JP4621597B2 - Steam turbine cycle - Google Patents

Steam turbine cycle Download PDF

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JP4621597B2
JP4621597B2 JP2006012124A JP2006012124A JP4621597B2 JP 4621597 B2 JP4621597 B2 JP 4621597B2 JP 2006012124 A JP2006012124 A JP 2006012124A JP 2006012124 A JP2006012124 A JP 2006012124A JP 4621597 B2 JP4621597 B2 JP 4621597B2
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feed water
steam
cycle
turbine
water heater
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JP2007192152A (en
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藤 功 一 後
田 信 雄 沖
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Toshiba Corp
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Toshiba Corp
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Priority to JP2006012124A priority Critical patent/JP4621597B2/en
Priority to KR1020087006641A priority patent/KR20080038233A/en
Priority to PCT/JP2006/325524 priority patent/WO2007083478A1/en
Priority to EP06843008.1A priority patent/EP1965043B1/en
Priority to CN2006800405919A priority patent/CN101300407B/en
Priority to US12/092,796 priority patent/US20090094983A1/en
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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01K—STEAM ENGINE PLANTS; STEAM ACCUMULATORS; ENGINE PLANTS NOT OTHERWISE PROVIDED FOR; ENGINES USING SPECIAL WORKING FLUIDS OR CYCLES
    • F01K7/00—Steam engine plants characterised by the use of specific types of engine; Plants or engines characterised by their use of special steam systems, cycles or processes; Control means specially adapted for such systems, cycles or processes; Use of withdrawn or exhaust steam for feed-water heating
    • F01K7/34—Steam engine plants characterised by the use of specific types of engine; Plants or engines characterised by their use of special steam systems, cycles or processes; Control means specially adapted for such systems, cycles or processes; Use of withdrawn or exhaust steam for feed-water heating the engines being of extraction or non-condensing type; Use of steam for feed-water heating
    • F01K7/40—Use of two or more feed-water heaters in series
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01K—STEAM ENGINE PLANTS; STEAM ACCUMULATORS; ENGINE PLANTS NOT OTHERWISE PROVIDED FOR; ENGINES USING SPECIAL WORKING FLUIDS OR CYCLES
    • F01K17/00—Using steam or condensate extracted or exhausted from steam engine plant
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01K—STEAM ENGINE PLANTS; STEAM ACCUMULATORS; ENGINE PLANTS NOT OTHERWISE PROVIDED FOR; ENGINES USING SPECIAL WORKING FLUIDS OR CYCLES
    • F01K7/00—Steam engine plants characterised by the use of specific types of engine; Plants or engines characterised by their use of special steam systems, cycles or processes; Control means specially adapted for such systems, cycles or processes; Use of withdrawn or exhaust steam for feed-water heating
    • F01K7/16—Steam engine plants characterised by the use of specific types of engine; Plants or engines characterised by their use of special steam systems, cycles or processes; Control means specially adapted for such systems, cycles or processes; Use of withdrawn or exhaust steam for feed-water heating the engines being only of turbine type
    • F01K7/22—Steam engine plants characterised by the use of specific types of engine; Plants or engines characterised by their use of special steam systems, cycles or processes; Control means specially adapted for such systems, cycles or processes; Use of withdrawn or exhaust steam for feed-water heating the engines being only of turbine type the turbines having inter-stage steam heating
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01K—STEAM ENGINE PLANTS; STEAM ACCUMULATORS; ENGINE PLANTS NOT OTHERWISE PROVIDED FOR; ENGINES USING SPECIAL WORKING FLUIDS OR CYCLES
    • F01K7/00—Steam engine plants characterised by the use of specific types of engine; Plants or engines characterised by their use of special steam systems, cycles or processes; Control means specially adapted for such systems, cycles or processes; Use of withdrawn or exhaust steam for feed-water heating
    • F01K7/34—Steam engine plants characterised by the use of specific types of engine; Plants or engines characterised by their use of special steam systems, cycles or processes; Control means specially adapted for such systems, cycles or processes; Use of withdrawn or exhaust steam for feed-water heating the engines being of extraction or non-condensing type; Use of steam for feed-water heating
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01K—STEAM ENGINE PLANTS; STEAM ACCUMULATORS; ENGINE PLANTS NOT OTHERWISE PROVIDED FOR; ENGINES USING SPECIAL WORKING FLUIDS OR CYCLES
    • F01K7/00—Steam engine plants characterised by the use of specific types of engine; Plants or engines characterised by their use of special steam systems, cycles or processes; Control means specially adapted for such systems, cycles or processes; Use of withdrawn or exhaust steam for feed-water heating
    • F01K7/34—Steam engine plants characterised by the use of specific types of engine; Plants or engines characterised by their use of special steam systems, cycles or processes; Control means specially adapted for such systems, cycles or processes; Use of withdrawn or exhaust steam for feed-water heating the engines being of extraction or non-condensing type; Use of steam for feed-water heating
    • F01K7/38—Steam engine plants characterised by the use of specific types of engine; Plants or engines characterised by their use of special steam systems, cycles or processes; Control means specially adapted for such systems, cycles or processes; Use of withdrawn or exhaust steam for feed-water heating the engines being of extraction or non-condensing type; Use of steam for feed-water heating the engines being of turbine type

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  • Engineering & Computer Science (AREA)
  • Chemical & Material Sciences (AREA)
  • Combustion & Propulsion (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Engine Equipment That Uses Special Cycles (AREA)
  • Control Of Turbines (AREA)
  • Control Of Steam Boilers And Waste-Gas Boilers (AREA)

Description

本発明は、サイクル熱効率の高い蒸気タービンサイクルに関する。   The present invention relates to a steam turbine cycle having high cycle thermal efficiency.

従来の技術として、火力発電プラント等で使用されている蒸気タービンサイクルの１つを、図１を用いて説明する。   As a conventional technique, one of steam turbine cycles used in a thermal power plant or the like will be described with reference to FIG.

ボイラ４にて燃料燃焼熱を利用する等してボイラ給水１４を加熱し、充分に高温である過熱蒸気(以下、主蒸気１６と記す)を発生させる。この過熱蒸気は超臨界圧流体である場合もある。   The boiler feed water 14 is heated by using, for example, fuel combustion heat in the boiler 4 to generate superheated steam (hereinafter referred to as main steam 16) having a sufficiently high temperature. This superheated steam may be a supercritical pressure fluid.

主蒸気１６は高圧タービン１に流入し、膨張しながら流れ、圧力、温度共に低下する。   The main steam 16 flows into the high-pressure turbine 1, flows while expanding, and decreases in both pressure and temperature.

高圧タービン１から流出した高圧タービン排気２１の大部分は再熱器５に流入し、より高温になり再熱蒸気１７となって中圧タービン２に流入する。   Most of the high-pressure turbine exhaust 21 that has flowed out of the high-pressure turbine 1 flows into the reheater 5, reaches a higher temperature, becomes reheated steam 17, and flows into the intermediate-pressure turbine 2.

中圧タービン２内で膨張しながら流れた蒸気は、圧力、温度共に低下して低圧タービン３に流入する。   The steam that flows while expanding in the intermediate pressure turbine 2 decreases in both pressure and temperature and flows into the low pressure turbine 3.

低圧タービン３内で膨張しながら流れた蒸気は、圧力、温度共に低下するが、一部、液体の水である飽和蒸気の状態になる場合が多い。この飽和蒸気は復水器１０にて、海水または大気２３等を用いて冷却され復水２５になる。復水２５は復水ポンプ１１により給水加熱器６に送られ、ボイラ給水１４になる。なお、中圧タービン２と低圧タービン３は再熱タービン２４である。   The steam that flows while expanding in the low-pressure turbine 3 decreases in both pressure and temperature, but in some cases, it is in a state of saturated steam that is partly liquid water. This saturated steam is cooled by the condenser 10 using seawater or the atmosphere 23 or the like to become the condensate 25. The condensate 25 is sent to the feed water heater 6 by the condensate pump 11 and becomes boiler feed water 14. The intermediate pressure turbine 2 and the low pressure turbine 3 are reheat turbines 24.

図１では８個の給水加熱器６が示されており、高圧タービン１、中圧タービン２及び低圧タービン３の流路中の抽気位置３１から抽出された抽気蒸気２０により、ボイラ給水１４を加熱する。ここで、より高圧の給水加熱器６ほど、より高圧の抽気が流入する構成になっている。   In FIG. 1, eight feed water heaters 6 are shown, and the boiler feed water 14 is heated by the extraction steam 20 extracted from the extraction position 31 in the flow path of the high pressure turbine 1, the intermediate pressure turbine 2, and the low pressure turbine 3. To do. Here, the higher-pressure feed water heater 6 has a configuration in which higher-pressure extraction air flows.

