WO2011055666A1 - Axial flow turbine - Google Patents

Axial flow turbine Download PDF

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Publication number
WO2011055666A1
WO2011055666A1 PCT/JP2010/069086 JP2010069086W WO2011055666A1 WO 2011055666 A1 WO2011055666 A1 WO 2011055666A1 JP 2010069086 W JP2010069086 W JP 2010069086W WO 2011055666 A1 WO2011055666 A1 WO 2011055666A1
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WO
WIPO (PCT)
Prior art keywords
turbine
stage
flow
working fluid
axial flow
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Application number
PCT/JP2010/069086
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French (fr)
Japanese (ja)
Inventor
茂樹 妹尾
英樹 小野
健 工藤
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株式会社日立製作所
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Publication of WO2011055666A1 publication Critical patent/WO2011055666A1/en

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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01D—NON-POSITIVE DISPLACEMENT MACHINES OR ENGINES, e.g. STEAM TURBINES
    • F01D11/00—Preventing or minimising internal leakage of working-fluid, e.g. between stages
    • F01D11/001—Preventing or minimising internal leakage of working-fluid, e.g. between stages for sealing space between stator blade and rotor
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01D—NON-POSITIVE DISPLACEMENT MACHINES OR ENGINES, e.g. STEAM TURBINES
    • F01D5/00—Blades; Blade-carrying members; Heating, heat-insulating, cooling or antivibration means on the blades or the members
    • F01D5/12—Blades
    • F01D5/14—Form or construction
    • F01D5/141—Shape, i.e. outer, aerodynamic form
    • F01D5/145—Means for influencing boundary layers or secondary circulations
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01D—NON-POSITIVE DISPLACEMENT MACHINES OR ENGINES, e.g. STEAM TURBINES
    • F01D9/00—Stators
    • F01D9/02—Nozzles; Nozzle boxes; Stator blades; Guide conduits, e.g. individual nozzles
    • F01D9/04—Nozzles; Nozzle boxes; Stator blades; Guide conduits, e.g. individual nozzles forming ring or sector
    • F01D9/041—Nozzles; Nozzle boxes; Stator blades; Guide conduits, e.g. individual nozzles forming ring or sector using blades

Definitions

  • the present invention relates to an axial flow turbine such as a steam turbine and a gas turbine.
  • Patent Document 1 a plurality of turbine stages each including a stationary blade fixed to an outer peripheral side diaphragm and an inner peripheral side diaphragm and a moving blade fixed to a turbine rotor rotating around a central axis of the turbine are provided in the working fluid flow path.
  • an axial flow turbine having a function of converting kinetic energy generated when a high pressure working fluid expands toward a low pressure portion in the flow path into a rotational force by a turbine stage composed of a stationary blade and a moving blade.
  • the ring area is the product of the blade length and the mean diameter obtained by adding the outer peripheral end diameter and the inner peripheral end diameter of the blade and dividing by 2 Therefore, in order to increase the ring area, it is practiced to increase the wing length and the average diameter.
  • the specific total enthalpy H0 which is the sum of the enthalpy (specific enthalpy) per unit mass of working fluid and the kinetic energy per unit mass divided by 2 of the flow velocity at the turbine stage inlet, is the rotational axis
  • the value is substantially constant from the inner circumferential side to the outer circumferential side close to
  • the specific enthalpy h1 between the stationary blade and the moving blade becomes larger toward the outer peripheral side than the inner peripheral side so as to be balanced with the swirling flow between the stationary and moving blades. Therefore, the specific enthalpy difference H0-h1 decreases toward the outer periphery.
  • the velocity of the flow out of the vane is proportional to the square root of this specific enthalpy difference H0-h1. That is, the vane outflow velocity decreases toward the outer periphery.
  • the first is that the relative inflow Mach number of the moving blade becomes supersonic, and the possibility of loss increases.
  • the peripheral speed which is the rotational speed of the moving blades.
  • the circumferential velocity of the moving blade is the largest at the outer peripheral end where the radial position is the largest, that is, at the blade tip. If the circumferential speed Mach number obtained by dividing the circumferential speed at the tip by the speed of sound exceeds 1 and becomes supersonic, the relative velocity of the stream flowing into the moving blade will be reduced if the rotational direction component of the flow from the stationary blade is insufficient. It becomes supersonic speed.
  • the second is that the possibility of the occurrence of peeling increases in the expanded flow passage portion on the outer peripheral side.
  • the rate of increase of the flow path height at the stage outlet with respect to the expansion rate of the meridional plane flow path, that is, the flow path height at the inlet of the turbine stage increases.
  • the axial length of the stage can not generally be increased because the length of the entire turbine is limited.
  • the increase is realized by increasing the spread angle of the shape of the meridional flow passage at the outer peripheral end or the inner peripheral end of the stationary blade portion.
  • an object of the present invention is to provide an axial flow turbine capable of suppressing the shock wave loss due to the increase of the annular area and the loss due to separation, and improving the turbine efficiency.
  • an axial flow turbine including a plurality of turbine stages each including a stationary blade fixed to a stationary body and a moving blade fixed to a turbine rotor in a working fluid flow path
  • a turbine is provided, and a part of the working fluid flowing in from the working fluid flow direction upstream side is bypassed to at least one outer peripheral side of the turbine stage and introduced to the moving blades downstream of the bypassed turbine paragraph working fluid flow direction
  • a paragraph bypass flow path is provided.
  • the present invention in the axial flow turbine, it is possible to suppress the shock wave loss due to the increase in the annular area and the loss due to the separation, and to improve the turbine efficiency.
  • FIG. 7 is a diagram schematically illustrating the relationship between the stator blade outflow velocity, the rotor circumferential velocity, and the relative inflow velocity of the moving blades when the circumferential velocity of the moving blades is high. It is a graph showing the blade height direction distribution of moving blade relative inflow Mach number in case the moving blade circumferential speed is large. It is a graph showing the specific enthalpy distribution of the wing height direction of the last paragraph part of the steam turbine concerning a 1st embodiment of the present invention.
  • FIG. 7 is a diagram schematically illustrating the relationship between the stator blade outflow velocity, the rotor circumferential velocity, and the relative inflow velocity of the moving blades when the circumferential velocity of the moving blades is high. It is a graph showing the blade height direction distribution of moving blade relative inflow Mach number in case the moving blade circumferential speed is large. It is a graph showing the specific enthalpy distribution of the wing height direction of the last paragraph part of the steam turbine concerning a 1st embodiment of
  • FIG. 3 is an explanatory view schematically showing the relationship among the stator blade outflow velocity, the moving blade peripheral velocity, and the moving blade relative inflow velocity of the steam turbine according to the first embodiment of the present invention. It is a meridional plane sectional view showing the important section structure of the turbine paragraph part of the steam turbine concerning a 1st embodiment of the present invention. It is a perspective view showing the principal part structure of the distribution plate of the steam turbine concerning a 1st embodiment of the present invention. It is a graph showing the specific enthalpy distribution of the last paragraph of the steam turbine concerning a 1st embodiment of the present invention, and the wing height direction of the one upper stream side paragraph.
  • FIG. 1 is a meridional cross-sectional view showing a main part structure of a turbine stage portion of a general steam turbine.
  • the high-pressure part P0 of the working fluid upstream of the working flow of the steam 1 (hereinafter referred to simply as upstream) and the downstream flow direction (hereinafter referred to simply as downstream) It is provided in the steam main flow path 2 between the steam low pressure portion P1 of
  • An outer peripheral diaphragm 4 is fixed to the inner peripheral side of the turbine casing 3.
  • the turbine stage comprises a stationary blade 6 fixed between the outer peripheral side diaphragm 4 and the inner peripheral side diaphragm 5 and a moving blade 7 fixed to the turbine rotor 8 so as to face the downstream side of the stationary blade 6.
  • the turbine stage is a multi-stage turbine including a plurality of stages
  • the turbine stage is provided by being repeated a plurality of times from the high pressure portion P0 to the low pressure portion P1 along the turbine center axis 9.
  • the paragraph closest to the steam inlet of the high pressure part P0 is called the first paragraph
  • the paragraph closest to the steam outlet of the low pressure part P1 is called the final paragraph.
  • the turbine rotor 8 is mechanically connected to a generator (not shown). Therefore, the steam turbine converts the kinetic energy generated when high-pressure steam expands toward the low pressure part into rotational force by the turbine stage composed of the stationary blades and the moving blades, and generates electrical energy from the generator via the turbine rotor. To generate electricity.
  • FIG. 2 is a graph showing the specific enthalpy distribution in the blade height direction of the final stage portion of the steam turbine illustrated in FIG.
  • the horizontal axis is the specific enthalpy, and the axial direction is determined so that the specific enthalpy increases as the working fluid flows from left to right in FIG.
  • the vertical axis represents the blade height direction
  • BH represents the blade outlet height.
  • H0 is the total specific enthalpy which is the sum of the enthalpy per unit mass at the stage inlet and the kinetic energy per unit mass divided by 2 of working fluid flow velocity divided by 2
  • h1 is the ratio between static and moving blades Enthalpy
  • h2 represents the specific enthalpy of the stage exit.
  • the total specific enthalpy H0 at the stage inlet is substantially constant in the wing height direction.
  • the specific enthalpy h1 between the stationary and moving blades becomes larger toward the outer peripheral side in the radial direction of the turbine (hereinafter, simply referred to as the outer peripheral side) so as to be balanced mainly with the centrifugal force due to the turning speed between the stationary blades.
  • the specific enthalpy difference ⁇ h applied to the outer peripheral side stationary blades decreases, and the stationary blade outflow velocity proportional to the square root of the specific enthalpy difference also decreases.
  • the specific enthalpy difference on the vane and the tendency of the vane outflow velocity to decrease are increased by the wing length and the average diameter of the outer peripheral end diameter and the inner peripheral end diameter of the wing divided by 2; The larger the annular area, the more prominent the outer peripheral end position of the wing becomes on the outer peripheral side.
  • FIG. 3 is a diagram schematically showing the relationship between the stator blade outflow velocity, the rotor peripheral velocity, and the relative inflow velocity of the moving blades when the peripheral velocity of the moving blades is large.
  • the high-pressure P3 steam 22 is accelerated and turned into a flow of velocity V by passing through the stationary blade 23 (hereinafter, this velocity is referred to as stationary blade outflow velocity V).
  • the specific enthalpy h1 between the stationary and moving blades is reduced to increase the specific enthalpy difference of the stationary blades in order to increase the vane outflow velocity V. It is necessary to increase ⁇ h.
  • the specific enthalpy between the static and moving blades increases toward the outer periphery due to the turning velocity field at the exit of the stationary blade, and the longer the blade length, the stronger the effect of the turning velocity field, making it difficult to reduce h1. . That is, as the blade length increases, it becomes more difficult to increase the specific enthalpy difference ⁇ h of the stator blades, and it becomes more difficult to increase the stator blade outflow velocity V.
  • FIG. 4 is a graph showing the blade height direction distribution of the rotor inflow relative Mach number when the rotor peripheral speed is high.
  • the Mach number exceeds 1.0 on the outer peripheral side of the wing, and supersonic inflow occurs.
  • the enthalpy difference H0-h2 itself of the turbine stage is increased in order to increase the specific enthalpy difference ⁇ h of the outer peripheral side stationary blades, the rotor inflow Mach number of the inner peripheral end exceeds 1.0 and the supersonic speed It is difficult to solve the problem of supersonic inflow by increasing the enthalpy difference throughout the paragraph to become the inflow.
  • FIG. 5 is a graph showing the specific enthalpy distribution in the blade height direction of the final stage of the steam turbine according to the first embodiment. Also in FIG. 5, the same shaft configuration as in FIG. 2 is used.
  • the specific enthalpy difference ⁇ h on the outer periphery side of the stationary blade is increased by increasing the total specific enthalpy H0 of the stage inlet portion toward the outer periphery side.
  • the stator blade outflow velocity V proportional to the square root of the specific enthalpy difference ⁇ h also increases. Therefore, as shown in FIG. 6, when the total specific enthalpy H0 at the stage inlet is constant, the stator blade outflow velocity, which is V1, can be increased to V2.
  • the velocity component in the swirling direction of the vane outflow velocity component V can also be increased from VT1 to VT2. As a result, it is possible to reduce the moving blade relative inflow velocity W1 flowing into the moving blade 24 to W2 despite the same circumferential velocity U.
  • the specific enthalpy difference ⁇ h of the stationary blade portion on the outer peripheral side of the blade becomes large, and the acceleration of the flow becomes large in the enlarged flow path portion of the outer peripheral end of the meridional flow path. It is possible to suppress the separation of the flow from the wing surface and the side wall surface constituting the main steam flow path wall, and to suppress the loss associated with the separation of the flow.
  • the present invention it is possible to suppress the shock wave loss due to the increase of the annular area and the loss due to the separation, and to improve the turbine stage efficiency.
  • FIG. 7 is a meridional cross-sectional view of the main part structure of the turbine stage portion of the steam turbine according to the first embodiment of the present invention.
  • a flow dividing plate 26 that divides the main steam flow passage 2 into two in the radial direction of the turbine is provided in the turbine stage portion one upstream side of the final stage.
