- or air-cooled condenser 20. By means of a condensate pump 21 acting downstream of this condenser 20, the condensate 32 is conveyed into the feed water tank and degasser 22 already mentioned, from where the circuit already described begins anew.

To improve the use of exergy in the evaporation cascade described, this can be done in more than two stages.

In order to achieve a good use of the exhaust gases 11, a separate one can of course be used in the heat recovery steam generator 15 Steam generating device are integrated, the steam either in the steam circuit III. managed, or in work in a separate expansion machine. However, a partial flow of the exhaust gases can also be branched off and used in a separate waste heat boiler. In this case, an ammonia / water mixture can preferably be used instead of water. However, other fluids such as freon, propane, etc. can also be used. A certain improvement in the use of the exhaust gases from the turbine down to a lower level can also be achieved in that the temperature level at the inlet of the waste heat steam generator is raised by an additional firing (not shown). However, this measure does not bring about any improvement in terms of the attainable efficiency.

FIG. 2 shows the H / T diagram, ie the course and the significant points of the feed water preheating and steam generation and steam superheating of a supercritical steam turbine process that have already been recognized in FIG. 1. The respective reference symbols of this figure are described in more detail in the list of reference symbols below. In addition to the explanations under Fig. 1, which are related to the reproduction of this diagram, the following is added. The feed water is introduced into A at, for example, 60 ° C at 300 bar, and it is to be thermally upgraded to F in steam of 540 ° C by means of gas turbine waste heat. After a first expansion stage in the high-pressure steam turbine, which leads up to 300 ° C, there should be an intermediate overheating from D to E, i.e. also to 540 ° C. The solid line 40 shows the resulting course of the heat absorption and the temperature. Assuming that the exhaust gases from the last gas turbine have a temperature of 620 ° C, and under the condition of a minimum temperature jump of 20 ° C for the heat transfer, these exhaust gases could reach point J, ie in this example only to 200 ° C be usefully cooled. In order to remedy this disadvantage, between points A and B increases the quantity of feed water so far, in the example to 180%, that the cooling curve 11/38 of the exhaust gases at point H experiences a kink as resultant 41 and extends up to I, ie up to 100 ° C. This additional feed water flow is taken off at B and fed to an evaporation cascade (cf. FIG. 1) in such a way that the steam produced can be fed to the medium and low-pressure part of the steam turbine, as can also be seen in FIG. 1. The assessment of the remaining points is also evident from the description of FIG. 1.

Reference list

I.

Gas turbine group

II.

Steam generation stage

III.

Steam cycle

1

compressor

2nd

First combustion chamber

3rd

First turbine

4th

Second combustion chamber

5

Second turbine

6

Intake air

7

Compressed air

8th

Hot gases

9

Partial clamping hot gases

10th

Hot gases

11

Exhaust gases

12th

fuel

13

fuel

14

generator

15

Heat recovery steam generator

15a

Economizer, in the lower temp. Area op. Heat exchange level

15b

Pipe bundle for supercritical high pressure steam

15c

Pipe bundle for reheated medium pressure steam

16

High pressure steam turbine

17th

Medium pressure steam turbine

18th

Low pressure steam turbine

19th

generator

20th

capacitor

21

Feed pump

22

Feed water tank and degasser

23

Feed pump

24th

Governing body

25th

Governing body

26

Evaporation bottle

27

Supercritical high pressure steam

28

Expanded steam from 16th

29

Reheated medium pressure steam

30th

Evaporation from 17 in 18

31a

Relaxed steam from 18

31b

Relaxed steam from 18

32

condensate

33

Steam from 22 in 18

34

Feed water

35

Small partial water flow

36

Hot residual water from 26 in 22

37

Steam from 26

38

Flue gases

39

Rotor shaft

40

Supercritical steam generation curve

41

Resultant

11/38

Cooling curve

A

Feed water after 22

B

Tapping point for pressurized water to 26

B-C

Sum of B-F + D-E, overheating and intermediate overheating

D-E

Reheat in 15c

F

Place supercritical high pressure steam

G

Entry of exhaust gases in 15

H

Flue gas temperature at extraction point B

I.

Exhaust gases from 15 = flue gases

J

Fictitious final flue gas value without withdrawal in B

FROM

Generally over 100%, in the example 180% water flow

B-F

100% water flow

Claims (6)

Method for operating a power plant, essentially consisting of a gas turbine group, a heat recovery steam generator downstream of the gas turbine group and a steam circuit downstream of the heat recovery steam generator, the gas turbine group consisting of at least one compressor unit, at least one combustion chamber, at least one turbine and at least one generator, the exhaust gases consisting of flow through the heat recovery steam generator in the last turbine, in which the generation of at least one steam for operating at least one steam turbine of the steam cycle takes place, characterized in that a heat exchange stage (15a) of the heat recovery steam generator (15) operating in the lower temperature range circulates an over 100% increased amount of liquid that the portion over 100% of this amount of liquid is branched off at the end of this heat exchange stage (15a) and evaporated in at least one pressure stage (26), that a vapor (37) of a D Ampfturbine (17) is supplied at a suitable point, that a still hot amount of liquid (36) from the pressure stage (26) is fed to a feed water tank and degasser (22), and that a steam (33) generated therein to another steam turbine (18) appropriate place is forwarded. A method according to claim 1, characterized in that the gas turbine group (I.) is operated with a sequential combustion. A method according to claim 1, characterized in that the 100% amount of liquid in a direct heat exchange stage (15a) following the heat exchange stage (15b) is processed into supercritical steam (27), which acts on a further steam turbine (16) in such a way that the steam (28) expanded in this steam turbine (16) is returned to the waste heat steam generator (15), that it is processed there in a further heat exchange stage (15c) to superheated steam (29), which then acts on a corresponding pressure stage of a downstream steam turbine (17). A method according to claim 1, characterized in that the feed water tank and degasser (22) is operated as the sole evaporation stage of the steam circuit (III.). A method according to claim 1, characterized in that the portion over 100% of the amount of liquid in a separate heat exchange element in parallel and / or in series with the heat exchange stage (15a) in the lower temperature range. A method according to claim 5, characterized in that the proportion over 100% of the amount of liquid differs from the fluid expanding in the steam circuit (III.), And that the thermal energy generated by the heat exchange is used in a separate working machine.

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