図１では低圧タービン３は複流として描いてあり、片側の低圧タービン３のみから抽気してあるように描いてあるが、実際は、両側の低圧タービン３から抽気して合流した後、給水加熱器６に流入させている。なお、給水加熱器６によって片側の低圧タービン３どちらか一方から抽気する構成にしてもよい。   In FIG. 1, the low-pressure turbine 3 is drawn as a double flow, and is drawn as if it was extracted from only one low-pressure turbine 3, but actually, after extracting and joining from the low-pressure turbine 3 on both sides, the feed water heater 6 is drawn. It is allowed to flow into. Note that the feed water heater 6 may extract air from either one of the low-pressure turbines 3 on one side.

給水加熱器６は表面式と混合式がある。表面式給水加熱器において抽気蒸気２０は、伝熱面を介した給水との熱交換により凝縮しドレン水１５となり、原則としてより高圧の給水加熱器６から順次流入合流しながらより低圧の給水加熱器６へ流れる。最も低圧の給水加熱器６のドレン水１５は復水器１０へ流れる。なお、ドレン水１５をドレン水ポンプ１３により給水に合流させてもよい。   The feed water heater 6 has a surface type and a mixing type. In the surface type water heater, the extracted steam 20 is condensed by heat exchange with the water supplied through the heat transfer surface to become drain water 15, and in principle, the lower pressure water heater is heated while sequentially flowing in from the higher pressure water heater 6. Flows to vessel 6. The drain water 15 of the lowest pressure feed water heater 6 flows to the condenser 10. Note that the drain water 15 may be joined to the water supply by the drain water pump 13.

混合式給水加熱器は、抽気が直接、給水と混合して加熱する構造であり、給水に溶存している酸素等を脱気させる脱気器９は、混合式給水加熱器に含まれる。   The mixed feed water heater has a structure in which the extracted air is directly mixed with the feed water and heated, and the deaerator 9 for degassing oxygen and the like dissolved in the feed water is included in the mixed feed water heater.

混合式給水加熱器の直後には、より高圧の給水加熱器６に給水を送るために、給水ポンプ１２が設けられる。図１では混合式給水加熱器への抽気は中圧タービン排気抽気３２だが、そうでなくてもよい。脱気器９はなくてもよいが、ない場合も複数の給水加熱器６の間の内、適当な位置に給水ポンプ１２を設ける。全ての給水加熱器６で加熱された給水は、ボイラ４に流入する。   Immediately after the mixed feed water heater, a feed water pump 12 is provided to send feed water to the higher pressure feed water heater 6. In FIG. 1, the bleed air to the mixed feed water heater is the medium pressure turbine exhaust bleed air 32, but this need not be the case. Although the deaerator 9 may not be provided, the feed water pump 12 is provided at an appropriate position among the plurality of feed water heaters 6 even when the deaerator 9 is not provided. The feed water heated by all the feed water heaters 6 flows into the boiler 4.

図１では、高圧タービン１と中圧タービン２と低圧タービン３は１つの回転軸１９で連結されており、発電機１８に接続されている。高圧タービン１、中圧タービン２、低圧タービン３の内部で膨張する事によって蒸気の保有するエンタルピが軸動力に変換され、発電機１８で発電する。各タービンを１つの回転軸１９で連結して１台の発電機１８に接続しなくてもよい。   In FIG. 1, the high-pressure turbine 1, the intermediate-pressure turbine 2, and the low-pressure turbine 3 are connected by a single rotating shaft 19 and are connected to a generator 18. By expanding inside the high-pressure turbine 1, the intermediate-pressure turbine 2, and the low-pressure turbine 3, the enthalpy held by the steam is converted into shaft power and is generated by the generator 18. Each turbine may be connected by one rotating shaft 19 and not connected to one generator 18.

図１では低圧タービン３は複流として描いてあるが、これは流入蒸気を２分割し２台の低圧タービン３に流入させる構造であり、４分割する構造でも、分割しなくてもよい。また図１では、中圧タービン２は単流として描いてあるが、複流でもよい。さらに図１では、中圧タービン２と低圧タービン３とは別々の蒸気タービンとして描いてあるが、再熱タービン２４、１台のみとしてもよい。   In FIG. 1, the low-pressure turbine 3 is depicted as a double flow, but this is a structure in which the incoming steam is divided into two parts and flows into the two low-pressure turbines 3, and may be divided into four parts or not. In FIG. 1, the intermediate pressure turbine 2 is depicted as a single flow, but it may be a double flow. Further, in FIG. 1, the intermediate pressure turbine 2 and the low pressure turbine 3 are depicted as separate steam turbines, but only one reheat turbine 24 may be provided.

このような抽気蒸気２０を用いた再生サイクルと、高圧タービン排気２１を再熱器５で加熱し再熱タービン２４に流入させる再熱サイクルは、どちらも変形ランキンサイクルであり、単純なランキンサイクルから熱効率を向上させる。なお発電プラントの場合、熱効率は発電量÷ボイラ入熱量とほぼ等しい。   Both the regeneration cycle using the extracted steam 20 and the reheat cycle in which the high-pressure turbine exhaust 21 is heated by the reheater 5 and flows into the reheat turbine 24 are modified Rankine cycles. Improve thermal efficiency. In the case of a power plant, the thermal efficiency is almost equal to the power generation amount divided by the boiler heat input amount.

さて、サイクル熱効率はサイクル構成によって変化するが、各抽気蒸気２０の温度や流量によっても変化する。特に近年、高温用材料の進歩に伴い、蒸気の高温化が進みサイクル熱効率は向上しつつあるが、蒸気の高温条件下にてサイクル構成を改善する余地がある。   The cycle thermal efficiency varies depending on the cycle configuration, but also varies depending on the temperature and flow rate of each extraction steam 20. Particularly, in recent years, with the progress of high temperature materials, the temperature of steam is increasing and the cycle thermal efficiency is improving. However, there is room for improving the cycle configuration under the high temperature conditions of steam.

なお、非特許文献には、「再熱点からの抽気によるヒータエンタルピ上昇はそれより低圧のヒータにおける平均エンタルピ上昇の約１.８倍が最適性能」という記述がある。
Bartlett著「Steam Turbine Performance and Economics」 Note that the non-patent document states that “the heater enthalpy rise due to extraction from the reheat point is about 1.8 times the average enthalpy rise in the heater at a lower pressure than that”.
"Steam Turbine Performance and Economics" by Bartlett

本発明は、サイクル熱効率の高い蒸気タービンサイクルを提供することを目的とする。   An object of this invention is to provide a steam turbine cycle with high cycle thermal efficiency.

請求項１に係る発明は、高圧タービンと、再熱タービンと、ボイラと、前記タービンからの抽気蒸気により前記ボイラへの給水を加熱する給水加熱器と、給水ポンプと、復水器とを備え、作動流体が水である１段再熱サイクルであり、かつ再生サイクルであるランキンサイクルにおいて、前記ボイラ出口の蒸気温度は、５９０℃以上であり、前記高圧タービンの排気からの抽気に対応した第１の給水加熱器における給水温度上昇と、前記第１の給水加熱器より給水が低圧である第２の給水加熱器における給水温度上昇の平均との、温度上昇比が１.９以上３.５以下であることを特徴とした蒸気タービンサイクルである。   The invention according to claim 1 includes a high-pressure turbine, a reheat turbine, a boiler, a feed water heater that heats feed water to the boiler with extracted steam from the turbine, a feed water pump, and a condenser. In the Rankine cycle, in which the working fluid is water, and the Rankine cycle is a regeneration cycle, the steam temperature at the boiler outlet is 590 ° C. or higher, corresponding to the extraction from the exhaust of the high-pressure turbine. The temperature rise ratio between the feed water temperature rise in the first feed water heater and the average feed water temperature rise in the second feed water heater whose feed water is lower than that of the first feed water heater is from 1.9 to 3.5. A steam turbine cycle characterized by the following.

請求項２に係る本発明は、高圧タービンと、再熱タービンと、ボイラと、前記タービンからの抽気蒸気により前記ボイラへの給水を加熱する給水加熱器と、給水ポンプと、復水器とを備え、作動流体が水である１段再熱サイクルであり、かつ再生サイクルであるランキンサイクルにおいて、前記ボイラ出口の蒸気温度は、５９０℃以上であり、前記高圧タービンの排気からの抽気に対応した第１の給水加熱器における給水の比エンタルピ上昇と、前記第１の給水加熱器より給水が低圧である第２の給水加熱器における給水の比エンタルピ上昇の平均との、比エンタルピ上昇比が１.９以上３.５以下であることを特徴とした蒸気タービンサイクルである。   The present invention according to claim 2 includes a high-pressure turbine, a reheat turbine, a boiler, a feed water heater that heats feed water to the boiler by extracted steam from the turbine, a feed pump, and a condenser. In the Rankine cycle, the working fluid is water, and the Rankine cycle is a regeneration cycle, the steam temperature at the boiler outlet is 590 ° C. or higher, corresponding to the extraction from the exhaust of the high-pressure turbine The specific enthalpy increase ratio between the increase in the specific enthalpy of the feed water in the first feed water heater and the average increase in the specific enthalpy of the feed water in the second feed water heater whose feed water is lower than that in the first feed water heater is 1. It is a steam turbine cycle characterized by being 9.9 or more and 3.5 or less.