  • the flow dividing plate 26 extends from the vanes 12 of the final stage to the vicinity of the blade 21 outlet of the turbine stage located two upstream sides of the final stage, and is supported by the vanes 12 constituting the final stage. It is done.
  • the flow dividing plate 26 is preferably provided on the inner peripheral side of the height where the inflow Mach number exceeds 1.0 as shown in FIG.
  • the upstream end of the flow dividing plate 26 is formed to gradually increase in thickness from the upstream side to the downstream side so as not to obstruct the flow of the working fluid.
  • the turbine stage one upstream side of the final stage is provided on the inner peripheral side of the flow dividing plate 26.
  • the stator vanes 27 are circumferentially arrayed on the inner peripheral side of the upstream end of the flow dividing plate 26, and the rotor vanes 28 are arranged downstream of the stator vanes 27 to configure a turbine stage one upstream side of the final paragraph doing.
  • the bypass flow passage 29 is formed on the outer peripheral side of the flow dividing plate 26, and the round flow passage 30 is formed on the inner peripheral side. Therefore, the outlet channel height of the moving blade 28 in the paragraph channel 30 is smaller than the outlet channel height of the rotor blade 21 in the upstream stage, and is compared with the outlet channel height of the rotor blade 13 in the downstream side Even small.
  • the flow dividing plate 26 may be warped by the weight of the vanes 27 provided on the inner circumferential side or by fluid force, or a large stress exceeding the allowable stress occurs in the stationary blade 12 fixing portion of the flow dividing plate 26
  • a plurality of cylindrical support members 31 extending from the outer peripheral side stationary portion such as the outer peripheral side diaphragm 4 may be installed in the circumferential direction to support the flow dividing plate 26.
  • FIG. 8 is a perspective view of the flow dividing plate 26 shown in FIG.
  • the flow dividing plate 26 is configured by connecting a plurality of plate members in the circumferential direction, and has a nearly conical shape whose inner diameter value gradually increases from the upstream side toward the downstream side. It is
  • the main steam discharged from the two upstream turbine stages in the final stage is divided into two in the radial direction of the turbine by the flow dividing plate 26, and the outer peripheral side of the flow dividing plate 26
  • the steam flow that has flowed into the bypass flow path 29 flows directly to the outer peripheral side of the stator vane 12 of the final stage.
  • the steam flow that has flowed into the stage flow passage 30 on the inner peripheral side of the flow dividing plate 26 flows only to the inner peripheral side of the final paragraph via the turbine stage on the upstream side of one final stage on the inner peripheral side of the flow dividing plate 26 Do.
  • FIG. 9 is a graph showing the specific enthalpy distribution in the blade height direction in the final stage of the steam turbine and the stage immediately upstream thereof according to the present embodiment.
  • H0 is the total specific enthalpy of the upstream side of the stationary blade 27 shown in FIG. This H0 becomes the inlet total specific enthalpy of the final paragraph on the outer peripheral side in the blade height direction from the diverting plate 26 position bh.
  • h1 is a specific enthalpy between the stationary blade 27 and the moving blade 28
  • H2 is a total inlet specific enthalpy of the inner peripheral side of the stationary blade 12
  • h3 is an inner peripheral side between the stationary blade 12 and the moving blade 13
  • the specific enthalpy, h5 is the specific enthalpy on the outer peripheral side between the stationary blade 12 and the moving blade 13
  • h5a is the specific enthalpy of the last stage of the general steam turbine shown in FIG. 1 corresponding to h5.
  • the specific enthalpy difference ⁇ h on the outer peripheral side of the stator blade can be increased by increasing the overall specific enthalpy of the stage inlet on the outer peripheral side.
  • FIG. 10 is a graph showing the distribution in the blade height direction of the relative inflowing Mach number of the moving blade in the final stage of the steam turbine according to the present embodiment shown in FIG. 7.
  • M1r shown by a solid line is the rotor inflow relative Mach number in the final stage of the steam turbine according to the present embodiment
  • M1ra is the rotor relative inflow Mach number in the general last stage of the steam turbine shown in FIG. Since the specific enthalpy difference ⁇ h of the stationary blades is increased on the outer peripheral side in the blade height direction from the flow dividing plate 26 position bh, the stationary blade outflow velocity on the outer peripheral side of the stationary blades increases, and as described using FIG. Sound velocity inflow can be avoided. Therefore, the efficiency of the turbine stage can be improved by the effect of reducing the moving blade relative inflow Mach number of the present invention.
  • FIG. 11 shows the blade height direction distribution of the moving blade inlet angle measured from the circumferential direction of the final stage of the steam turbine according to the present embodiment.
  • .Alpha..sub.in indicated by a solid line is a blade inlet angle
  • .beta..sub.in indicated by a dotted line is a relative inflow angle of steam to the blade.
  • the relative inflow angle also becomes smaller discontinuously in the portion where the outer entrance side specific inlet enthalpy becomes discontinuous, the inlet angle of the moving blade is continuous.
  • the difference between the inlet angle and the relative inflow angle is referred to as the incident angle, but if the incident angle is within a range of plus or minus, the sharp increase in wing loss does not occur.
  • the range is particularly large, for example, in the range of plus and minus 40 degrees of the incidence angle, a rapid increase in loss does not occur. Therefore, as shown in FIG. 11, the efficiency does not greatly deteriorate even if the blade inlet angle and the blade relative inflow angle do not match.
  • the specific enthalpy difference ⁇ h on the outer periphery side of the stationary blade is increased by increasing the total specific enthalpy H0 on the outer peripheral side of the final stage inlet portion.
  • the specific enthalpy difference ⁇ h on the outer peripheral side of the stationary blade becomes large, and the acceleration of the flow becomes large in the enlarged flow path portion of the outer peripheral end of the meridional flow path. Can suppress the separation from the wing surface and the side wall surface constituting the steam main flow path wall, and can suppress the loss accompanying the flow separation.
  • the rotational force extracted by the paragraph consisting of the stator vanes 27 and the moving blades 28 becomes smaller, the reduction thereof can be taken out as the rotational force at the paragraph consisting of the stationary blades 12 and the moving vanes 13 As it can, the torque of the entire turbine is not reduced. Rather, the rotational force can be increased by the amount of loss reduction.
  • FIG. 12 is a meridional cross-sectional view showing a main part structure of a turbine stage portion of a steam turbine according to a second embodiment of the present invention.
  • symbol is attached
  • the present embodiment is different from the first embodiment shown in FIG. 7 in that the main blade 2 of the steam in the final stage on the downstream side of the flow dividing plate 26 is divided into two on the inner circumferential side and the outer circumferential side.
  • the second flow dividing plate 32 to be divided is provided.
  • the second flow dividing plate 32 is fixed to each of the moving blades 13 arranged in a row in the circumferential direction, and is connected in contact with the second flow dividing plate 32 of the adjacent wing.
  • it can utilize also for providing a vibration damping mechanism, adjusting the vibration mode of a wing
  • FIG. 13 shows the blade height direction distribution of the moving blade inlet angle measured from the circumferential direction of the final stage portion of the steam turbine according to the present embodiment.
  • the provision of the second flow dividing plate 32 eliminates the need for the blades to be smoothly connected on the inner and outer peripheral sides of the flow dividing plate, so the inner and outer peripheral sides of the blade are independently provided. It is possible to design according to each flow, and as shown in FIG. 13, it is possible to match the blade inlet angle ⁇ in with the blade relative inflow angle ⁇ in of the steam in the final paragraph. Therefore, according to the second embodiment, in addition to the effects of the first embodiment, it is possible to suppress the loss caused by the blade inlet angle being discontinuous in the blade height direction. Further, the second flow dividing plate 32 can be expected to damp the vibration of the moving blade 13 and can be used to adjust the vibration mode of the blade to avoid resonance.
  • FIG. 14 is a meridional cross-sectional view showing a main part structure of a turbine stage portion of a steam turbine according to a third embodiment of the present invention.
  • symbol is attached
  • the present embodiment is an application example of the second embodiment shown in FIG. 12, and a seal structure 33 is provided between the first flow dividing plate 26 and the second flow dividing plate 32, and steam is transmitted between the flow dividing plates. It is suppressing that a leak flow arises.
  • FIG. 15 is an enlarged view of the seal structure 33 of the steam turbine shown in FIG.
  • the first flow dividing plate 26 has an extending portion 35 extending toward the moving blade 13 in the inter-static-moving blade channel 34 in the final paragraph, and seals the inner peripheral side of the extending portion 35 Fins 36 are provided.
  • the second flow dividing plate 32 also has an extension 37 that extends toward the stationary blade 12 on the inner peripheral side of the extension 35. The extension 35 and the extension 37 are radially juxtaposed to each other.
  • the seal structure 33 can be provided to suppress the occurrence of the leak flow of the steam between the diverting plates. There is no need to continue the pressure on the inner circumferential side 38 and the outer circumferential side 39 in the blade length direction, and the final paragraph can be designed completely independently on the inner circumferential side and the outer circumferential side.
  • FIG. 16 shows the specific enthalpy distribution in the final stage of the steam turbine of this embodiment and in the blade height direction in the upstream stage of one stage thereof.
  • the static / moving blade specific enthalpy in the final paragraph is discontinuous at the connection portion between the inner circumferential side h3 and the outer circumferential side h5.
  • or 3 performs paragraph design on the conditions that the flow volume per unit area is equal at the last paragraph exit, and has shown the characteristic.
  • the present embodiment is an application example of the first embodiment, and the basic structure is the same as the first embodiment shown in FIG.
  • the blade inlet angle can be continuous.
  • the flow rate per unit area of the outer circumferential side flow passage 42 (bypass flow passage 29) of the final paragraph is set to the flow rate per unit area of the inner circumference side flow passage 41 (paragraph flow passage 30) of the final paragraph.
  • the blade inlet angle was made continuous by setting it small.
  • the flow rate per unit area of the turbine paragraph in the paragraph flow path 30 of FIG. In order to change, the position of the upstream end of the flow dividing plate 26 is determined so that the flow ratio between the inner circumference and the outer circumference of the final paragraph and the flow ratio flowing through the paragraph flow passage 30 and the bypass flow passage 29 become equal.
  • FIG. 18 is a meridional cross-sectional view showing a main part structure of a turbine stage portion of a steam turbine according to a fifth embodiment of the present invention.
  • the same reference numerals as in FIG. 7 denote the same parts.
  • a plurality of flow dividing plates similar to those of the first embodiment are provided in the steam main flow passage 2 in the radial direction of the turbine, and the steam main flow passage 2 is divided into a plurality of blocks 46, 47, 48, 49 in the blade height direction.
  • the bypass steam from the upstream turbine stage is made to flow into the moving blades 13 toward the outer peripheral side.
  • the block 49 on the innermost circumferential side is formed between the diverting plate 45 and the inner circumferential side wall surface of the main steam passage 2.
  • the flow dividing plate 45 is supported by the vanes 13 of the final stage.
  • a stationary blade 50 and a moving blade 51 that constitute a turbine stage one upstream side of the final stage are provided on the inner peripheral side of the upstream end portion of the flow dividing plate 45.
  • the vane 50 is fixed between the diverting plate 45 and the inner diaphragm 53.
  • the flow dividing plate 45 is supported by a support member 31 extending from an outer peripheral side fixing member 52 constituting an outer peripheral side wall surface of the steam main flow channel 2.
  • a shroud member 54 is provided at the outer peripheral tip of the moving blade 51.
  • one block 48 on the outer peripheral side of the block 49 is formed between the diverter plate 45 and the diverter plate 44 on the outer peripheral side of the diverter plate 45.
  • the flow dividing plate 44 is supported by the vanes 13.
  • On the inner peripheral side of the upstream end portion of the flow dividing plate 44 there are provided a vane 55 and a moving blade 56 which constitute a two upstream turbine stage of the final stage.
  • the vane 55 is fixed between the diverting plate 44 and the inner diaphragm 57.
  • the flow dividing plate 44 is supported by a support member 31 extending from an outer peripheral fixing member 58 which constitutes an outer peripheral side wall surface of the steam main flow channel 2.
  • a shroud member 59 is also provided at the outer circumferential tip of the moving blade 56.
  • one upstream block 47 of the block 48 is formed between the diverter plate 44 and the diverter plate 43 on the outer peripheral side of the diverter plate 44.
  • the flow dividing plate 43 is supported by the vanes 13.
  • On the inner peripheral side of the upstream end portion of the flow dividing plate 43 there are provided a stator blade 60 and a rotor blade 61 which constitute three upstream turbine stages of the final paragraph.
  • the vane 60 is fixed between the diverting plate 43 and the inner diaphragm 62.
  • the flow dividing plate 43 is supported by a support member 31 extending from an outer peripheral side fixing member 63 which constitutes an outer peripheral side wall surface of the steam main flow channel 2.
  • a shroud member 67 is also provided at the outer circumferential tip of the moving blade 61.
  • the block 46 on the outermost side is formed between the flow dividing plate 43 and the outer peripheral side wall surface of the main steam channel 2.
  • the vapor bypassed from the upstream side flows into the final paragraph as it goes to the outer peripheral side.
  • FIG. 19 shows the specific enthalpy distribution in the blade height direction of the final stage portion of the steam turbine according to the present embodiment.
  • the distribution in the height direction of the total specific enthalpy of the most ideal final stage inlet of the present invention is a distribution in which the total specific enthalpy H0 on the outer peripheral side increases continuously as shown in FIG.
  • the final paragraph is divided into a plurality of blocks in the blade height direction, and by using the bypass steam from the turbine paragraph on the more upstream side of the block on the outer circumferential side, FIG.
  • FIG. A distribution is obtained in which the total specific enthalpy H0 increases in a stepwise manner toward the outer peripheral side in the span direction as shown in FIG.
  • the stator blade outflow velocity can be increased by increasing the specific enthalpy difference ⁇ h of the vanes on the outer peripheral side of the blade length. Therefore, since the stator blade outflow velocity can be increased, the rotor relative inflow velocity can be suppressed, and the same effect as that of the first embodiment can be obtained.
  • the moving blade can be increased by increasing the specific enthalpy difference ⁇ h of the stationary blades to increase the flowing velocity of the stationary blades in accordance with the increase in the peripheral velocity of the moving blades toward the outer peripheral side.
  • Changes in the relative inflow angle with respect to In this case in consideration of the incident angle range in which the above-mentioned loss does not increase sharply, for example, even if the blade length is large, using a moving blade having a constant, non-twisting with the blade inlet angle 90 and having high efficiency It becomes possible.
  • blades having no twist angle also have the advantage of being easy to manufacture.
  • FIG. 20 is a meridional cross-sectional view showing a main part structure of a turbine stage portion of a steam turbine according to a sixth embodiment of the present invention.
  • symbol is attached
  • the present embodiment is an application example of the first embodiment shown in FIG. 1 and differs from the first embodiment in that, as indicated by the arrow 81 on the outer peripheral side of the bypass flow passage 29, It is in the point which provided the extraction slit 64 which extracts a part to the turbine exterior.
  • the bleed slit is provided with a small opening length in the axial direction due to the restriction of the overall length of the turbine rotor in a general turbine. Therefore, when the amount of bleed air flow is large, the flow velocity increases at the bleed slit and the loss increases.
  • the present invention is not limited to the low pressure turbine and the last paragraph, and the effect of the present invention can be applied to an intermediate stage. You can get it. Further, the effect of the present invention is effective regardless of working fluid such as steam and air.
  • the first advantage is the reduction of wetness loss.
  • the water film attached to the blade surface of the moving blade 7 constituting the upstream side paragraph of the final paragraph is collected on the outer peripheral side by the centrifugal force, and is directed to the vane 12 of the final paragraph. Released. Therefore, the degree of wetness increases on the outer peripheral side of the final stage inlet, which causes an increase in wet loss and an increase in erosion in the final stage where the moving blade circumferential speed is large.
  • the present invention when the present invention is applied to the final stage of the steam turbine, the degree of wetness, which is the mass fraction of the liquid phase, decreases because the total inlet enthalpy on the outer side of the final stage is large.
  • the wet loss is reduced and the occurrence of erosion can also be suppressed. Therefore, the turbine efficiency can be improved, and the reliability of the steam turbine can also be improved.
  • the second advantage is the ability to improve wing reliability.
  • the Wilson line which transitions from the superheated steam of the steam turbine to wet steam in a two-phase flow state, is often located one turbine stage upstream of the final stage.
  • the Wilson line moves in the flow direction depending on the turbine load and steam conditions, so in the turbine stage where the Wilson line exists, the state of dry steam and wet steam is repeated and corrosion pits are likely to occur.
  • the turbine stage one upstream side of the final paragraph where the Wilson line occurs has a small blade length, so the stress applied to the blade can be reduced and the reliability of the blade is reduced due to corrosion pits. Can be suppressed.

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  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
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  • Turbine Rotor Nozzle Sealing (AREA)

Abstract

An axial flow turbine wherein a shock wave loss due to an increase in an annular band area and a loss due to separation are minimized to improve turbine efficiency. The axial flow turbine which is provided with turbine stages comprising stator blades fixed to a stationary body, and moving blades fixed to a turbine rotor, and a working fluid flow path (2) having a plurality of the turbine stages in the axial direction of the turbine is equipped with a turbine stage bypass flow path (29) which is provided within the working fluid flow path (2) so as to allow part of the working fluid flowing from the upstream side in the flow direction of the working fluid to bypass the outer peripheral side of at least one of the turbine stages to be introduced into a moving blade (13) on the downstream side in the flow direction of the working fluid of the bypassed turbine stage.

Description

軸流タービンAxial flow turbine
　本発明は、蒸気タービンや、ガスタービン等の軸流タービンに関する。 The present invention relates to an axial flow turbine such as a steam turbine and a gas turbine.
　特許文献１には、外周側ダイアフラムおよび内周側ダイアフラムに固定された静翼と、タービン中心軸回りに回転するタービンロータに固定された動翼とからなるタービン段落を作動流体流路内に複数備え、高圧の作動流体が流路内低圧部に向かって膨張する時に生じる運動エネルギーを、静翼と動翼とから構成されるタービン段落により回転力に変える機能を持つ軸流タービンが開示されている。 In Patent Document 1, a plurality of turbine stages each including a stationary blade fixed to an outer peripheral side diaphragm and an inner peripheral side diaphragm and a moving blade fixed to a turbine rotor rotating around a central axis of the turbine are provided in the working fluid flow path. Disclosed is an axial flow turbine having a function of converting kinetic energy generated when a high pressure working fluid expands toward a low pressure portion in the flow path into a rotational force by a turbine stage composed of a stationary blade and a moving blade. There is.
　軸流タービンでは、段落当たりの出力を増加させるために、単位時間当たりに流れる作動流体の質量である流量を増加させたい要求がある。流量を増加させ、段落当たりの出力を増加させることによって、段落数を変えずに発電量を増加させることが可能となる。 In an axial flow turbine, there is a need to increase the flow rate, which is the mass of working fluid flowing per unit time, in order to increase the power per stage. By increasing the flow rate and increasing the output per paragraph, it is possible to increase the amount of power generation without changing the number of paragraphs.
　ここで、流量を増加させるためには、作動流体が流れる部分のタービン回転軸方向からみた面積である環帯面積を大きくすることが有効であることが知られている。そこで、軸流タービンの場合には、環帯面積は翼長と、翼の外周端直径と内周端直径とを足して２で割った平均直径との積に円周率を掛けたものとなるため、環帯面積の増加のために、翼長と平均直径を大きくすることが行われている。 Here, in order to increase the flow rate, it is known that it is effective to increase an annular area which is an area viewed from the direction of the turbine rotation axis of a portion through which the working fluid flows. Therefore, in the case of an axial flow turbine, the ring area is the product of the blade length and the mean diameter obtained by adding the outer peripheral end diameter and the inner peripheral end diameter of the blade and dividing by 2 Therefore, in order to increase the ring area, it is practiced to increase the wing length and the average diameter.
特開２００３－２７９０１号公報Japanese Patent Application Publication No. 2003-27901
　一般的に、タービン段落入口における、作動流体の単位質量当たりのエンタルピー（比エンタルピー）と、流速の二乗を２で割った単位質量当たりの運動エネルギーとの和である比全エンタルピーＨ０は、回転軸に近い内周側から外周側にかけて、略一定の値とされる。一方、静翼と動翼との間の比エンタルピーｈ１は、静動翼間の旋回流とバランスするように内周側に比べ外周側にいくほど大きくなる。従って、比エンタルピー差Ｈ０－ｈ１は、外周側ほど小さくなる。静翼から出る流れの速度は、この比エンタルピー差Ｈ０－ｈ１の二乗根に比例する。即ち、静翼流出速度は外周側ほど小さくなる。 Generally, the specific total enthalpy H0, which is the sum of the enthalpy (specific enthalpy) per unit mass of working fluid and the kinetic energy per unit mass divided by 2 of the flow velocity at the turbine stage inlet, is the rotational axis The value is substantially constant from the inner circumferential side to the outer circumferential side close to On the other hand, the specific enthalpy h1 between the stationary blade and the moving blade becomes larger toward the outer peripheral side than the inner peripheral side so as to be balanced with the swirling flow between the stationary and moving blades. Therefore, the specific enthalpy difference H0-h1 decreases toward the outer periphery. The velocity of the flow out of the vane is proportional to the square root of this specific enthalpy difference H0-h1. That is, the vane outflow velocity decreases toward the outer periphery.
　ところで、上述したように、環帯面積を大きくする、すなわち翼長や平均直径を大きくすると、外周側の比エンタルピー差Ｈ０－ｈ１は、さらに小さくなり、静翼流出速度も小さくなる。このように、環帯面積を大きくすることにより、外周側の比エンタルピー差Ｈ０－ｈ１と静翼流出速度が小さくなることは、以下に述べるような二つの問題を引き起こす可能性がある。 By the way, as described above, when the annular zone area is increased, that is, the blade length and the average diameter are increased, the specific enthalpy difference H0-h1 on the outer peripheral side is further reduced, and the stator blade outflow velocity is also reduced. In this way, the specific enthalpy difference H0-h1 on the outer peripheral side and the decrease in the stator blade outflow velocity by increasing the annular area can cause the following two problems.
　一つ目は、動翼の相対流入マッハ数が超音速となり、損失が増加する可能性が増えることである。翼長や、平均直径を大きくすると、動翼の回転速度である周速が大きくなる。動翼の周速は、半径位置が一番大きい外周端、すなわち動翼先端部で最も大きくなる。先端部の周速を音速で割った周速マッハ数が１を超えて超音速となると、静翼からの流れの回転方向成分が十分でないと、動翼に流入してくる流れの相対速度が超音速となる。相対流入速度が超音速となると、動翼上流側で不連続な圧力上昇を伴う衝撃波が発生し、衝撃波そのものによるエントロピー上昇に加え、衝撃波が翼面の境界層と干渉して、その不連続な圧力上昇により境界層厚さが増加する。さらにははく離を生じさせることなどによるエントロピー上昇も生じる。この衝撃波によるエントロピー上昇により、タービン段落の環帯面積を増加させ、作動流体の流量を増加させたにも関わらず、増加流量に相当する回転力すなわち出力が増えないことがある。そのため、限界周速を超えて環帯面積を大きくすることにより、段落当たりの出力増加を実現するためには、動翼流入部で生じる衝撃波を無くす、もしくは弱くすることが重要であり、そのためには、動翼相対流入速度を小さくする必要がある。 The first is that the relative inflow Mach number of the moving blade becomes supersonic, and the possibility of loss increases. When the blade length or the average diameter is increased, the peripheral speed, which is the rotational speed of the moving blades, increases. The circumferential velocity of the moving blade is the largest at the outer peripheral end where the radial position is the largest, that is, at the blade tip. If the circumferential speed Mach number obtained by dividing the circumferential speed at the tip by the speed of sound exceeds 1 and becomes supersonic, the relative velocity of the stream flowing into the moving blade will be reduced if the rotational direction component of the flow from the stationary blade is insufficient. It becomes supersonic speed. When the relative inflow velocity becomes supersonic, a shock wave with a discontinuous pressure rise occurs on the upstream side of the moving blade, and in addition to the entropic rise due to the shock wave itself, the shock wave interferes with the boundary layer of the wing surface, and the discontinuity The boundary layer thickness increases due to the pressure increase. Furthermore, there is also an increase in entropy due to the occurrence of peeling or the like. Although the entropy increase due to the shock wave increases the annular area of the turbine stage and increases the flow rate of the working fluid, the rotational force or output corresponding to the increase flow may not increase. Therefore, it is important to eliminate or weaken the shock wave generated at the inflow portion of the moving blade in order to realize an increase in power per paragraph by increasing the annular area beyond the limit circumferential speed. Needs to reduce the relative inflow velocity of the moving blades.
　二つ目は、外周側の拡大流路部で、はく離が起きる可能性が増大することである。タービン段落の環帯面積を大きくすると、子午面流路の拡大率、すなわちタービン段落入口の流路高さに対する、段落出口の流路高さの増加率が大きくなる。一方、タービン段落の環帯面積を大きくしても、段落の軸方向長さは、タービン全体の長さに制約があるために一般的にはあまり大きくできず、子午面流路の拡大率の増大は、静翼部の外周端や内周端の子午面流路形状の広がり角を大きくすることで実現されることが一般的である。子午面流路形状の広がり角が大きくても、静翼部の比エンタルピー差Ｈ０－ｈ１が大きければ、翼間で流れが加速されるため、はく離が起きるなどの問題は起きないが、環帯面積を大きくするために、翼長や平均直径を大きくすると、外周側の静翼部の比エンタルピー差Ｈ０－ｈ１が小さくなり、子午面流路の外周端の拡大流路部で、流れの加速が小さくなり、流れが翼面や側壁面からはく離して、損失が増大する可能性が大きくなる。 The second is that the possibility of the occurrence of peeling increases in the expanded flow passage portion on the outer peripheral side. When the annular zone area of the turbine stage is increased, the rate of increase of the flow path height at the stage outlet with respect to the expansion rate of the meridional plane flow path, that is, the flow path height at the inlet of the turbine stage increases. On the other hand, even if the annular zone area of the turbine stage is increased, the axial length of the stage can not generally be increased because the length of the entire turbine is limited. Generally, the increase is realized by increasing the spread angle of the shape of the meridional flow passage at the outer peripheral end or the inner peripheral end of the stationary blade portion. Even if the divergence angle of the meridional flow channel shape is large, if the specific enthalpy difference H0-h1 of the stationary blade portion is large, the flow is accelerated between the blades, and thus problems such as peeling do not occur. When the blade length and the average diameter are increased in order to increase the area, the specific enthalpy difference H0-h1 of the stator vane portion on the outer peripheral side decreases, and the flow is accelerated in the expanded flow path portion of the outer peripheral end of the meridional flow path. Becomes smaller, and the flow separates from the wing and side wall surfaces, increasing the possibility of increased losses.