請求項３に係る発明は、高圧タービンと、再熱タービンと、ボイラと、前記タービンからの抽気蒸気により前記ボイラへの給水を加熱する給水加熱器と、給水ポンプと、復水器とを備え、作動流体が水である１段再熱サイクルであり、かつ再生サイクルであるランキンサイクルにおいて、前記ボイラ出口の蒸気温度は、５９０℃以上であり、前記高圧タービンの排気からの抽気に対応した第１の給水加熱器における給水温度上昇と、前記第１の給水加熱器を除く給水加熱器における給水温度上昇の平均との、温度上昇比が１.９以上３.５以下であることを特徴とした蒸気タービンサイクルである。   The invention according to claim 3 includes a high-pressure turbine, a reheat turbine, a boiler, a feed water heater that heats feed water to the boiler by extracted steam from the turbine, a feed water pump, and a condenser. In the Rankine cycle, in which the working fluid is water, and the Rankine cycle is a regeneration cycle, the steam temperature at the boiler outlet is 590 ° C. or higher, corresponding to the extraction from the exhaust of the high-pressure turbine. A temperature rise ratio between a feed water temperature rise in one feed water heater and an average feed water temperature rise in a feed water heater excluding the first feed water heater is 1.9 to 3.5. Steam turbine cycle.

請求項４に係る発明は、高圧タービンと、再熱タービンと、ボイラと、前記タービンからの抽気蒸気により前記ボイラへの給水を加熱する給水加熱器と、給水ポンプと、復水器とを備え、作動流体が水である１段再熱サイクルであり、かつ再生サイクルであるランキンサイクルにおいて、前記ボイラ出口の蒸気温度が５９０℃以上であり、前記高圧タービンの排気からの抽気に対応した第１の給水加熱器における給水の比エンタルピ上昇と、前記第１の給水加熱器を除く給水加熱器における給水の比エンタルピ上昇の平均との、比エンタルピ上昇比が１.９以上３.５以下であることを特徴とした蒸気タービンサイクルである。   The invention according to claim 4 includes a high-pressure turbine, a reheat turbine, a boiler, a feed water heater that heats feed water to the boiler by extracted steam from the turbine, a feed water pump, and a condenser. In the Rankine cycle in which the working fluid is water and the Rankine cycle is a regeneration cycle, the steam temperature at the boiler outlet is 590 ° C. or higher, and the first corresponding to the extraction from the exhaust of the high-pressure turbine. The specific enthalpy increase ratio between the specific enthalpy increase in the feed water in the feed water heater and the average of the specific enthalpy rise in the feed water heater excluding the first feed water heater is 1.9 to 3.5 It is the steam turbine cycle characterized by this.

本発明によれば、サイクル熱効率の高い蒸気タービンサイクルを提供することができる。   According to the present invention, a steam turbine cycle with high cycle thermal efficiency can be provided.

第１の実施の形態
以下、本発明に係る蒸気タービンサイクルの第１の実施の形態について、図面を参照して説明する。ここで、図１は本発明の第１の実施の形態を示す図である。 First Embodiment Hereinafter, a first embodiment of a steam turbine cycle according to the present invention will be described with reference to the drawings. Here, FIG. 1 is a diagram showing a first embodiment of the present invention.

本実施の形態の蒸気タービンサイクルは、 高圧タービン１と、再熱タービン２４と、ボイラ４と、高圧タービン１及び再熱タービン２４からの抽気蒸気によりボイラ４への給水を加熱する給水加熱器６と、給水ポンプ１２と、復水器１０とを備え、作動流体が水である１段再熱サイクルであり、かつ再生サイクルであるランキンサイクルを構成する。   The steam turbine cycle of the present embodiment includes a high-pressure turbine 1, a reheat turbine 24, a boiler 4, and a feed water heater 6 that heats feed water to the boiler 4 using extracted steam from the high pressure turbine 1 and the reheat turbine 24. And a water supply pump 12 and a condenser 10, and constitutes a Rankine cycle that is a one-stage reheat cycle in which the working fluid is water and a regeneration cycle.

また、ボイラ４出口の蒸気温度は５９０℃以上となっており、さらに高圧タービン排気２１からの抽気（高圧タービン排気抽気）２２に対応した第１の給水加熱器７における給水温度上昇と、第１の給水加熱器７より給水が低圧である第２の給水加熱器８における給水温度上昇の平均との温度上昇比が１.９以上３.５以下となっている。   In addition, the steam temperature at the outlet of the boiler 4 is 590 ° C. or higher, and the feed water temperature rise in the first feed water heater 7 corresponding to the bleed air (high pressure turbine exhaust bleed air) 22 from the high pressure turbine exhaust 21, The temperature rise ratio with respect to the average of the feed water temperature rise in the second feed water heater 8 whose feed water is lower than that of the feed water heater 7 is 1.9 to 3.5.

各抽気蒸気２０の流量と、各抽気位置３１を調節することにより、第１の給水加熱器７及び第２の給水加熱器８における給水温度上昇を調整することができる。なお、高圧タービン排気抽気２２や中圧タービン排気抽気３２に関して温度変更する場合は、高圧タービン１と中圧タービン２の排気仕様を変更する事になる。   By adjusting the flow rate of each extraction steam 20 and each extraction position 31, it is possible to adjust the feed water temperature rise in the first feed water heater 7 and the second feed water heater 8. Note that when the temperature is changed with respect to the high-pressure turbine exhaust bleed air 22 and the intermediate-pressure turbine exhaust bleed air 32, the exhaust specifications of the high-pressure turbine 1 and the intermediate-pressure turbine 2 are changed.

ボイラ４入口の給水温度はボイラ４側で規定される事が多いので、値を固定して最適化計算をしたところ、１.９以上３.５以下の温度上昇比条件の時、サイクル熱効率は最大になった。   Since the boiler 4 inlet water supply temperature is often specified on the boiler 4 side, the optimization calculation was performed with the value fixed. When the temperature rise ratio was 1.9 to 3.5, the cycle thermal efficiency was It became the maximum.

第１の給水加熱器７における給水温度上昇と、第２の給水加熱器８における給水温度上昇の平均との温度上昇比は、給水加熱器６の個数や、排気損失等といった蒸気タービンの機械上の差異や、発電プラントにおける発電出力に相当する規模、細かい構成の差異等の影響によって、範囲幅がある。   The temperature rise ratio between the feed water temperature rise in the first feed water heater 7 and the average of the feed water temperature rise in the second feed water heater 8 depends on the number of feed water heaters 6, exhaust loss, etc. There is a range width due to the influence of the difference in size, the scale corresponding to the power generation output in the power plant, the difference in the detailed configuration, and the like.

なお前述したように非特許文献では、比エンタルピ上昇比として１.８が最良という旨の記述があるが、温度上昇比の範囲では、比エンタルピ上昇比が１.８になる事は、通常の発電プラントではない。   As described above, in non-patent literature, there is a description that 1.8 is the best as the specific enthalpy increase ratio. However, within the range of the temperature increase ratio, the specific enthalpy increase ratio is 1.8. It is not a power plant.

この事象は以下の理由によると推定される。   This event is presumed to be due to the following reason.

蒸気タービンの出力は、各タービンの各段落における「熱落差即ち比エンタルピ減少量×蒸気質量流量」の総和なので、なるべく比エンタルピの低い位置から抽気した方が、蒸気タービンが仕事をした後にボイラ給水１４を暖めることになるので、効率が高くなる効果がある。しかしながら一方で、ボイラ給水１４のボイラ４入口温度が高い方が再生サイクルとしての効率は高くなるため、その効果も考慮する必要がある。   The output of the steam turbine is the sum of the “heat drop, ie, specific enthalpy reduction amount x steam mass flow rate” in each stage of each turbine, so it is better to extract from the position where the specific enthalpy is as low as possible after the steam turbine has worked. Since 14 is warmed, the efficiency is increased. However, on the other hand, the higher the boiler 4 inlet temperature of the boiler feed water 14 is, the higher the efficiency as the regeneration cycle is.

ボイラ４入口温度を規定すると、最も器内蒸気圧力が高い給水加熱器２６のボイラ４入口温度とほぼ同じ飽和温度である蒸気が必要で、抽気蒸気２０の圧力が規定される。他の給水加熱器７，８はそこに至るまでの温度上昇を段階的に賄う。   When the boiler 4 inlet temperature is defined, steam having a saturation temperature substantially the same as the boiler 4 inlet temperature of the feed water heater 26 having the highest internal steam pressure is required, and the pressure of the extracted steam 20 is defined. Other feed water heaters 7 and 8 cover the temperature rise up to that step by step.