　そこで、本発明の目的は、環帯面積の増加による衝撃波損失とはく離による損失を抑制し、タービン効率を向上させることができる軸流タービンを提供することにある。 Therefore, an object of the present invention is to provide an axial flow turbine capable of suppressing the shock wave loss due to the increase of the annular area and the loss due to separation, and improving the turbine efficiency.
　上記目的を達成するため、静止体に固定された静翼と、タービンロータに固定された動翼とからなるタービン段落を作動流体流路中に複数備える軸流タービンにおいて、作動流体流路中に設けられ、作動流体流れ方向上流側から流入する作動流体の一部を、タービン段落の少なくとも一段の外周側をバイパスさせ、バイパスしたタービン段落の作動流体流れ方向下流側にある動翼に導入するタービン段落バイパス流路を備える。 In order to achieve the above object, in an axial flow turbine including a plurality of turbine stages each including a stationary blade fixed to a stationary body and a moving blade fixed to a turbine rotor in a working fluid flow path, A turbine is provided, and a part of the working fluid flowing in from the working fluid flow direction upstream side is bypassed to at least one outer peripheral side of the turbine stage and introduced to the moving blades downstream of the bypassed turbine paragraph working fluid flow direction A paragraph bypass flow path is provided.
　本発明によれば、軸流タービンにおいて、環帯面積の増加による衝撃波損失とはく離による損失を抑制し、タービン効率を向上させることができる。 According to the present invention, in the axial flow turbine, it is possible to suppress the shock wave loss due to the increase in the annular area and the loss due to the separation, and to improve the turbine efficiency.
蒸気タービンのタービン段落部の要部構造を表す子午面断面図である。It is a meridional plane sectional view showing the important section structure of the turbine section of a steam turbine. 蒸気タービンのタービン最終段落部の、比エンタルピーの翼高さ方向分布を表したグラフである。It is a graph showing blade height direction distribution of specific enthalpy of a turbine last paragraph part of a steam turbine. 動翼の周速が大きい場合の、静翼流出速度と、動翼周速と、動翼の相対流入速度との関係を模式的に表す図である。FIG. 7 is a diagram schematically illustrating the relationship between the stator blade outflow velocity, the rotor circumferential velocity, and the relative inflow velocity of the moving blades when the circumferential velocity of the moving blades is high. 動翼周速が大きい場合の、動翼相対流入マッハ数の翼高さ方向分布を表すグラフである。It is a graph showing the blade height direction distribution of moving blade relative inflow Mach number in case the moving blade circumferential speed is large. 本発明の第１の実施の形態に係る蒸気タービン最終段落部の翼高さ方向の比エンタルピー分布を表すグラフである。It is a graph showing the specific enthalpy distribution of the wing height direction of the last paragraph part of the steam turbine concerning a 1st embodiment of the present invention. 本発明の第１の実施の形態に係る蒸気タービンの静翼流出速度、動翼周速、動翼相対流入速度の関係を模式的に表した説明図である。FIG. 3 is an explanatory view schematically showing the relationship among the stator blade outflow velocity, the moving blade peripheral velocity, and the moving blade relative inflow velocity of the steam turbine according to the first embodiment of the present invention. 本発明の第１の実施の形態に係る蒸気タービンのタービン段落部の要部構造を表す子午面断面図である。It is a meridional plane sectional view showing the important section structure of the turbine paragraph part of the steam turbine concerning a 1st embodiment of the present invention. 本発明の第１の実施の形態に係る蒸気タービンの分流板の要部構造を表す斜視図である。It is a perspective view showing the principal part structure of the distribution plate of the steam turbine concerning a 1st embodiment of the present invention. 本発明の第１の実施の形態に係る蒸気タービン最終段落と、その１つ上流側段落の翼高さ方向の、比エンタルピー分布を表すグラフである。It is a graph showing the specific enthalpy distribution of the last paragraph of the steam turbine concerning a 1st embodiment of the present invention, and the wing height direction of the one upper stream side paragraph. 本発明の第１の実施の形態に係る蒸気タービン最終段落部の動翼相対流入マッハ数の翼高さ方向分布を表すグラフである。It is a graph showing blade height direction distribution of bucket relative inflow Mach number of a steam turbine last paragraph part concerning a 1st embodiment of the present invention. 本発明の第１の実施の形態に係る蒸気タービン最終段落部の周方向から測った動翼入口角の翼高さ方向分布を表すグラフである。It is a graph showing blade height direction distribution of the moving blade inlet angle measured from the circumferential direction of the steam turbine last paragraph part concerning a 1st embodiment of the present invention. 本発明の第２の実施の形態に係る蒸気タービンのタービン段落部の要部構造を表す子午面断面図である。It is a meridional plane sectional view showing the important section structure of the turbine section of the steam turbine concerning a 2nd embodiment of the present invention. 本発明の第２の実施の形態に係る蒸気タービン最終段落部の周方向から測った動翼入口角の翼高さ方向分布を表すグラフである。It is a graph showing the blade height direction distribution of the moving blade inlet angle measured from the circumferential direction of the steam turbine last stage part which concerns on the 2nd Embodiment of this invention. 本発明の第３の実施の形態に係る蒸気タービンのタービン段落部の要部構造を表す子午面断面図である。It is a meridional plane sectional view showing the important section structure of the turbine section of the steam turbine concerning a 3rd embodiment of the present invention. 本発明の第３の実施の形態に係る蒸気タービンのシール構造部の拡大図である。It is an enlarged view of the seal structure part of the steam turbine concerning a 3rd embodiment of the present invention. 本発明の第３の実施の形態に係る蒸気タービン最終段落と、その１つ上流側段落の翼高さ方向の、比エンタルピー分布を表すグラフである。It is a graph showing the specific enthalpy distribution of the last paragraph of the steam turbine concerning a 3rd embodiment of the present invention, and the wing height direction of the one upper stream side paragraph. 本発明の第４の実施の形態に係る蒸気タービン最終段落部の周方向から測った動翼入口角の翼高さ方向分布を表すグラフである。It is a graph showing the blade height direction distribution of the moving blade inlet angle measured from the circumferential direction of the steam turbine last stage part which concerns on the 4th Embodiment of this invention. 本発明の第５の実施の形態に係る蒸気タービンのタービン段落部の要部構造を表す子午面断面図である。It is a meridional plane sectional view showing the important section structure of the turbine paragraph part of the steam turbine concerning a 5th embodiment of the present invention. 本発明の第５の実施の形態に係る蒸気タービン最終段落部の翼高さ方向の比エンタルピー分布を表すグラフである。It is a graph showing the specific enthalpy distribution of the blade height direction of the last paragraph part of the steam turbine concerning a 5th embodiment of the present invention. 本発明の第６の実施の形態に係る蒸気タービンのタービン段落部の要部構造を表す子午面断面図である。It is a meridional plane sectional view showing the important section structure of the turbine paragraph part of the steam turbine concerning a 6th embodiment of the present invention.
　以下、本発明を実施するための形態について、適宜図を参照して詳細に説明する。なお、各図面を通し、同等の構成要素には同一の符号を付してある。 Hereinafter, embodiments of the present invention will be described in detail with reference to the drawings. The same reference numerals are given to the same components throughout the drawings.
　本発明の第１の実施の形態として、本発明を蒸気タービンの最終段落に適用した例について、以下説明する。 As a first embodiment of the present invention, an example in which the present invention is applied to the final paragraph of a steam turbine will be described below.
　初めに、図１乃至図４を用いて、一般的な蒸気タービン段落部の基本構成および動作について説明する。 First, the basic configuration and operation of a general steam turbine stage will be described using FIGS. 1 to 4.
　図１は、一般的な蒸気タービンのタービン段落部の要部構造を表す子午面断面図である。蒸気タービンのタービン段落は、作動流体である蒸気１の流れ方向上流側（以下、単に上流側と記載する）の蒸気高圧部Ｐ０と、蒸気流れ方向下流側（以下、単に下流側と記載する）の蒸気低圧部Ｐ１との間の蒸気主流路２に設けられている。タービンケーシング３の内周側に外周側ダイアフラム４が固定されている。タービン段落は、外周側ダイアフラム４と内周側ダイアフラム５との間に固定された静翼６と、静翼６の下流側に対向するようにタービンロータ８に固定された動翼７とで構成される。タービン段落が複数の段落から構成される多段落型タービンである場合、タービン中心軸９に沿って、高圧部Ｐ０から低圧部Ｐ１に向かって、タービン段落が複数回繰り返されて設けられる。高圧部Ｐ０の蒸気入口に最も近い段落を初段落といい、低圧部Ｐ１の蒸気出口に最も近い段落を最終段落という。タービンロータ８は、図示しない発電機に機械的に接続されている。従って、蒸気タービンは、高圧の蒸気が低圧部に向かって膨張する時に生じる運動エネルギーを、静翼と動翼から構成されるタービン段落により回転力に変え、タービンロータを介して発電機で電気エネルギーに変換して発電を行う。 FIG. 1 is a meridional cross-sectional view showing a main part structure of a turbine stage portion of a general steam turbine. In the turbine stage of the steam turbine, the high-pressure part P0 of the working fluid upstream of the working flow of the steam 1 (hereinafter referred to simply as upstream) and the downstream flow direction (hereinafter referred to simply as downstream) It is provided in the steam main flow path 2 between the steam low pressure portion P1 of An outer peripheral diaphragm 4 is fixed to the inner peripheral side of the turbine casing 3. The turbine stage comprises a stationary blade 6 fixed between the outer peripheral side diaphragm 4 and the inner peripheral side diaphragm 5 and a moving blade 7 fixed to the turbine rotor 8 so as to face the downstream side of the stationary blade 6. Be done. When the turbine stage is a multi-stage turbine including a plurality of stages, the turbine stage is provided by being repeated a plurality of times from the high pressure portion P0 to the low pressure portion P1 along the turbine center axis 9. The paragraph closest to the steam inlet of the high pressure part P0 is called the first paragraph, and the paragraph closest to the steam outlet of the low pressure part P1 is called the final paragraph. The turbine rotor 8 is mechanically connected to a generator (not shown). Therefore, the steam turbine converts the kinetic energy generated when high-pressure steam expands toward the low pressure part into rotational force by the turbine stage composed of the stationary blades and the moving blades, and generates electrical energy from the generator via the turbine rotor. To generate electricity.
　図２は、図１に図示した蒸気タービンのタービン最終段落部の、翼高さ方向の比エンタルピー分布を表したグラフである。横軸は、比エンタルピーであり、図１において作動流体である蒸気が左から右に流れていくのに合わせて、左に行くほど比エンタルピーが大きくなるように軸方向を決めている。縦軸は、翼高さ方向を表し、ＢＨは、動翼出口高さを表す。 FIG. 2 is a graph showing the specific enthalpy distribution in the blade height direction of the final stage portion of the steam turbine illustrated in FIG. The horizontal axis is the specific enthalpy, and the axial direction is determined so that the specific enthalpy increases as the working fluid flows from left to right in FIG. The vertical axis represents the blade height direction, and BH represents the blade outlet height.
　図２において、Ｈ０は、段落入口における単位質量当たりのエンタルピーと作動流体流速の二乗を２で割った単位質量当たりの運動エネルギーとの和である全比エンタルピー、ｈ１は静・動翼間の比エンタルピー、ｈ２は段落出口の比エンタルピーを表す。段落入口の全比エンタルピーＨ０は、翼高さ方向に略一定である。静・動翼間の比エンタルピーｈ１は、主に静動翼間の旋回速度による遠心力とバランスするようにタービン半径方向外周側（以下、単に外周側と記載する）ほど大きくなる。結果として、外周側の静翼にかかる比エンタルピー差Δｈが小さくなり、比エンタルピー差の二乗根に比例する静翼流出速度も小さくなる。この静翼にかかる比エンタルピー差と静翼流出速度が小さくなる傾向は、翼長や、翼の外周端直径と内周端直径とを足して２で割った平均直径が大きくなることにより、即ち環帯面積が大きくなることにより翼の外周端位置がより外周側になるほど顕著となる。 In FIG. 2, H0 is the total specific enthalpy which is the sum of the enthalpy per unit mass at the stage inlet and the kinetic energy per unit mass divided by 2 of working fluid flow velocity divided by 2, h1 is the ratio between static and moving blades Enthalpy, h2 represents the specific enthalpy of the stage exit. The total specific enthalpy H0 at the stage inlet is substantially constant in the wing height direction. The specific enthalpy h1 between the stationary and moving blades becomes larger toward the outer peripheral side in the radial direction of the turbine (hereinafter, simply referred to as the outer peripheral side) so as to be balanced mainly with the centrifugal force due to the turning speed between the stationary blades. As a result, the specific enthalpy difference Δh applied to the outer peripheral side stationary blades decreases, and the stationary blade outflow velocity proportional to the square root of the specific enthalpy difference also decreases. The specific enthalpy difference on the vane and the tendency of the vane outflow velocity to decrease are increased by the wing length and the average diameter of the outer peripheral end diameter and the inner peripheral end diameter of the wing divided by 2; The larger the annular area, the more prominent the outer peripheral end position of the wing becomes on the outer peripheral side.
　図３は、動翼の周速が大きい場合の、静翼流出速度と、動翼周速と、動翼の相対流入速度との関係を模式的に表す図である。高圧Ｐ３の蒸気２２は、静翼２３を通過することによって、加速、転向され速度Ｖの流れとなる（以下、この速度を静翼流出速度Ｖと記載する）。この静翼流出速度Ｖを動翼２４と一緒に回転する相対座標系で見ると、動翼２４はタービン中心軸周りに周速Ｕで回転しているため（回転方向を矢印２５で表す）、静翼流出速度ベクトルＶと周速ベクトルＵとの合成により、動翼２４への流入する蒸気は相対的に速度Ｗの流れとなる（以下、動翼相対流入速度Ｗと記載する）。この静翼流出速度ベクトルＶ、周速ベクトルＵ、動翼相対流入速度ベクトルＷによって形成される三角形を速度三角形と呼ぶ。速度三角形から明らかなように、動翼周速Ｕが大きくなると、動翼２４に流入する動翼相対速度Ｗは大きくなる。動翼相対流入速度Ｗを小さくするためには、静翼流出速度Ｖを大きくする必要がある。 FIG. 3 is a diagram schematically showing the relationship between the stator blade outflow velocity, the rotor peripheral velocity, and the relative inflow velocity of the moving blades when the peripheral velocity of the moving blades is large. The high-pressure P3 steam 22 is accelerated and turned into a flow of velocity V by passing through the stationary blade 23 (hereinafter, this velocity is referred to as stationary blade outflow velocity V). When this stationary blade outflow velocity V is viewed in a relative coordinate system rotating with the moving blades 24, the moving blades 24 are rotated at a circumferential velocity U around the central axis of the turbine (the direction of rotation is represented by the arrow 25), Due to the combination of the stator blade outflow velocity vector V and the circumferential velocity vector U, the steam flowing into the moving blades 24 relatively flows at a velocity W (hereinafter referred to as the relative blade velocity inflow W). A triangle formed by the stationary blade outflow velocity vector V, the circumferential velocity vector U, and the moving blade relative inflow velocity vector W is referred to as a velocity triangle. As apparent from the velocity triangle, as the circumferential velocity U of the moving blade increases, the relative velocity W of the moving blade flowing into the moving blade 24 increases. In order to reduce the rotor relative inflow velocity W, it is necessary to increase the stator vane outflow velocity V.
　ここで、段落入口の蒸気の状態量が固定されているとき、静翼流出速度Ｖを大きくするためには、静・動翼間での比エンタルピーｈ１を小さくして、静翼の比エンタルピー差Δｈを大きくする必要がある。しかしながら、静・動翼間での比エンタルピーは静翼出口の旋回速度場によって外周側ほど大きくなり、翼長が長くなるほど、旋回速度場の影響が強くなるので、ｈ１を小さくすることは難しくなる。すなわち、翼長が長くなるほど、静翼の比エンタルピー差Δｈを大きくすることは難しくなり、静翼流出速度Ｖを大きくすることは難しくなる。 Here, when the quantity of state of the steam at the inlet of the stage is fixed, the specific enthalpy h1 between the stationary and moving blades is reduced to increase the specific enthalpy difference of the stationary blades in order to increase the vane outflow velocity V. It is necessary to increase Δh. However, the specific enthalpy between the static and moving blades increases toward the outer periphery due to the turning velocity field at the exit of the stationary blade, and the longer the blade length, the stronger the effect of the turning velocity field, making it difficult to reduce h1. . That is, as the blade length increases, it becomes more difficult to increase the specific enthalpy difference Δh of the stator blades, and it becomes more difficult to increase the stator blade outflow velocity V.
　図４は、動翼周速が大きい場合の、動翼相対流入マッハ数の翼高さ方向分布を表すグラフである。 FIG. 4 is a graph showing the blade height direction distribution of the rotor inflow relative Mach number when the rotor peripheral speed is high.
　図４に示したように、翼の外周側では、マッハ数が１.０を超え、超音速流入となっていることがわかる。ここで、外周側の静翼の比エンタルピー差Δｈを大きくするために、タービン段落のエンタルピー差Ｈ０－ｈ２自体を大きくすると、内周端の動翼相対流入マッハ数が１.０を超え超音速流入となるために、段落全体のエンタルピー差を大きくすることでは、超音速流入の問題を解決することは難しい。 As shown in FIG. 4, it can be seen that the Mach number exceeds 1.0 on the outer peripheral side of the wing, and supersonic inflow occurs. Here, if the enthalpy difference H0-h2 itself of the turbine stage is increased in order to increase the specific enthalpy difference Δh of the outer peripheral side stationary blades, the rotor inflow Mach number of the inner peripheral end exceeds 1.0 and the supersonic speed It is difficult to solve the problem of supersonic inflow by increasing the enthalpy difference throughout the paragraph to become the inflow.
　以上を踏まえて、本発明の第１の実施形態ついて、図面を用いて説明する。 Based on the above, a first embodiment of the present invention will be described using the drawings.
　図５は、第１の実施形態に係る蒸気タービン最終段落の翼高さ方向の比エンタルピー分布を表すグラフである。なお、図５においても図２と同様の軸構成を用いている。 FIG. 5 is a graph showing the specific enthalpy distribution in the blade height direction of the final stage of the steam turbine according to the first embodiment. Also in FIG. 5, the same shaft configuration as in FIG. 2 is used.
　本発明では、段落入口部の全比エンタルピーＨ０を外周側ほど大きくすることで、静翼外周側の比エンタルピー差Δｈを大きくしている。比エンタルピー差Δｈを大きくすることで、比エンタルピー差Δｈの二乗根に比例する静翼流出速度Ｖも大きくなる。よって、図６に示すように、段落入口部の全比エンタルピーＨ０を一定にした場合Ｖ１となる静翼流出速度をＶ２へと大きくすることができる。静翼流出速度成分Ｖの旋回方向の速度成分もＶＴ１からＶＴ２へ大きくすることができる。その結果、周速Ｕが同じにも係らず、動翼２４に流入する動翼相対流入速度Ｗ１をＷ２に減速させることができる。 In the present invention, the specific enthalpy difference Δh on the outer periphery side of the stationary blade is increased by increasing the total specific enthalpy H0 of the stage inlet portion toward the outer periphery side. By increasing the specific enthalpy difference Δh, the stator blade outflow velocity V proportional to the square root of the specific enthalpy difference Δh also increases. Therefore, as shown in FIG. 6, when the total specific enthalpy H0 at the stage inlet is constant, the stator blade outflow velocity, which is V1, can be increased to V2. The velocity component in the swirling direction of the vane outflow velocity component V can also be increased from VT1 to VT2. As a result, it is possible to reduce the moving blade relative inflow velocity W1 flowing into the moving blade 24 to W2 despite the same circumferential velocity U.
　よって、本発明によれば、段落入口の全比エンタルピーＨ０を、動翼相対流入速度Ｗ２が音速以下となるまで大きくすることで、超音速流入を回避でき、動翼入口の衝撃波の発生を抑制して衝撃波の発生に伴う損失を抑制できる。 Therefore, according to the present invention, supersonic flow can be avoided by increasing the total specific enthalpy H0 of the stage inlet until the blade relative inflow velocity W2 becomes lower than the speed of sound, thereby suppressing the generation of shock waves at the blade inlet. Thus, it is possible to suppress the loss associated with the generation of shock waves.
　また、本発明によれば、翼外周側の静翼部の比エンタルピー差Δｈが大きくなり、子午面流路の外周端の拡大流路部で、流れの加速が大きくなり、作動流体である蒸気流れが翼面や蒸気主流路壁を構成する側壁面からはく離することを抑制し、流れのはく離に伴う損失を抑制できる。 Further, according to the present invention, the specific enthalpy difference Δh of the stationary blade portion on the outer peripheral side of the blade becomes large, and the acceleration of the flow becomes large in the enlarged flow path portion of the outer peripheral end of the meridional flow path. It is possible to suppress the separation of the flow from the wing surface and the side wall surface constituting the main steam flow path wall, and to suppress the loss associated with the separation of the flow.
　本発明によれば、環帯面積の増加による衝撃波損失とはく離による損失を抑制し、タービン段落効率を向上させることができる。 According to the present invention, it is possible to suppress the shock wave loss due to the increase of the annular area and the loss due to the separation, and to improve the turbine stage efficiency.
　次に、本発明の第１の実施形態に係る蒸気タービンの基本構成および動作について説明する。 Next, the basic configuration and operation of the steam turbine according to the first embodiment of the present invention will be described.
　図７は、本発明の第１の実施形態に係る蒸気タービンのタービン段落部の要部構造の子午面断面図である。図７に示すように、本実施形態に係る蒸気タービンでは、最終段落の１つ上流側のタービン段落部に蒸気主流路２をタービン半径方向に二分割する分流板２６が設けられている。本実施形態では、分流板２６は、最終段落の静翼１２から最終段落の２つ上流側にあるタービン段落の動翼２１出口付近まで延伸しており、最終段落を構成する静翼１２に支持されている。分流板２６は、図１０に示すように流入マッハ数が１.０を超える高さより内周側に設けることが望ましい。分流板２６の上流側端部は、作動流体の流れを阻害しないように、上流側から下流側に向かって徐々に厚さを大きくするように形成されている。 FIG. 7 is a meridional cross-sectional view of the main part structure of the turbine stage portion of the steam turbine according to the first embodiment of the present invention. As shown in FIG. 7, in the steam turbine according to the present embodiment, a flow dividing plate 26 that divides the main steam flow passage 2 into two in the radial direction of the turbine is provided in the turbine stage portion one upstream side of the final stage. In the present embodiment, the flow dividing plate 26 extends from the vanes 12 of the final stage to the vicinity of the blade 21 outlet of the turbine stage located two upstream sides of the final stage, and is supported by the vanes 12 constituting the final stage. It is done. The flow dividing plate 26 is preferably provided on the inner peripheral side of the height where the inflow Mach number exceeds 1.0 as shown in FIG. The upstream end of the flow dividing plate 26 is formed to gradually increase in thickness from the upstream side to the downstream side so as not to obstruct the flow of the working fluid.
　本実施形態では、最終段落の１つ上流側のタービン段落は、分流板２６の内周側に設けられている。分流板２６の上流側端部の内周側に静翼２７を周方向に列設し、静翼２７の下流側に動翼２８を配置して最終段落の１つ上流側のタービン段落を構成している。本実施形態では、分流板２６を設けることによって、分流板２６の外周側に、バイパス流路２９を形成し、内周側に、段落流路３０を形成している。よって、段落流路３０内の動翼２８の出口流路高さは、上流側段落の動翼２１出口流路高さと比較して小さく、かつ下流側段落の動翼１３出口流路高さと比較しても小さい。 In the present embodiment, the turbine stage one upstream side of the final stage is provided on the inner peripheral side of the flow dividing plate 26. The stator vanes 27 are circumferentially arrayed on the inner peripheral side of the upstream end of the flow dividing plate 26, and the rotor vanes 28 are arranged downstream of the stator vanes 27 to configure a turbine stage one upstream side of the final paragraph doing. In the present embodiment, by providing the flow dividing plate 26, the bypass flow passage 29 is formed on the outer peripheral side of the flow dividing plate 26, and the round flow passage 30 is formed on the inner peripheral side. Therefore, the outlet channel height of the moving blade 28 in the paragraph channel 30 is smaller than the outlet channel height of the rotor blade 21 in the upstream stage, and is compared with the outlet channel height of the rotor blade 13 in the downstream side Even small.
　なお、分流板２６が、内周側に設けられた静翼２７の重みや、流体力でたわむ可能性がある場合、または分流板２６の静翼１２固定部分に許容応力以上の大きな応力が生じる場合には、外周側ダイアフラム４等の外周側静止部から伸びる円柱状のサポート部材３１を周方向に複数本設置して分流板２６を支持しても良い。 Note that if there is a possibility that the flow dividing plate 26 may be warped by the weight of the vanes 27 provided on the inner circumferential side or by fluid force, or a large stress exceeding the allowable stress occurs in the stationary blade 12 fixing portion of the flow dividing plate 26 In this case, a plurality of cylindrical support members 31 extending from the outer peripheral side stationary portion such as the outer peripheral side diaphragm 4 may be installed in the circumferential direction to support the flow dividing plate 26.