高圧タービン排気抽気２２は、比較的高圧でありながら比エンタルピが低い蒸気であり、また再熱器５で加熱された後の蒸気からの抽気でないので、この高圧タービン排気抽気２２のエンタルピを用いてボイラ給水１４を加熱する事を多めにすると、サイクル全体の熱効率が高くなる。   The high-pressure turbine exhaust bleed air 22 is a steam having a relatively high pressure and a low specific enthalpy, and is not extracted from the steam heated by the reheater 5, so that the enthalpy of the high-pressure turbine exhaust bleed air 22 is used. If heating the boiler feed water 14 is increased, the thermal efficiency of the entire cycle is increased.

即ち、図３に概略を示すように、温度上昇比には熱効率が最も良くなる所定の最適値があり、この最適値は温度上昇比が１より充分に高い方が良い。この最適値は主蒸気１６の条件によって異なり、高温蒸気ではより高い値になると推定される。   That is, as schematically shown in FIG. 3, the temperature rise ratio has a predetermined optimum value at which the thermal efficiency is the best, and this optimum value should be sufficiently higher than 1. This optimum value varies depending on the conditions of the main steam 16, and is estimated to be a higher value for high-temperature steam.

上述したように、第１の給水加熱器７における給水温度上昇と、第２の給水加熱器８における給水温度上昇の平均との温度上昇比を１.９以上３.５以下にすることによって、サイクル熱効率を向上させることができる。   As described above, by setting the temperature rise ratio between the feed water temperature rise in the first feed water heater 7 and the average feed water temperature rise in the second feed water heater 8 to 1.9 to 3.5, Cycle heat efficiency can be improved.

第２の実施の形態
次に図１により本発明の第２の実施の形態について説明する。 Second Embodiment Next, a second embodiment of the present invention will be described with reference to FIG.

本実施の形態の蒸気タービンサイクルは、高圧タービン１と、再熱タービン２４と、ボイラ４と、高圧タービン１及び再熱タービン２４からの抽気蒸気によりボイラ４への給水を加熱する給水加熱器６と、給水ポンプ１２と、復水器１０とを備え、作動流体が水である１段再熱サイクルであり、かつ再生サイクルであるランキンサイクルを構成する。   The steam turbine cycle of the present embodiment includes a high-pressure turbine 1, a reheat turbine 24, a boiler 4, and a feed water heater 6 that heats the feed water to the boiler 4 using the extracted steam from the high pressure turbine 1 and the reheat turbine 24. And a water supply pump 12 and a condenser 10, and constitutes a Rankine cycle that is a one-stage reheat cycle in which the working fluid is water and a regeneration cycle.

また、ボイラ４出口の蒸気温度は５９０℃以上であり、高圧タービン排気抽気２２に対応した第１の給水加熱器７における給水の比エンタルピ上昇と、第１の給水加熱器７より給水が低圧である第２の給水加熱器８における給水の比エンタルピ上昇の平均との比エンタルピ上昇比が１.９以上３.５以下となっている。   Further, the steam temperature at the outlet of the boiler 4 is 590 ° C. or higher, the specific enthalpy of the feed water in the first feed water heater 7 corresponding to the high pressure turbine exhaust bleed 22 is increased, and the feed water is lower in pressure from the first feed water heater 7. The specific enthalpy increase ratio with respect to the average specific enthalpy increase of the feed water in a certain second feed water heater 8 is 1.9 or more and 3.5 or less.

各抽気蒸気２０の流量と、各抽気位置３１を調節することによって、第１の給水加熱器７及び第２の給水加熱器８における給水の比エンタルピ上昇を調整することができる。なお、高圧タービン排気抽気２２や中圧タービン排気抽気３２からの抽気蒸気２０に関して比エンタルピを変更する場合は、高圧タービン１と中圧タービン２の排気仕様を変更する事になる。   By adjusting the flow rate of each extraction steam 20 and each extraction position 31, it is possible to adjust the specific enthalpy increase of water supply in the first feed water heater 7 and the second feed water heater 8. When the specific enthalpy is changed with respect to the extracted steam 20 from the high-pressure turbine exhaust bleed air 22 or the intermediate-pressure turbine exhaust bleed air 32, the exhaust specifications of the high-pressure turbine 1 and the intermediate-pressure turbine 2 are changed.

ボイラ４入口の給水温度はボイラ４側で規定される事が多いので、値を固定して最適化計算をしたところ、１.９以上３.５以下の比エンタルピ上昇比条件の時、サイクル熱効率は最大になった。   Since the boiler 4 inlet water supply temperature is often specified on the boiler 4 side, optimization calculation was performed with a fixed value. When the specific enthalpy increase ratio was 1.9 or more and 3.5 or less, the cycle thermal efficiency Became the maximum.

比エンタルピ上昇比には、給水加熱器６の個数や、排気損失等といった蒸気タービンの機械上の差異や、発電プラントにおける発電出力に相当する規模、細かい構成の差異等の影響による範囲幅がある。   The specific enthalpy increase ratio has a range width due to the influence of the difference in steam turbine machinery such as the number of feed water heaters 6 and exhaust loss, the scale corresponding to the power generation output in the power plant, and the difference in fine configuration. .

なお前述したように非特許文献では、比エンタルピ上昇比として１.８が最良という旨の記述があるが、この文献では主蒸気１６の温度が述べられておらず、想定している主蒸気１６温度が異なるために、比エンタルピ上昇比の最適値が異なっていると考えられる。   As described above, in the non-patent document, there is a description that 1.8 is the best as the specific enthalpy increase ratio, but in this document, the temperature of the main steam 16 is not described, and the assumed main steam 16 Since the temperature is different, the optimum value of the specific enthalpy increase ratio is considered to be different.

本実施の形態の場合も、前述した第１の実施の形態と同様に、図３に概略を示すように、比エンタルピ上昇比には熱効率が最も良くなる所定の最適値があり、この最適値は比エンタルピ上昇比が１より充分に高い方が良い。この最適値は主蒸気１６の条件によって異なり、高温蒸気ではより高い値になると推定される。   Also in the case of the present embodiment, like the first embodiment described above, as shown schematically in FIG. 3, the specific enthalpy increase ratio has a predetermined optimum value that provides the best thermal efficiency. It is better that the specific enthalpy increase ratio is sufficiently higher than 1. This optimum value varies depending on the conditions of the main steam 16, and is estimated to be a higher value for high-temperature steam.

上述したように、第１の給水加熱器７における給水の比エンタルピ上昇と、第２の給水加熱器８における給水の比エンタルピ上昇の平均との比エンタルピ上昇比を１.９以上３.５以下にすることによって、サイクル熱効率が向上させることができる。   As described above, the specific enthalpy increase ratio between the specific enthalpy increase of the feed water in the first feed water heater 7 and the average of the specific enthalpy increase of the feed water in the second feed water heater 8 is 1.9 or more and 3.5 or less. By doing so, the cycle thermal efficiency can be improved.

第３の実施の形態
次に図１により本発明の第３の実施の形態について説明する。 Third Embodiment Next, a third embodiment of the present invention will be described with reference to FIG.

本実施の形態の蒸気タービンサイクルは、高圧タービン１と、再熱タービン２４と、ボイラ４と、高圧タービン１及び再熱タービン２４からの抽気蒸気によりボイラ４への給水を加熱する給水加熱器６と、給水ポンプ１２と、復水器１０とを備え、作動流体が水である１段再熱サイクルであり、かつ再生サイクルであるランキンサイクルとなっている。   The steam turbine cycle of the present embodiment includes a high-pressure turbine 1, a reheat turbine 24, a boiler 4, and a feed water heater 6 that heats the feed water to the boiler 4 using the extracted steam from the high pressure turbine 1 and the reheat turbine 24. And a water supply pump 12 and a condenser 10, a one-stage reheat cycle in which the working fluid is water, and a Rankine cycle that is a regeneration cycle.

また、ボイラ４出口の蒸気温度が５９０℃以上であり、高圧タービン排気抽気２２に対応した第１の給水加熱器７における給水温度上昇と、第１の給水加熱器７を除く給水加熱器における給水温度上昇の平均との温度上昇比が１.９以上３.５以下となっている。   Further, the steam temperature at the outlet of the boiler 4 is 590 ° C. or higher, the feed water temperature rise in the first feed water heater 7 corresponding to the high pressure turbine exhaust bleed air 22, and the feed water in the feed water heater excluding the first feed water heater 7 The temperature rise ratio with respect to the average temperature rise is 1.9 or more and 3.5 or less.

ここで、第１の給水加熱器７を除く給水加熱器とは、第１の給水加熱器７より給水が低圧である第２の給水加熱器８と、第１の給水加熱器７より給水が高圧である第３の給水加熱器２６を合わせた全てのものである。なお、第３の給水加熱器２６は高圧タービン１内部からの抽気蒸気により給水を加熱する。   Here, the feed water heaters excluding the first feed water heater 7 are the second feed water heater 8 whose feed water is lower in pressure than the first feed water heater 7 and the feed water from the first feed water heater 7. All of the high-pressure third feed water heaters 26 are combined. The third feed water heater 26 heats the feed water with the extracted steam from the inside of the high-pressure turbine 1.

各抽気蒸気２０の流量と、各抽気圧力位置３１を調節する事によって、第１の給水加熱器７、第２の給水加熱器８及び第３の給水加熱器２６における給水温度上昇を調整することができる。なお、高圧タービン排気抽気２２や中圧タービン排気抽気３２に関して温度変更する場合は、高圧タービン１と中圧タービン２の排気仕様を変更する事になる。   Adjusting the feed water temperature rise in the first feed water heater 7, the second feed water heater 8, and the third feed water heater 26 by adjusting the flow rate of each extraction steam 20 and each extraction pressure position 31 Can do. Note that when the temperature is changed with respect to the high-pressure turbine exhaust bleed air 22 and the intermediate-pressure turbine exhaust bleed air 32, the exhaust specifications of the high-pressure turbine 1 and the intermediate-pressure turbine 2 are changed.

ボイラ４入口の給水温度はボイラ４側で規定される事が多いので、値を固定して最適化計算をしたところ、１.９以上３.５以下の温度上昇比条件の時、サイクル熱効率は最大になった。   Since the boiler 4 inlet water supply temperature is often specified on the boiler 4 side, the optimization calculation was performed with the value fixed. When the temperature rise ratio was 1.9 to 3.5, the cycle thermal efficiency was It became the maximum.

第１の給水加熱器７における給水温度上昇と、第１の給水加熱器７を除く給水加熱器８，２６における給水温度上昇の平均との温度上昇比には、給水加熱器６の個数や、排気損失等といった蒸気タービンの機械上の差異や、発電プラントにおける発電出力に相当する規模、細かい構成の差異等の影響による範囲幅がある。   The temperature rise ratio between the feed water temperature rise in the first feed water heater 7 and the average feed water temperature rise in the feed water heaters 8 and 26 excluding the first feed water heater 7 includes the number of feed water heaters 6, There is a range of ranges due to the effects of differences in steam turbine machinery such as exhaust loss, the scale corresponding to the power generation output in the power plant, and the difference in fine structure.

上述したように、第１の給水加熱器７における給水温度上昇と、第１の給水加熱器７を除く給水加熱器８，２６における給水温度上昇の平均との温度上昇比を１.９以上３.５以下にすることによって、実施例１と同様にサイクル熱効率を向上させることができる。   As described above, the temperature increase ratio between the feed water temperature rise in the first feed water heater 7 and the average feed water temperature rise in the feed water heaters 8 and 26 excluding the first feed water heater 7 is 1.9 or more 3 By making it 0.5 or less, the cycle thermal efficiency can be improved as in the first embodiment.

第４の実施の形態
次に図１により本発明の第４の実施の形態について説明する。 Fourth Embodiment Next, a fourth embodiment of the present invention will be described with reference to FIG.

本実施の形態の蒸気タービンサイクルは、高圧タービン１と、再熱タービン２４と、ボイラ４と、高圧タービン１及び再熱タービン２４からの抽気蒸気によりボイラ４への給水を加熱する給水加熱器６と、給水ポンプ１２と、復水器１０とを備え、作動流体が水である１段再熱サイクルであり、かつ再生サイクルであるランキンサイクルを構成する。   The steam turbine cycle of the present embodiment includes a high-pressure turbine 1, a reheat turbine 24, a boiler 4, and a feed water heater 6 that heats the feed water to the boiler 4 using the extracted steam from the high pressure turbine 1 and the reheat turbine 24. And a water supply pump 12 and a condenser 10, and constitutes a Rankine cycle that is a one-stage reheat cycle in which the working fluid is water and a regeneration cycle.

ボイラ４出口の蒸気温度は５９０℃以上であり、高圧タービン排気抽気２２に対応した第１の給水加熱器７における給水の比エンタルピ上昇と、第１の給水加熱器７を除く給水加熱器８，２６における給水の比エンタルピ上昇の平均との比エンタルピ上昇比が１.９以上３.５以下となっている。   The steam temperature at the outlet of the boiler 4 is 590 ° C. or higher, the specific enthalpy increase of the feed water in the first feed water heater 7 corresponding to the high pressure turbine exhaust bleed air 22, and the feed water heaters 8 excluding the first feed water heater 7, The specific enthalpy increase ratio with the average of the specific enthalpy increase of the water supply at 26 is 1.9 or more and 3.5 or less.

各抽気蒸気２０の流量と、各抽気位置３１を調節する事によって、第１の給水加熱器７、第２の給水加熱器８及び第３の給水加熱器２６における給水の比エンタルピ上昇を調整することができる。なお、高圧タービン排気抽気２２や中圧タービン排気抽気３２に関して比エンタルピを変更する場合は、高圧タービン１と中圧タービン２の排気仕様を変更する事になる。   By adjusting the flow rate of each extraction steam 20 and each extraction position 31, the specific enthalpy increase of the feed water in the first feed water heater 7, the second feed water heater 8, and the third feed water heater 26 is adjusted. be able to. Note that when the specific enthalpy is changed with respect to the high-pressure turbine exhaust bleed 22 or the intermediate-pressure turbine exhaust bleed 32, the exhaust specifications of the high-pressure turbine 1 and the intermediate-pressure turbine 2 are changed.

ボイラ４入口の給水温度はボイラ４側で規定される事が多いので、値を固定して最適化計算をしたところ、１.９以上３.５以下の比エンタルピ上昇比条件の時、サイクル熱効率は最大になった。   Since the boiler 4 inlet water supply temperature is often specified on the boiler 4 side, optimization calculation was performed with a fixed value. When the specific enthalpy increase ratio was 1.9 or more and 3.5 or less, the cycle thermal efficiency Became the maximum.

第１の給水加熱器７における給水の比エンタルピ上昇と、第１の給水加熱器７を除く給水加熱器８，２６における給水の比エンタルピ上昇の平均との比エンタルピ上昇比には、給水加熱器６の個数や、排気損失等といった蒸気タービンの機械上の差異や、発電プラントにおける発電出力に相当する規模、細かい構成の差異等の影響による範囲幅がある。   The specific enthalpy increase ratio between the specific enthalpy increase in the feed water in the first feed water heater 7 and the average of the specific enthalpy increases in the feed water heaters 8 and 26 excluding the first feed water heater 7 includes the feed water heater There are range widths due to mechanical differences such as the number of 6 and exhaust loss, the scale corresponding to the power generation output in the power plant, and the difference in fine configuration.

上述のように、第１の給水加熱器７における給水の比エンタルピ上昇と、第１の給水加熱器７を除く給水加熱器８，２６における給水の比エンタルピ上昇の平均との比エンタルピ上昇比を１.９以上３.５以下とすることによって、第２の実施の形態と同様に、サイクル熱効率を向上させることができる。   As described above, the specific enthalpy increase ratio between the specific enthalpy increase of the feed water in the first feed water heater 7 and the average of the specific enthalpy increase of the feed water in the feed water heaters 8 and 26 excluding the first feed water heater 7 is By setting it to 1.9 or more and 3.5 or less, cycle thermal efficiency can be improved similarly to 2nd Embodiment.

第５の実施の形態
次に図１により本発明の第５の実施の形態について説明する。図１に示す第５の実施の形態は、第２の給水加熱器８における給水温度上昇が、給水ポンプ１２による給水の温度上昇を加えて算出されたものであり、他は第１の実施の形態と略同一である。 Fifth Embodiment Next, a fifth embodiment of the present invention will be described with reference to FIG. In the fifth embodiment shown in FIG. 1, the feed water temperature rise in the second feed water heater 8 is calculated by adding the feed water temperature rise by the feed water pump 12, and the others are the first embodiment. It is almost the same as the form.

給水ポンプ１２は給水を加熱してしまうため、給水が温度上昇する。このため、この温度上昇を加えて、第２の給水加熱器８の１台当たりの平均温度上昇を算出する。   Since the feed water pump 12 heats the feed water, the temperature of the feed water rises. For this reason, by adding this temperature rise, an average temperature rise per one second feed water heater 8 is calculated.

また、第３の実施の形態において、第１の給水加熱器７を除く給水加熱器８，２６における給水温度上昇を、給水ポンプ１２による給水の温度上昇を加えて算出することもできる。   Moreover, in 3rd Embodiment, the feed water temperature rise in feed water heaters 8 and 26 except the 1st feed water heater 7 can also be calculated by adding the rise in feed water temperature by the feed water pump 12.