　図８は、図７に示した分流板２６の斜視図である。図８に示すように、分流板２６は、周方向に複数の板部材を連結して構成されており、上流側から下流側に向かって、徐々に内径値が大きくなる円錐状に近い形状を成している。 FIG. 8 is a perspective view of the flow dividing plate 26 shown in FIG. As shown in FIG. 8, the flow dividing plate 26 is configured by connecting a plurality of plate members in the circumferential direction, and has a nearly conical shape whose inner diameter value gradually increases from the upstream side toward the downstream side. It is
　本実施形態に係る蒸気タービンによれば、最終段落の二つ上流側のタービン段落から吐出した蒸気主流は、分流板２６によってタービン半径方向に二つの流れに二分割され、分流板２６外周側のバイパス流路２９に流入した蒸気流は、直接最終段落の静翼１２の外周側に流入する。一方、分流板２６内周側の段落流路３０に流入した蒸気流は、分流板２６内周側の最終段落の１つ上流側のタービン段落を経由して最終段落の内周側にのみ流入する。 According to the steam turbine according to the present embodiment, the main steam discharged from the two upstream turbine stages in the final stage is divided into two in the radial direction of the turbine by the flow dividing plate 26, and the outer peripheral side of the flow dividing plate 26 The steam flow that has flowed into the bypass flow path 29 flows directly to the outer peripheral side of the stator vane 12 of the final stage. On the other hand, the steam flow that has flowed into the stage flow passage 30 on the inner peripheral side of the flow dividing plate 26 flows only to the inner peripheral side of the final paragraph via the turbine stage on the upstream side of one final stage on the inner peripheral side of the flow dividing plate 26 Do.
　図９は、本実施の形態に係る蒸気タービン最終段落と、その１つ上流側段落の、翼高さ方向の比エンタルピー分布を表すグラフである。Ｈ０は、図７に示した静翼２７の上流側の全比エンタルピーである。分流板２６位置ｂｈより翼高さ方向外周側ではこのＨ０が、最終段落の入口全比エンタルピーとなる。ｈ１は、静翼２７と動翼２８との間の比エンタルピー、Ｈ２は、静翼１２の内周側の入口全比エンタルピー、ｈ３は静翼１２と動翼１３との間の内周側の比エンタルピー、ｈ５は静翼１２と動翼１３との間の外周側の比エンタルピー、ｈ５ａはｈ５に相当する図１に示した一般的な蒸気タービン最終段落の比エンタルピーである。外周側の段落入口全比エンタルピーを大きくすることで、静翼の外周側の比エンタルピー差Δｈを大きくできる。このように外周側においても、大きな比エンタルピー差を確保できるため、静翼外周側で子午面流路の広がり角が大きい場合においても、はく離による性能低下を抑制できる。 FIG. 9 is a graph showing the specific enthalpy distribution in the blade height direction in the final stage of the steam turbine and the stage immediately upstream thereof according to the present embodiment. H0 is the total specific enthalpy of the upstream side of the stationary blade 27 shown in FIG. This H0 becomes the inlet total specific enthalpy of the final paragraph on the outer peripheral side in the blade height direction from the diverting plate 26 position bh. h1 is a specific enthalpy between the stationary blade 27 and the moving blade 28, H2 is a total inlet specific enthalpy of the inner peripheral side of the stationary blade 12, h3 is an inner peripheral side between the stationary blade 12 and the moving blade 13 The specific enthalpy, h5 is the specific enthalpy on the outer peripheral side between the stationary blade 12 and the moving blade 13, and h5a is the specific enthalpy of the last stage of the general steam turbine shown in FIG. 1 corresponding to h5. The specific enthalpy difference Δh on the outer peripheral side of the stator blade can be increased by increasing the overall specific enthalpy of the stage inlet on the outer peripheral side. As described above, a large specific enthalpy difference can be secured also on the outer peripheral side, and therefore, even when the spread angle of the meridional surface flow channel is large on the stator outer peripheral side, performance deterioration due to peeling can be suppressed.
　図１０は、図７に示した本実施の形態に係る蒸気タービン最終段落の動翼相対流入マッハ数の翼高さ方向分布を表すグラフである。実線で示したＭ１ｒが本実施の形態に係る蒸気タービン最終段落の動翼相対流入マッハ数、Ｍ１ｒａが図１に示した一般的な蒸気タービン最終段落の動翼相対流入マッハ数である。分流板２６位置ｂｈより翼高さ方向外周側で静翼の比エンタルピー差Δｈを大きくしたため、静翼外周側の静翼流出速度が大きくなり、図６を用いて説明したように動翼に対する超音速流入が回避できている。よって、本発明の動翼相対流入マッハ数を低減する効果により、タービン段落の効率を向上できる。 FIG. 10 is a graph showing the distribution in the blade height direction of the relative inflowing Mach number of the moving blade in the final stage of the steam turbine according to the present embodiment shown in FIG. 7. M1r shown by a solid line is the rotor inflow relative Mach number in the final stage of the steam turbine according to the present embodiment, and M1ra is the rotor relative inflow Mach number in the general last stage of the steam turbine shown in FIG. Since the specific enthalpy difference Δh of the stationary blades is increased on the outer peripheral side in the blade height direction from the flow dividing plate 26 position bh, the stationary blade outflow velocity on the outer peripheral side of the stationary blades increases, and as described using FIG. Sound velocity inflow can be avoided. Therefore, the efficiency of the turbine stage can be improved by the effect of reducing the moving blade relative inflow Mach number of the present invention.
　図１１に、本実施の形態に係る蒸気タービン最終段落の周方向から測った動翼入口角の翼高さ方向分布を示す。実線で示したαｉｎが動翼入口角、点線で示したβｉｎが動翼に対する蒸気の相対流入角である。外周側の段落入口全比エンタルピーが不連続となる部分で、相対流入角も不連続に小さくなるが、動翼の入口角は連続としている。入口角と相対流入角の違いを入射角というが、入射角がプラス・マイナスのある範囲内であれば、急減な翼の損失増加は起きない。翼の入口角が９０度付近では、特にその範囲は大きく、例えば入射角がプラス・マイナス４０度の範囲内では、急激な損失増加は起きない。そのため、図１１に示すように、動翼入口角と動翼相対流入角が一致していなくても、効率が大きく悪化することはない。 FIG. 11 shows the blade height direction distribution of the moving blade inlet angle measured from the circumferential direction of the final stage of the steam turbine according to the present embodiment. .Alpha..sub.in indicated by a solid line is a blade inlet angle, and .beta..sub.in indicated by a dotted line is a relative inflow angle of steam to the blade. Although the relative inflow angle also becomes smaller discontinuously in the portion where the outer entrance side specific inlet enthalpy becomes discontinuous, the inlet angle of the moving blade is continuous. The difference between the inlet angle and the relative inflow angle is referred to as the incident angle, but if the incident angle is within a range of plus or minus, the sharp increase in wing loss does not occur. Especially in the vicinity of 90 degrees of the inlet angle of the wing, the range is particularly large, for example, in the range of plus and minus 40 degrees of the incidence angle, a rapid increase in loss does not occur. Therefore, as shown in FIG. 11, the efficiency does not greatly deteriorate even if the blade inlet angle and the blade relative inflow angle do not match.
　本実施の形態に係る蒸気タービンによれば、最終段落入口部の外周側の全比エンタルピーＨ０を大きくすることで、静翼外周側の比エンタルピー差Δｈを大きくしている。これにより静翼流出速度を大きくすることができる。従って、静翼流出速度成分Ｖの旋回方向の速度成分も大きくすることができ、周速Ｕが同じにも係らず、最終段落の動翼に流入する動翼相対流入速度を減速させることができる。よって、動翼への超音速流入を回避でき、動翼入口の衝撃波の発生を抑制し、衝撃波の発生に伴う損失を抑制できる。 According to the steam turbine according to the present embodiment, the specific enthalpy difference Δh on the outer periphery side of the stationary blade is increased by increasing the total specific enthalpy H0 on the outer peripheral side of the final stage inlet portion. This makes it possible to increase the vane outflow velocity. Therefore, it is possible to increase the velocity component in the swirling direction of the stator blade outflow velocity component V, and to reduce the relative inflow velocity of the moving blade flowing into the moving blade in the final stage despite the same circumferential velocity U. . Therefore, supersonic flow into the moving blades can be avoided, generation of a shock wave at the moving blade inlet can be suppressed, and loss associated with the generation of shock waves can be suppressed.
　また、本実施の形態の蒸気タービンによれば、静翼外周側の比エンタルピー差Δｈが大きくなり、子午面流路の外周端の拡大流路部で、流れの加速が大きくなるので、蒸気流れが翼面や蒸気主流路壁を構成する側壁面からはく離することを抑制し、流れのはく離に伴う損失を抑制できる。 Further, according to the steam turbine of the present embodiment, the specific enthalpy difference Δh on the outer peripheral side of the stationary blade becomes large, and the acceleration of the flow becomes large in the enlarged flow path portion of the outer peripheral end of the meridional flow path. Can suppress the separation from the wing surface and the side wall surface constituting the steam main flow path wall, and can suppress the loss accompanying the flow separation.
　なお、本実施形態では、静翼２７と動翼２８からなる段落により取り出される回転力は小さくなるが、その低下分は、静翼１２と動翼１３とからなる段落で回転力として取り出すことができるため、タービン全体としての回転力は減らない。むしろ、損失の低下した分、回転力は増加させることが可能となる。 In this embodiment, although the rotational force extracted by the paragraph consisting of the stator vanes 27 and the moving blades 28 becomes smaller, the reduction thereof can be taken out as the rotational force at the paragraph consisting of the stationary blades 12 and the moving vanes 13 As it can, the torque of the entire turbine is not reduced. Rather, the rotational force can be increased by the amount of loss reduction.
　次に、本発明の第２の実施形態について図面を用いて説明する。図１２は、本発明の第２の実施形態に係る蒸気タービンのタービン段落部の要部構造を表す子午面断面図である。なお、第１の実施形態と同等の構成要素には同一の符号を付し、説明を省略する。 Next, a second embodiment of the present invention will be described using the drawings. FIG. 12 is a meridional cross-sectional view showing a main part structure of a turbine stage portion of a steam turbine according to a second embodiment of the present invention. In addition, the same code | symbol is attached | subjected to a component equivalent to 1st Embodiment, and description is abbreviate | omitted.
　本実施形態が、図７に示した第１の実施形態と異なるのは、分流板２６の下流側にある最終段落の動翼１３にも、蒸気主流路２を内周側と外周側に二分割する第２の分流板３２を設けている点である。第２の分流板３２は、周方向に列設された動翼１３一本一本に固設され、隣接翼の第２の分流板３２と接触連結されている。なお、接触連結させることで、振動減衰機構を持たせたり、翼の振動モードを調節して共振回避設計したりすることにも利用することができる。 The present embodiment is different from the first embodiment shown in FIG. 7 in that the main blade 2 of the steam in the final stage on the downstream side of the flow dividing plate 26 is divided into two on the inner circumferential side and the outer circumferential side. The second flow dividing plate 32 to be divided is provided. The second flow dividing plate 32 is fixed to each of the moving blades 13 arranged in a row in the circumferential direction, and is connected in contact with the second flow dividing plate 32 of the adjacent wing. In addition, by making it contact-connect, it can utilize also for providing a vibration damping mechanism, adjusting the vibration mode of a wing | blade, and carrying out a resonance avoidance design.
　図１３に、本実施形態に係る蒸気タービン最終段落部の周方向から測った動翼入口角の翼高さ方向分布を示す。本実施形態によれば、第２の分流板３２を設けたことで、分流板の内周側と外周側で翼が滑らかにつながる必要がなくなるため、翼の内周側と外周側を独立に、それぞれの流れに合わせて設計することが可能となり、図１３に示すように最終段落における動翼入口角αｉｎと蒸気の動翼相対流入角βｉｎを一致させることができる。よって、第２の実施形態によれば、第１の実施形態の効果に加えて、翼高さ方向において、翼入口角が不連続であることによって生じる損失を抑制できる。また、第２の分流板３２は、動翼１３の振動減衰効果が期待でき、翼の振動モードを調節して共振回避することに用いることもできる。 FIG. 13 shows the blade height direction distribution of the moving blade inlet angle measured from the circumferential direction of the final stage portion of the steam turbine according to the present embodiment. According to the present embodiment, the provision of the second flow dividing plate 32 eliminates the need for the blades to be smoothly connected on the inner and outer peripheral sides of the flow dividing plate, so the inner and outer peripheral sides of the blade are independently provided. It is possible to design according to each flow, and as shown in FIG. 13, it is possible to match the blade inlet angle α in with the blade relative inflow angle β in of the steam in the final paragraph. Therefore, according to the second embodiment, in addition to the effects of the first embodiment, it is possible to suppress the loss caused by the blade inlet angle being discontinuous in the blade height direction. Further, the second flow dividing plate 32 can be expected to damp the vibration of the moving blade 13 and can be used to adjust the vibration mode of the blade to avoid resonance.
　次に、本発明の第３の実施形態について図面を用いて説明する。図１４は、本発明の第３の実施形態に係る蒸気タービンのタービン段落部の要部構造を表す子午面断面図である。なお、先に説明した実施形態と同等の構成要素には同一の符号を付し、説明を省略する。 Next, a third embodiment of the present invention will be described using the drawings. FIG. 14 is a meridional cross-sectional view showing a main part structure of a turbine stage portion of a steam turbine according to a third embodiment of the present invention. In addition, the same code | symbol is attached | subjected to the component equivalent to embodiment described previously, and description is abbreviate | omitted.