やはり、給水ポンプ１２は給水を加熱してしまうため、給水が温度上昇する。このため、この温度上昇を加えて、第１の給水加熱器７を除く給水加熱器８，２６の１台当たりの平均温度上昇を算出する。   Again, since the feed water pump 12 heats the feed water, the temperature of the feed water rises. For this reason, this temperature rise is added and the average temperature rise per feed water heater 8, 26 excluding the first feed water heater 7 is calculated.

ボイラ４入口の給水温度はボイラ４側で規定される事が多いので、値を固定して最適化計算をしたところ、１.９以上３.５以下の温度上昇比条件の時、サイクル熱効率は最大になった。   Since the boiler 4 inlet water supply temperature is often specified on the boiler 4 side, the optimization calculation was performed with the value fixed. When the temperature rise ratio was 1.9 to 3.5, the cycle thermal efficiency was It became the maximum.

なお、第１の実施の形態にて記載した理由に加えて、給水ポンプ１２の機械上の差異による発熱差の影響もあるので、上述のように温度上昇比には範囲幅が生じてしまう。   In addition to the reason described in the first embodiment, there is also an influence of a heat generation difference due to a mechanical difference of the feed water pump 12, so that the range of the temperature rise ratio is generated as described above.

上述のように、第２の給水加熱器８における給水温度上昇を、給水ポンプ１２による給水の温度上昇を加えて算出したうえで、温度上昇比を規定することによって、第１の実施の形態及び第３の実施の形態と同様に、サイクル熱効率を向上させることができる。   As described above, by calculating the temperature increase of the feed water in the second feed water heater 8 by adding the temperature rise of the feed water by the feed water pump 12, by defining the temperature rise ratio, the first embodiment and As in the third embodiment, cycle thermal efficiency can be improved.

第６の実施の形態
次に図１により本発明の第６の実施の形態について説明する。図１に示す第６の実施の形態は、第２の給水加熱器８における給水の比エンタルピ上昇が、給水ポンプ１２による給水の比エンタルピ上昇を加えて算出されたものであり、他は第１の実施の形態と略同一である。 Sixth Embodiment Next, a sixth embodiment of the present invention will be described with reference to FIG. In the sixth embodiment shown in FIG. 1, the specific enthalpy increase of the feed water in the second feed water heater 8 is calculated by adding the specific enthalpy increase of the feed water by the feed water pump 12, and the others are the first. This is substantially the same as the embodiment.

給水ポンプ１２は給水を昇圧する物であり、さらに第３の実施の形態で記載したように給水を加熱してしまうため、給水の比エンタルピは上昇する。この比エンタルピ上昇を加えて、第２の給水加熱器８の１台当たりの平均比エンタルピ上昇を算出する。   The feed water pump 12 boosts the feed water and further heats the feed water as described in the third embodiment, so that the specific enthalpy of the feed water rises. By adding this specific enthalpy increase, the average specific enthalpy increase per one second feed water heater 8 is calculated.

また、本実施の形態は、前述の第４の実施の形態において、第１の給水加熱器７を除く給水加熱器８，２６における給水の比エンタルピ上昇を、給水ポンプ１２による給水の比エンタルピ上昇を加えて算出することもできる。   Further, in the present embodiment, the specific enthalpy increase of the feed water in the feed water heaters 8 and 26 excluding the first feed water heater 7 is increased in the specific enthalpy of the feed water by the feed pump 12 in the fourth embodiment. It is also possible to calculate by adding.

やはり、給水ポンプ１２は給水を昇圧する物であり、さらに前述の第３の実施の形態で記載したように給水を加熱してしまうため、給水の比エンタルピは上昇する。この比エンタルピ上昇を加えて、第１の給水加熱器７を除く給水加熱器８，２６の１台当たりの平均比エンタルピ上昇を算出する。   After all, the feed water pump 12 boosts the feed water, and further heats the feed water as described in the third embodiment, so that the specific enthalpy of the feed water rises. By adding this specific enthalpy increase, the average specific enthalpy increase per feed water heater 8, 26 excluding the first feed water heater 7 is calculated.

ボイラ４入口の給水温度はボイラ４側で規定される事が多いので、値を固定して最適化計算をしたところ、１.９以上３.５以下の比エンタルピ上昇比条件の時、サイクル熱効率は最大になった。   Since the boiler 4 inlet water supply temperature is often specified on the boiler 4 side, optimization calculation was performed with a fixed value. When the specific enthalpy increase ratio was 1.9 or more and 3.5 or less, the cycle thermal efficiency Became the maximum.

本実施の形態は、前述した第１の実施の形態で記載した理由に加えて、給水ポンプ１２の機械上の差異による発熱差の影響もあるので、上述の方に比エンタルピ上昇比には範囲幅が生じてしまう。   In addition to the reason described in the first embodiment described above, the present embodiment is also affected by the difference in heat generation due to the mechanical difference of the feed water pump 12, so the specific enthalpy increase ratio is in the above range. A width will occur.

上述のように、給水の比エンタルピ上昇を、給水ポンプ１２による給水の比エンタルピ上昇を加えて算出した上で、温度上昇比を規定することによって、第２の実施の形態及び第４の実施の形態と同様に、サイクル熱効率を向上させることができる。   As described above, the specific enthalpy increase of the water supply is calculated by adding the specific enthalpy increase of the water supply by the water supply pump 12, and then the temperature increase ratio is defined to thereby determine the second embodiment and the fourth embodiment. Similar to the embodiment, the cycle thermal efficiency can be improved.

第７の実施の形態
次に図１により本発明の第７の実施の形態について説明する。 Seventh Embodiment Next, a seventh embodiment of the present invention will be described with reference to FIG.

第１の実施の形態、第３の実施の形態及び第５の実施の形態において、給水加熱器６の総数を８個とし、温度上昇比が１.９以上３.５以下となるサイクル構成にする。経済性を考慮すると大型火力発電所では給水加熱器６の個数は８個が良いとされているためである。   In the first embodiment, the third embodiment, and the fifth embodiment, the total number of feed water heaters 6 is 8, and the temperature rise ratio is 1.9 or more and 3.5 or less. To do. This is because it is considered that the number of the feed water heaters 6 is good in large thermal power plants in consideration of economy.

図１では中圧タービン２からの抽気は排気を含めて２箇所、低圧タービン３からの抽気は４箇所だが、合計６箇所であればどちらが何箇所でもよい。   In FIG. 1, the extraction from the intermediate pressure turbine 2 includes two places including exhaust, and the extraction from the low pressure turbine 3 includes four places.

図１では脱気器９への抽気は中圧タービン排気抽気３２だが、そうでなくてもよい。給水加熱器６を８個と限定して最適化計算したところ、１.９以上３.５以下の温度上昇比条件の時、サイクル熱効率が最大になった。   In FIG. 1, the bleed air to the deaerator 9 is the medium pressure turbine exhaust bleed air 32, but this need not be the case. When optimization calculation was performed by limiting the number of feed water heaters 6 to 8, the cycle thermal efficiency was maximized when the temperature increase ratio was 1.9 or more and 3.5 or less.

上述のように、第１の実施の形態、第３の実施の形態及び第５の実施の形態において、給水加熱器６の総数を８個とし、温度上昇比を１.９以上３.５以下となるようなサイクル構成にすることによって、第１の実施の形態、第３の実施の形態及び第５の実施の形態と同様に、サイクル熱効率を向上させることができる。   As described above, in the first embodiment, the third embodiment, and the fifth embodiment, the total number of the feed water heaters 6 is 8, and the temperature increase ratio is 1.9 or more and 3.5 or less. By adopting such a cycle configuration, cycle thermal efficiency can be improved as in the first embodiment, the third embodiment, and the fifth embodiment.

第８の実施の形態
次に図１により本発明の第８の実施の形態について説明する。 Eighth Embodiment Next, an eighth embodiment of the present invention will be described with reference to FIG.

第２の実施の形態、第４の実施の形態及び第６の実施の形態において、給水加熱器６の総数を８個とし、比エンタルピ上昇比が１.９以上３.５以下となるサイクル構成にする。経済性を考慮すると大型火力発電所では給水加熱器６の個数は８個が良いとされているためである。   In the second embodiment, the fourth embodiment, and the sixth embodiment, the total number of the feed water heaters 6 is eight, and the specific enthalpy increase ratio is 1.9 or more and 3.5 or less. To. This is because it is considered that the number of the feed water heaters 6 is good in large thermal power plants in consideration of economy.

図１では中圧タービン２からの抽気は排気を含めて２箇所、低圧タービン３からの抽気は４箇所だが、合計６箇所であればどちらが何箇所でもよい。   In FIG. 1, the extraction from the intermediate pressure turbine 2 includes two places including exhaust, and the extraction from the low pressure turbine 3 includes four places.

図１では脱気器９への抽気は中圧タービン排気抽気３２だが、そうでなくてもよい。給水加熱器６が８個と限定して最適化計算したところ、１.９以上３.５以下の比エンタルピ上昇比条件の時、サイクル熱効率は最大になった。   In FIG. 1, the bleed air to the deaerator 9 is the medium pressure turbine exhaust bleed air 32, but this need not be the case. When optimization calculation was performed with the number of feed water heaters 6 limited to 8, the cycle thermal efficiency was maximized when the specific enthalpy increase ratio was 1.9 or more and 3.5 or less.

上述のように、第２の実施の形態、第４の実施の形態及び第６の実施の形態において、給水加熱器６の総数を８個とし、比エンタルピ上昇比を１.９以上３.５以下となるようなサイクル構成にすることによって、第２の実施の形態、第４の実施の形態及び第６の実施の形態と同様に、サイクル熱効率を向上させることができる。   As described above, in the second embodiment, the fourth embodiment, and the sixth embodiment, the total number of feed water heaters 6 is 8, and the specific enthalpy increase ratio is 1.9 or more and 3.5. By adopting a cycle configuration as described below, cycle thermal efficiency can be improved in the same manner as in the second, fourth, and sixth embodiments.

第９の実施の形態
次に図２により本発明の第９の実施の形態について説明する。図２において、図１に示した部分と同一部分には同一符号を付して詳細な説明は省略する。 Ninth Embodiment Next, a ninth embodiment of the present invention will be described with reference to FIG. In FIG. 2, the same parts as those shown in FIG.

本実施の形態は、前述した第１の実施の形態、第３の実施の形態及び第５の実施の形態における、給水加熱器６の総数を８個から１個増やして９個とし、温度上昇比が１．９以上かつ３.５以下となるサイクル構成にする。経済性を考慮すると大型火力発電所では給水加熱器６の個数は８個が良いとされているが、高効率化、高出力化、主蒸気高温化が進むにつれ、９個が良いという場合もあるためである。   In the present embodiment, the total number of feed water heaters 6 in the first embodiment, the third embodiment, and the fifth embodiment described above is increased from 8 to 9 to increase the temperature. The cycle configuration is such that the ratio is 1.9 or more and 3.5 or less. Considering economy, it is said that the number of feed water heaters 6 is good in large thermal power plants, but there are cases where nine are good as efficiency, power output, and main steam temperature increase. Because there is.

図２では中圧タービン２からの抽気は排気を含めて３箇所、低圧タービン３からの抽気は４箇所だが、合計７箇所であればどちらが何箇所でもよい。   In FIG. 2, the extraction from the intermediate pressure turbine 2 includes three places including exhaust, and the extraction from the low pressure turbine 3 includes four places.

図２では脱気器９への抽気は中圧タービン排気３２だが、そうでなくてもよい。給水加熱器６が９個と限定して最適化計算したところ、１.９以上３.５以下の温度上昇比条件の時、サイクル熱効率は最大になった。   In FIG. 2, the bleed air to the deaerator 9 is the intermediate pressure turbine exhaust 32, but this need not be the case. When optimization calculation was performed by limiting the number of feed water heaters to nine, the cycle thermal efficiency was maximized when the temperature increase ratio was 1.9 or more and 3.5 or less.

上述のように、本実施の形態は、第１の実施の形態、第３の実施の形態及び第５の実施の形態における、給水加熱器６の総数を８個から１個増やして９個とし、温度上昇比が１．９以上かつ３.５以下となるようなサイクル構成にすることによって、第１の実施の形態、第３の実施の形態及び第５の実施の形態と同様に、サイクル熱効率を向上させることができる。   As described above, in the present embodiment, the total number of feed water heaters 6 in the first embodiment, the third embodiment, and the fifth embodiment is increased from eight to nine to nine. As in the first embodiment, the third embodiment, and the fifth embodiment, the cycle configuration is such that the temperature rise ratio is 1.9 or more and 3.5 or less. Thermal efficiency can be improved.

第１０の実施の形態
次に図２により本発明の第１０の実施の形態について説明する。 Tenth Embodiment Next, a tenth embodiment of the present invention will be described with reference to FIG.

本実施の形態は、前述した第２の実施の形態、第４の実施の形態及び第６の実施の形態における、給水加熱器６の総数を８個から１個増やして９個とし、温度上昇比が１.９以上３.５以下となるサイクル構成にする。経済性を考慮すると大型火力発電所では給水加熱器６の個数は８個が良いとされているが、高効率化、高出力化、主蒸気高温化が進むにつれ、９個が良いという場合もあるためである。   In this embodiment, the total number of feed water heaters 6 in the second embodiment, the fourth embodiment, and the sixth embodiment described above is increased from 8 to 9 to increase the temperature. The cycle configuration is such that the ratio is 1.9 or more and 3.5 or less. Considering economy, it is said that the number of feed water heaters 6 is good in large thermal power plants, but there are cases where nine are good as efficiency, power output, and main steam temperature increase. Because there is.

図２では中圧タービン２からの抽気は排気を含めて３箇所、低圧タービン３からの抽気は４箇所だが、合計７箇所であればどちらが何箇所でもよい。   In FIG. 2, the extraction from the intermediate pressure turbine 2 includes three places including exhaust, and the extraction from the low pressure turbine 3 includes four places.

図２では脱気器９への抽気は中圧タービン排気３２だが、そうでなくてもよい。給水加熱器６が９個と限定して、最適化計算したところ、１.９以上３.５以下の比エンタルピ上昇比条件の時、サイクル熱効率は最大になった。   In FIG. 2, the bleed air to the deaerator 9 is the intermediate pressure turbine exhaust 32, but this need not be the case. When optimization calculation was performed with the feed water heater 6 limited to nine, the cycle thermal efficiency was maximized when the specific enthalpy increase ratio was 1.9 or more and 3.5 or less.

上述したように第２の実施の形態、第４の実施の形態及び第６の実施の形態における、給水加熱器６の総数を８個から１個増やして９個とし、比エンタルピ上昇比を１.９以上３.５以下となるようなサイクル構成にすることによって、第２の実施の形態、第４の実施の形態及び第６の実施の形態と同様に、サイクル熱効率を向上させることができる。   As described above, the total number of feed water heaters 6 in the second embodiment, the fourth embodiment, and the sixth embodiment is increased from 8 to 9, and the specific enthalpy increase ratio is 1. By adopting a cycle configuration such that the ratio is 9.9 or more and 3.5 or less, the cycle thermal efficiency can be improved in the same manner as the second embodiment, the fourth embodiment, and the sixth embodiment. .

第１１の実施の形態
次に図１及び図２により本発明の第１１の実施の形態について説明する。 Eleventh Embodiment Next, an eleventh embodiment of the present invention will be described with reference to FIGS.

第１乃至１０の実施の形態において、ボイラ４出口の蒸気温度が６００℃以上であるサイクル構成にする。主蒸気１６が６００℃以上の場合、より効果が顕著になるためである。主蒸気１６温度の高温化によるサイクル熱効率の向上効果が、抽気蒸気２０の条件設定により損なわず生かされる。   In 1st thru | or 10th Embodiment, it is set as the cycle structure whose steam temperature of the boiler 4 exit is 600 degreeC or more. This is because the effect becomes more remarkable when the main steam 16 is 600 ° C. or higher. The effect of improving the cycle thermal efficiency by increasing the temperature of the main steam 16 is utilized without being impaired by the condition setting of the extraction steam 20.

第１乃至１０の実施の形態において、ボイラ４出口の蒸気温度が６００℃以上であるサイクル構成にすることによって、第１乃至１０の実施の形態と同様に、サイクル熱効率を向上させることができる。   In the first to tenth embodiments, by adopting a cycle configuration in which the steam temperature at the outlet of the boiler 4 is 600 ° C. or higher, the cycle thermal efficiency can be improved as in the first to tenth embodiments.

本発明による蒸気タービンサイクルの第１乃至８の実施の形態及び第１１の実施の形態と、従来技術を示す概略図。The schematic which shows 1st-8th Embodiment of the steam turbine cycle by this invention, 11th Embodiment, and a prior art. 本発明による蒸気タービンサイクルの第９乃至１１の実施の形態を示す概略図。Schematic which shows 9th thru | or 11th embodiment of the steam turbine cycle by this invention. 温度上昇比と熱効率の関係を示した概略図。Schematic showing the relationship between temperature rise ratio and thermal efficiency.