　本実施形態は、図１２に示した第２の実施形態の応用例であり、第１の分流板２６と第２の分流板３２との間にシール構造３３を設け、分流板間に蒸気の漏れ流れが生じることを抑制している。 The present embodiment is an application example of the second embodiment shown in FIG. 12, and a seal structure 33 is provided between the first flow dividing plate 26 and the second flow dividing plate 32, and steam is transmitted between the flow dividing plates. It is suppressing that a leak flow arises.
　図１５は、図１４に示した蒸気タービンのシール構造３３の拡大図である。本実施形態では、第１の分流板２６は、最終段落の静・動翼間流路３４に、動翼１３に向かって延伸する延長部３５を有し、延長部３５の内周側にシールフィン３６を設けている。一方、第２の分流板３２も、延長部３５の内周側で静翼１２に向かって延伸する、延長部３７を有する。延長部３５と延長部３７は、互いに半径方向に並設している。 FIG. 15 is an enlarged view of the seal structure 33 of the steam turbine shown in FIG. In the present embodiment, the first flow dividing plate 26 has an extending portion 35 extending toward the moving blade 13 in the inter-static-moving blade channel 34 in the final paragraph, and seals the inner peripheral side of the extending portion 35 Fins 36 are provided. On the other hand, the second flow dividing plate 32 also has an extension 37 that extends toward the stationary blade 12 on the inner peripheral side of the extension 35. The extension 35 and the extension 37 are radially juxtaposed to each other.
　本実施の形態によれば、第２の実施形態の効果に加えて、シール構造３３を設けて、分流板間に蒸気の漏れ流れが生じることを抑制できるので、最終段落の静・動翼間の内周側３８と、外周側３９の圧力を翼長方向に連続する必要も無くなり、最終段落は内周側と外周側で完全に独立して段落設計することができる。図１６に、本実施形態の蒸気タービン最終段落と、その１つ上流側段落の翼高さ方向の、比エンタルピー分布を示す。第１の実施形態と異なり、本実施形態では、最終段落の静・動翼間比エンタルピーが内周側ｈ３と外周側ｈ５との接続部で不連続となっている。 According to the present embodiment, in addition to the effects of the second embodiment, the seal structure 33 can be provided to suppress the occurrence of the leak flow of the steam between the diverting plates. There is no need to continue the pressure on the inner circumferential side 38 and the outer circumferential side 39 in the blade length direction, and the final paragraph can be designed completely independently on the inner circumferential side and the outer circumferential side. FIG. 16 shows the specific enthalpy distribution in the final stage of the steam turbine of this embodiment and in the blade height direction in the upstream stage of one stage thereof. Unlike the first embodiment, in this embodiment, the static / moving blade specific enthalpy in the final paragraph is discontinuous at the connection portion between the inner circumferential side h3 and the outer circumferential side h5.
　以上、実施例１乃至３に示した実施形態は、最終段落出口で単位面積当たりの流量が等しいという条件で段落設計を行い、その特性を示している。 As mentioned above, the embodiment shown in Example 1 thru | or 3 performs paragraph design on the conditions that the flow volume per unit area is equal at the last paragraph exit, and has shown the characteristic.
　次に、本発明の第４の実施形態について図面を用いて説明する。本実施形態は、第１の実施形態の応用例であり、基本的な構造は、図７に示した第１の実施形態と同一である。 Next, a fourth embodiment of the present invention will be described using the drawings. The present embodiment is an application example of the first embodiment, and the basic structure is the same as the first embodiment shown in FIG.
　本実施形態では、最終段の内周側と外周側とで、単位面積当たりの流量を変えることで、図９に示すような、比エンタルピー分布で設計しても、図１７に示すように、動翼入口角を連続とすることが可能である。本実施形態では、最終段落の内周側流路４１（段落流路３０）の単位面積当たりの流量に対し、最終段落の外周側流路４２（バイパス流路２９）の単位面積当たりの流量を小さく設定することで、動翼入口角を連続とした。 In this embodiment, even if the specific enthalpy distribution as shown in FIG. 9 is designed by changing the flow rate per unit area between the inner peripheral side and the outer peripheral side of the final stage, as shown in FIG. The blade inlet angle can be continuous. In this embodiment, the flow rate per unit area of the outer circumferential side flow passage 42 (bypass flow passage 29) of the final paragraph is set to the flow rate per unit area of the inner circumference side flow passage 41 (paragraph flow passage 30) of the final paragraph. The blade inlet angle was made continuous by setting it small.
　なお、図７の段落流路３０内のタービン段落を、単位面積当たりの流量が等しいという条件で段落設計を行っている場合、最終段落の内周側と外周側とで単位面積当たりの流量を変えるには、最終段落の内周側と外周側の流量比と、段落流路３０およびバイパス流路２９を流れる流量比が等しくなるように、分流板２６の上流側端部の位置を決める。 When the paragraph design is performed on the condition that the flow rate per unit area is equal, the flow rate per unit area of the turbine paragraph in the paragraph flow path 30 of FIG. In order to change, the position of the upstream end of the flow dividing plate 26 is determined so that the flow ratio between the inner circumference and the outer circumference of the final paragraph and the flow ratio flowing through the paragraph flow passage 30 and the bypass flow passage 29 become equal.
　本実施形態においても、第１の実施形態と同様の効果が得られる。 Also in this embodiment, the same effect as that of the first embodiment can be obtained.
　次に、本発明の第５の実施形態について図面を用いて説明する。図１８は、本発明の第５の実施形態に係る蒸気タービンのタービン段落部の要部構造を表す子午面断面図である。なお、図７と同一符号は、同一部分を示している。 Next, a fifth embodiment of the present invention will be described using the drawings. FIG. 18 is a meridional cross-sectional view showing a main part structure of a turbine stage portion of a steam turbine according to a fifth embodiment of the present invention. The same reference numerals as in FIG. 7 denote the same parts.
　本実施形態では、蒸気主流路２内に第１の実施形態と同様の分流板をタービン半径方向に複数設けて、蒸気主流路２を翼高さ方向に複数のブロック４６，４７，４８，４９に区分けし、外周側ほどより上流側のタービン段落からのバイパス蒸気が動翼１３に流入するようにしている。 In the present embodiment, a plurality of flow dividing plates similar to those of the first embodiment are provided in the steam main flow passage 2 in the radial direction of the turbine, and the steam main flow passage 2 is divided into a plurality of blocks 46, 47, 48, 49 in the blade height direction. The bypass steam from the upstream turbine stage is made to flow into the moving blades 13 toward the outer peripheral side.
　まず最内周側のブロック４９は、分流板４５と蒸気主流路２の内周側壁面との間に形成される。分流板４５は、最終段落の静翼１３に支持されている。分流板４５上流側端部の内周側には、最終段落の１つ上流側のタービン段落を構成する静翼５０と動翼５１が設けられている。静翼５０は、分流板４５と内周側ダイアフラム５３との間に固定されている。分流板４５は、蒸気主流路２の外周側壁面を構成する外周側固定部材５２から伸びるサポート部材３１によって支持されている。なお、動翼５１の外周側先端には、シュラウド部材５４が設けられている。 First, the block 49 on the innermost circumferential side is formed between the diverting plate 45 and the inner circumferential side wall surface of the main steam passage 2. The flow dividing plate 45 is supported by the vanes 13 of the final stage. On the inner peripheral side of the upstream end portion of the flow dividing plate 45, a stationary blade 50 and a moving blade 51 that constitute a turbine stage one upstream side of the final stage are provided. The vane 50 is fixed between the diverting plate 45 and the inner diaphragm 53. The flow dividing plate 45 is supported by a support member 31 extending from an outer peripheral side fixing member 52 constituting an outer peripheral side wall surface of the steam main flow channel 2. A shroud member 54 is provided at the outer peripheral tip of the moving blade 51.
　次にブロック４９の１つ外周側のブロック４８は、分流板４５と分流板４５の１つ外周側の分流板４４との間に形成される。分流板４４は、静翼１３に支持されている。分流板４４の上流側端部の内周側には、最終段落の２つ上流側のタービン段落を構成する静翼５５と動翼５６が設けられている。静翼５５は、分流板４４と内周側ダイアフラム５７との間に固定されている。分流板４４は、蒸気主流路２の外周側壁面を構成する外周側固定部材５８から伸びるサポート部材３１によって支持されている。動翼５６の外周側先端にもシュラウド部材５９が設けられている。 Next, one block 48 on the outer peripheral side of the block 49 is formed between the diverter plate 45 and the diverter plate 44 on the outer peripheral side of the diverter plate 45. The flow dividing plate 44 is supported by the vanes 13. On the inner peripheral side of the upstream end portion of the flow dividing plate 44, there are provided a vane 55 and a moving blade 56 which constitute a two upstream turbine stage of the final stage. The vane 55 is fixed between the diverting plate 44 and the inner diaphragm 57. The flow dividing plate 44 is supported by a support member 31 extending from an outer peripheral fixing member 58 which constitutes an outer peripheral side wall surface of the steam main flow channel 2. A shroud member 59 is also provided at the outer circumferential tip of the moving blade 56.
　次にブロック４８の１つ上流側のブロック４７は、分流板４４と分流板４４の１つ外周側の分流板４３との間に形成される。分流板４３は、静翼１３に支持されている。分流板４３上流側端部の内周側には、最終段落の３つ上流側のタービン段落を構成する静翼６０と動翼６１が設けられている。静翼６０は、分流板４３と内周側ダイアフラム６２との間に固定されている。分流板４３は、蒸気主流路２の外周側壁面を構成する外周側固定部材６３から伸びるサポート部材３１によって支持されている。動翼６１の外周側先端にもシュラウド部材６７が設けられている。 Next, one upstream block 47 of the block 48 is formed between the diverter plate 44 and the diverter plate 43 on the outer peripheral side of the diverter plate 44. The flow dividing plate 43 is supported by the vanes 13. On the inner peripheral side of the upstream end portion of the flow dividing plate 43, there are provided a stator blade 60 and a rotor blade 61 which constitute three upstream turbine stages of the final paragraph. The vane 60 is fixed between the diverting plate 43 and the inner diaphragm 62. The flow dividing plate 43 is supported by a support member 31 extending from an outer peripheral side fixing member 63 which constitutes an outer peripheral side wall surface of the steam main flow channel 2. A shroud member 67 is also provided at the outer circumferential tip of the moving blade 61.
　最外周側のブロック４６は、分流板４３と蒸気主流路２外周側壁面との間に形成されている。 The block 46 on the outermost side is formed between the flow dividing plate 43 and the outer peripheral side wall surface of the main steam channel 2.
　本実施形態の構成によれば、外周側に行くほどより上流側からバイパスされた蒸気が最終段落に流入する。 According to the configuration of the present embodiment, the vapor bypassed from the upstream side flows into the final paragraph as it goes to the outer peripheral side.
　図１９に本実施形態に係る蒸気タービン最終段落部の翼高さ方向の比エンタルピー分布を示す。本発明のもっとも理想的な最終段落入口の全比エンタルピーの翼高さ方向分布は、図５に示すとおり連続的に外周側の全比エンタルピーＨ０が大きくなる分布である。本実施形態によれば、図１８に示すように、最終段落を翼高さ方向に複数のブロックに分け、外周側のブロックほどより上流側のタービン段落からのバイパス蒸気を用いることで、図１９に示すような、翼長方向外周側に向かって階段状に全比エンタルピーＨ０が大きくなる分布が得られる。本実施形態によれば、翼長外周側の静翼の比エンタルピー差Δｈを大きくすることで静翼流出速度を大きくできる。よって、静翼流出速度を大きくできるので、動翼相対流入速度を抑制でき、第１の実施形態と同様の効果を得ることができる。 FIG. 19 shows the specific enthalpy distribution in the blade height direction of the final stage portion of the steam turbine according to the present embodiment. The distribution in the height direction of the total specific enthalpy of the most ideal final stage inlet of the present invention is a distribution in which the total specific enthalpy H0 on the outer peripheral side increases continuously as shown in FIG. According to the present embodiment, as shown in FIG. 18, the final paragraph is divided into a plurality of blocks in the blade height direction, and by using the bypass steam from the turbine paragraph on the more upstream side of the block on the outer circumferential side, FIG. A distribution is obtained in which the total specific enthalpy H0 increases in a stepwise manner toward the outer peripheral side in the span direction as shown in FIG. According to the present embodiment, the stator blade outflow velocity can be increased by increasing the specific enthalpy difference Δh of the vanes on the outer peripheral side of the blade length. Therefore, since the stator blade outflow velocity can be increased, the rotor relative inflow velocity can be suppressed, and the same effect as that of the first embodiment can be obtained.