符号の説明Explanation of symbols

１ 高圧タービン
２ 中圧タービン
３ 低圧タービン
４ ボイラ
５ 再熱器
６ 給水加熱器
７ 第１の給水加熱器
８ 第２の給水加熱器
９ 脱気器
１０ 復水器
１１ 復水ポンプ
１２ 給水ポンプ
１３ ドレン水ポンプ
１４ ボイラ給水
１５ ドレン水
１６ 主蒸気
１７ 再熱蒸気
１８ 発電機
１９ 回転軸
２０ 抽気蒸気
２１ 高圧タービン排気
２２ 高圧タービン排気抽気
２３ 海水または大気
２４ 再熱タービン
２５ 復水
２６ 第３の給水加熱器
３１ 抽気位置
３２ 中圧タービン排気抽気 DESCRIPTION OF SYMBOLS 1 High pressure turbine 2 Medium pressure turbine 3 Low pressure turbine 4 Boiler 5 Reheater 6 Feed water heater 7 1st feed water heater 8 2nd feed water heater 9 Deaerator 10 Condenser 11 Condensate pump 12 Feed pump 13 Drain water pump 14 Boiler feed water 15 Drain water 16 Main steam 17 Reheat steam 18 Generator 19 Rotating shaft 20 Extraction steam 21 High pressure turbine exhaust 22 High pressure turbine exhaust extraction 23 Seawater or atmosphere 24 Reheat turbine 25 Condensate 26 Third water supply Heater 31 Extraction position 32 Medium pressure turbine exhaust extraction

Claims (11)

高圧タービンと、再熱タービンと、ボイラと、前記タービンからの抽気蒸気により前記ボイラへの給水を加熱する給水加熱器と、給水ポンプと、復水器とを備え、作動流体が水である１段再熱サイクルであり、かつ再生サイクルであるランキンサイクルにおいて、
前記ボイラ出口の蒸気温度は、５９０℃以上であり、
前記高圧タービンの排気からの抽気に対応した第１の給水加熱器における給水温度上昇と、前記第１の給水加熱器より給水が低圧である第２の給水加熱器における給水温度上昇の平均との、温度上昇比が１.９以上３.５以下であることを特徴とした蒸気タービンサイクル。 A high-pressure turbine, a reheat turbine, a boiler, a feed water heater that heats feed water to the boiler by extracted steam from the turbine, a feed water pump, and a condenser, and the working fluid is water 1 In the Rankine cycle, which is a stage reheat cycle and a regeneration cycle,
The steam temperature at the boiler outlet is 590 ° C or higher,
A rise in feed water temperature in the first feed water heater corresponding to extraction from the exhaust of the high pressure turbine, and an average feed water temperature rise in the second feed water heater whose feed water is at a lower pressure than the first feed water heater. A steam turbine cycle characterized in that the temperature rise ratio is 1.9 or more and 3.5 or less. 高圧タービンと、再熱タービンと、ボイラと、前記タービンからの抽気蒸気により前記ボイラへの給水を加熱する給水加熱器と、給水ポンプと、復水器とを備え、作動流体が水である１段再熱サイクルであり、かつ再生サイクルであるランキンサイクルにおいて、
前記ボイラ出口の蒸気温度は、５９０℃以上であり、
前記高圧タービンの排気からの抽気に対応した第１の給水加熱器における給水の比エンタルピ上昇と、前記第１の給水加熱器より給水が低圧である第２の給水加熱器における給水の比エンタルピ上昇の平均との、比エンタルピ上昇比が１.９以上３.５以下であることを特徴とした蒸気タービンサイクル。 A high-pressure turbine, a reheat turbine, a boiler, a feed water heater that heats feed water to the boiler by extracted steam from the turbine, a feed water pump, and a condenser, and the working fluid is water 1 In the Rankine cycle, which is a stage reheat cycle and a regeneration cycle,
The steam temperature at the boiler outlet is 590 ° C or higher,
Increase in specific enthalpy of feed water in the first feed water heater corresponding to extraction from the exhaust of the high pressure turbine, and rise in specific enthalpy of feed water in the second feed water heater whose feed water is at a lower pressure than the first feed water heater A steam turbine cycle having a specific enthalpy increase ratio of 1.9 to 3.5 with respect to an average of 高圧タービンと、再熱タービンと、ボイラと、前記タービンからの抽気蒸気により前記ボイラへの給水を加熱する給水加熱器と、給水ポンプと、復水器とを備え、作動流体が水である１段再熱サイクルであり、かつ再生サイクルであるランキンサイクルにおいて、
前記ボイラ出口の蒸気温度は、５９０℃以上であり、
前記高圧タービンの排気からの抽気に対応した第１の給水加熱器における給水温度上昇と、前記第１の給水加熱器を除く給水加熱器における給水温度上昇の平均との、温度上昇比が１.９以上３.５以下であることを特徴とした蒸気タービンサイクル。 A high-pressure turbine, a reheat turbine, a boiler, a feed water heater that heats feed water to the boiler by extracted steam from the turbine, a feed water pump, and a condenser, and the working fluid is water 1 In the Rankine cycle, which is a stage reheat cycle and a regeneration cycle,
The steam temperature at the boiler outlet is 590 ° C or higher,
A temperature increase ratio between the feed water temperature rise in the first feed water heater corresponding to the extraction from the exhaust of the high-pressure turbine and the average feed water temperature rise in the feed water heater excluding the first feed water heater is 1. A steam turbine cycle characterized by being 9 or more and 3.5 or less. 高圧タービンと、再熱タービンと、ボイラと、前記タービンからの抽気蒸気により前記ボイラへの給水を加熱する給水加熱器と、給水ポンプと、復水器とを備え、作動流体が水である１段再熱サイクルであり、かつ再生サイクルであるランキンサイクルにおいて、
前記ボイラ出口の蒸気温度が５９０℃以上であり、
前記高圧タービンの排気からの抽気に対応した第１の給水加熱器における給水の比エンタルピ上昇と、前記第１の給水加熱器を除く給水加熱器における給水の比エンタルピ上昇の平均との、比エンタルピ上昇比が１.９以上３.５以下であることを特徴とした蒸気タービンサイクル。 A high-pressure turbine, a reheat turbine, a boiler, a feed water heater that heats feed water to the boiler by extracted steam from the turbine, a feed water pump, and a condenser, and the working fluid is water 1 In the Rankine cycle, which is a stage reheat cycle and a regeneration cycle,
The steam temperature at the boiler outlet is 590 ° C or higher,
The specific enthalpy of the specific enthalpy increase of the feed water in the first feed water heater corresponding to the extraction from the exhaust of the high-pressure turbine and the average of the increase in the specific enthalpy of the feed water in the feed water heater excluding the first feed water heater A steam turbine cycle characterized in that the increase ratio is 1.9 or more and 3.5 or less. 第２の給水加熱器における給水温度上昇は、給水ポンプによる給水の温度上昇を加えて算出されることを特徴とした請求項１又は３のいずれかに記載の蒸気タービンサイクル。   The steam turbine cycle according to claim 1, wherein the feed water temperature rise in the second feed water heater is calculated by adding the feed water temperature rise by the feed pump. 第２の給水加熱器における給水の比エンタルピ上昇は、給水ポンプによる給水の比エンタルピ上昇を加えて算出されることを特徴とした請求項２又は４のいずれかに記載の蒸気タービンサイクル。   5. The steam turbine cycle according to claim 2, wherein the specific enthalpy increase in the feed water in the second feed water heater is calculated by adding the specific enthalpy increase in the feed water by the feed water pump. 前記給水加熱器の総数は８個であり、
前記温度上昇比が１.９以上３.５以下であることを特徴とした請求項１、３又は５のいずれかに記載の蒸気タービンサイクル。 The total number of the water heaters is 8,
The steam turbine cycle according to any one of claims 1, 3, and 5, wherein the temperature increase ratio is 1.9 or more and 3.5 or less. 前記給水加熱器の総数は８個であり、
前記比エンタルピ上昇比が１.９以上３.５以下であることを特徴とした請求項２、４又は６のいずれかに記載の蒸気タービンサイクル。 The total number of the water heaters is 8,
The steam turbine cycle according to claim 2, wherein the specific enthalpy increase ratio is 1.9 or more and 3.5 or less. 前記給水加熱器の総数は９個であり、
前記温度上昇比が１.９以上３.５以下であることを特徴とした請求項１、３又は５のいずれかに記載の蒸気タービンサイクル。 The total number of water heaters is nine;
The steam turbine cycle according to any one of claims 1, 3, and 5, wherein the temperature increase ratio is 1.9 or more and 3.5 or less. 前記給水加熱器の総数は９個であり、
前記比エンタルピ上昇比が１.９以上３.５以下であることを特徴とした請求項２、４又は６のいずれかに記載の蒸気タービンサイクル。 The total number of water heaters is nine;
The steam turbine cycle according to claim 2, wherein the specific enthalpy increase ratio is 1.9 or more and 3.5 or less. 前記ボイラ出口の蒸気温度は、６００℃以上であることを特徴とした請求項１乃至１０のいずれか１項に記載の蒸気タービンサイクル。   The steam turbine cycle according to any one of claims 1 to 10, wherein the steam temperature at the boiler outlet is 600 ° C or higher.
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