　また、本実施形態によれば、外周側に行くほど動翼の周速が大きくなるのに合わせて、静翼の比エンタルピー差Δｈを大きくして静翼流出速度を大きくすることで、動翼に対する相対流入角の変化を小さくすることができる。この場合、前述した損失が急激に大きくならない入射角範囲を考慮すると、翼長が大きいにも係らず、例えば動翼入口角９０と一定でねじれが無く、かつ効率の良い動翼を用いることが可能となる。ねじれ角のない動翼は、局所応力が大きくならないという利点の他に、製作も簡単となる利点がある。 Further, according to the present embodiment, the moving blade can be increased by increasing the specific enthalpy difference Δh of the stationary blades to increase the flowing velocity of the stationary blades in accordance with the increase in the peripheral velocity of the moving blades toward the outer peripheral side. Changes in the relative inflow angle with respect to In this case, in consideration of the incident angle range in which the above-mentioned loss does not increase sharply, for example, even if the blade length is large, using a moving blade having a constant, non-twisting with the blade inlet angle 90 and having high efficiency It becomes possible. In addition to the advantage that local stresses do not increase, blades having no twist angle also have the advantage of being easy to manufacture.
　次に、本発明の第６の実施形態について図面を用いて説明する。図２０は、本発明の第６の実施形態に係る蒸気タービンのタービン段落部の要部構造を表す子午面断面図である。なお、第１の実施形態と同等の構成要素には同一の符号を付し、説明を省略する。 Next, a sixth embodiment of the present invention will be described using the drawings. FIG. 20 is a meridional cross-sectional view showing a main part structure of a turbine stage portion of a steam turbine according to a sixth embodiment of the present invention. In addition, the same code | symbol is attached | subjected to a component equivalent to 1st Embodiment, and description is abbreviate | omitted.
　本実施形態は、図１に示した第１の実施形態の応用例であり、第１の実施形態と異なるのは、バイパス流路２９の外周側に、矢印８１で示したように、蒸気の一部をタービン外部に抽気する抽気スリット６４を設けた点にある。 The present embodiment is an application example of the first embodiment shown in FIG. 1 and differs from the first embodiment in that, as indicated by the arrow 81 on the outer peripheral side of the bypass flow passage 29, It is in the point which provided the extraction slit 64 which extracts a part to the turbine exterior.
　抽気スリットは、一般のタービンではタービンロータ全長の制約から軸方向には小さい開口長さで設けられる。そのため、抽気流量が多い場合には、抽気スリットで流速が大きくなり損失が増大する。一方、図２０に示す本実施形態では、バイパス流路２９として軸方向に一段落分の大きな空間があり、軸方向に大きく開口する抽気スリット６４を設けることができる。そのため、抽気流量が増加した場合でも、抽気流８１の流速を小さくすることができ、抽気スリットでの損失の増加を抑制することができる。 The bleed slit is provided with a small opening length in the axial direction due to the restriction of the overall length of the turbine rotor in a general turbine. Therefore, when the amount of bleed air flow is large, the flow velocity increases at the bleed slit and the loss increases. On the other hand, in the present embodiment shown in FIG. 20, there is a large space of one paragraph in the axial direction as the bypass flow passage 29, and a bleed slit 64 that opens wide in the axial direction can be provided. Therefore, even when the amount of bleed flow increases, the flow velocity of the bleed flow 81 can be reduced, and an increase in loss at the bleed slit can be suppressed.
　従って、本実施形態によれば、第１の実施形態と同様の効果が得られるのに加えて、抽気スリットでの損失の増加を抑制することができる。 Therefore, according to this embodiment, in addition to the same effect as that of the first embodiment can be obtained, it is possible to suppress an increase in loss in the bleed slit.
　以上説明した各実施例は、低圧タービンの最終段落に適用した例であるが、本発明は、低圧タービン、および最終段落に限定したものではなく、中間段落に適用しても本発明の効果は得ることができる。また本発明の効果は、蒸気，空気等の作動流体によらず有効である。 Although each embodiment described above is an example applied to the last paragraph of a low pressure turbine, the present invention is not limited to the low pressure turbine and the last paragraph, and the effect of the present invention can be applied to an intermediate stage. You can get it. Further, the effect of the present invention is effective regardless of working fluid such as steam and air.
　なお、本発明を蒸気タービンの最終段落に適用した場合には、先に各実施形態で説明した利点の他に、以下の二つの利点がある。 When the present invention is applied to the final paragraph of the steam turbine, the following two advantages are obtained in addition to the advantages described in the above embodiments.
　一つ目の利点は、湿り損失が低減することである。図１に示した一般的な蒸気タービンでは、最終段落の上流側段落を構成する動翼７翼面に付着した水膜が遠心力により外周側に集められ、最終段落の静翼１２に向かって放出される。そのため、最終段落入口の外周側で湿り度が大きくなり、これが動翼周速の大きい最終段落での湿り損失増加や、エロージョン増加の原因となる。一方、本発明を蒸気タービンの最終段落に適用した場合、最終段落外周側の入口全比エンタルピーが大きいために、液相の質量分率である湿り度が小さ
くなる。 The first advantage is the reduction of wetness loss. In the general steam turbine shown in FIG. 1, the water film attached to the blade surface of the moving blade 7 constituting the upstream side paragraph of the final paragraph is collected on the outer peripheral side by the centrifugal force, and is directed to the vane 12 of the final paragraph. Released. Therefore, the degree of wetness increases on the outer peripheral side of the final stage inlet, which causes an increase in wet loss and an increase in erosion in the final stage where the moving blade circumferential speed is large. On the other hand, when the present invention is applied to the final stage of the steam turbine, the degree of wetness, which is the mass fraction of the liquid phase, decreases because the total inlet enthalpy on the outer side of the final stage is large.
　湿り度が小さくなる結果、本発明では湿り損失が小さくなり、エロージョンの発生も抑制できる。そのため、タービン効率を向上でき、蒸気タービンの信頼性も向上できる。 As a result of the decrease in the degree of wetness, in the present invention, the wet loss is reduced and the occurrence of erosion can also be suppressed. Therefore, the turbine efficiency can be improved, and the reliability of the steam turbine can also be improved.
　二つ目の利点は、翼の信頼性を向上できることである。蒸気タービンの過熱蒸気から二相流状態である湿り蒸気に移行するウイルソン線は、最終段落の１つ上流側のタービン段落に位置することが多い。ウイルソン線は、タービン負荷や蒸気条件によって、流れ方向に動くため、ウイルソン線が存在するタービン段落では、乾き蒸気と湿り蒸気の状態が繰り返され、腐食ピットが発生しやすい。しかしながら、本発明を最終段落に適用した場合、ウイルソン線が生じる最終段落の１つ上流側のタービン段落は、翼長が小さいため、翼にかかる応力を小さくでき、腐食ピットによる翼の信頼性低下を抑制できる。 The second advantage is the ability to improve wing reliability. The Wilson line, which transitions from the superheated steam of the steam turbine to wet steam in a two-phase flow state, is often located one turbine stage upstream of the final stage. The Wilson line moves in the flow direction depending on the turbine load and steam conditions, so in the turbine stage where the Wilson line exists, the state of dry steam and wet steam is repeated and corrosion pits are likely to occur. However, when the present invention is applied to the final paragraph, the turbine stage one upstream side of the final paragraph where the Wilson line occurs has a small blade length, so the stress applied to the blade can be reduced and the reliability of the blade is reduced due to corrosion pits. Can be suppressed.
２　蒸気主流路
３　タービンケーシング
４，１０，１４，１８　外周側ダイアフラム
５，１１，１５，１９，５３，５７，６２　内周側ダイアフラム
６，１２，１６，２０，２３，２７，５０，５５，６０　静翼
７，１３，１７，２１，２４，２８，５１，５６，６１　動翼
８　タービンロータ
９　タービン中心軸
２６，４３，４４，４５　分流板　
２９　バイパス流路
３０　段落流路
３１　サポート部材
３２　第２の分流板
３３　シール構造
３５　延長部
３６　シールフィン
３７　延長部
４０　分流板上流端
４１　内周側流路
４２　外周側流路
５２，５８，６３　外周側固定部材
６４　抽気スリット 2 steam main flow path 3 turbine casings 4, 10, 14, 18 outer peripheral side diaphragms 5, 11, 15, 19, 53, 57, 62 inner peripheral side diaphragms 6, 12, 16, 20, 23, 27, 27, 50, 55, 60 stator vanes 7, 13, 17, 21, 24, 28, 51, 56, 61 moving blades 8 turbine rotors 9 turbine center shaft 26, 43, 44, 45 divided flow plate
29 bypass flow passage 30 flow passage 31 support member 32 second flow dividing plate 33 seal structure 35 extension part 36 seal fin 37 extension part 40 flow dividing plate upstream end 41 inner peripheral side flow passage 42 outer peripheral side flow passage 52, 58, 63 Outer peripheral side fixing member 64 bleed slit

Claims (14)

  1. 　静止体に固定された静翼と、タービンロータに固定された動翼とからなるタービン段落と、
    　前記タービン段落をタービン軸方向に複数有する作動流体流路とを備える軸流タービンであって、
    　前記作動流体流路中に設けられ、作動流体流れ方向上流側から流入する作動流体の一部を、前記タービン段落の少なくとも一段の外周側をバイパスさせ、バイパスした前記タービン段落の作動流体流れ方向下流側にある前記動翼に導入するタービン段落バイパス流路を備えることを特徴とする軸流タービン。     A turbine stage comprising a stationary blade fixed to a stationary body and a moving blade fixed to a turbine rotor;
    An axial flow turbine comprising: a working fluid flow passage having a plurality of the turbine stages in the axial direction of the turbine;
    A portion of the working fluid provided in the working fluid flow path and flowing in from the upstream side of the working fluid flow direction is bypassed by at least the outer peripheral side of at least one stage of the turbine stage, and the working fluid flow direction downstream of the turbine stage bypassed An axial flow turbine comprising a turbine stage bypass flow passage introduced to the moving blade located on the side.
  2. 　請求項１記載の軸流タービンであって、
    　前記バイパスしたタービン段落の動翼出口流路高さが、上流側段落の動翼出口流路高さと比較して小さく、かつ下流側段落の動翼出口流路高さと比較しても小さいことを特徴とする軸流タービン。     An axial flow turbine according to claim 1, wherein
    The blade outlet channel height of the bypassed turbine stage is smaller compared to the blade outlet channel height of the upstream stage, and smaller than the blade outlet channel height of the downstream stage Features axial flow turbines.
  3. 　請求項１記載の軸流タービンであって、
    　前記タービン段落バイパス流路は、
    　前記静止体と、
    　前記作動流体流路中に設けられ、前記作動流体流路を流下する作動流体の流れをタービン半径方向に二分割し、内周側に前記タービン段落を有する分流板との間に設けられていることを特徴とする軸流タービン。     An axial flow turbine according to claim 1, wherein
    The turbine stage bypass flow path is
    Said stationary body,
    The flow of working fluid provided in the working fluid flow path and flowing down the working fluid flow path is divided in two in the radial direction of the turbine, and provided between the flow dividing plate having the turbine stage on the inner circumferential side An axial flow turbine characterized by
  4. 　請求項３記載の軸流タービンであって、前記タービン段落バイパス流路を流下した作動流体が導入される前記動翼は、前記作動流体を前記タービン段落バイパス流路を流下した流れと、その他の流れとに分流する第２の分流板を備えることを特徴とする軸流タービン。 The axial flow turbine according to claim 3, wherein the moving blade into which the working fluid having flowed down the turbine stage bypass flow path is introduced is the flow having flowed the working fluid down the turbine stage bypass flow path, and the like. An axial flow turbine comprising: a second flow dividing plate which is branched into a flow and a flow.
  5. 　請求項４記載の軸流タービンであって、
    　前記分流板と前記第２の分流板とは、互いに間隙を空けて並行する延長部をそれぞれ有し、
    　前記延長部のいずれか一方または両方にシール装置を設けたことを特徴とする軸流タービン。     The axial flow turbine according to claim 4, wherein
    The diverting plate and the second diverting plate respectively have extensions extending parallel to each other,
    An axial flow turbine provided with a seal device in one or both of the extensions.
  6. 　請求項３記載の軸流タービンであって、
    　前記タービン段落バイパス流路の外周側に、前記タービン段落バイパス流路を流下する作動流体の一部を抽気し、タービン外部へ供給する抽気手段を備えることを特徴とする軸流タービン。     The axial flow turbine according to claim 3, wherein
    An axial flow turbine comprising an extraction means for extracting a part of the working fluid flowing down the turbine stage bypass channel and supplying it to the outside of the turbine on the outer peripheral side of the turbine stage bypass channel.
  7. 　静止体に固定された静翼と、タービンロータに固定された動翼とからなるタービン段落と、
    　前記タービン段落をタービン軸方向に複数有する作動流体流路とを備える軸流タービンであって、
    　第１のタービン段落と、
    　該第１のタービン段落の作動流体流れ方向上流側に設けられた第２のタービン段落と、
    　前記作動流体流路中に設けられ、前記第１のタービン段落の前記静翼に支持され、前記第１のタービン段落から作動流体流れ方向上流側に延伸する円環状の分流板とを備え、
    　前記第２のタービン段落は、前記分流板の内周側に設けられていることを特徴とする軸流タービン。     A turbine stage comprising a stationary blade fixed to a stationary body and a moving blade fixed to a turbine rotor;
    An axial flow turbine comprising: a working fluid flow passage having a plurality of the turbine stages in the axial direction of the turbine;
    With the first turbine stage,
    A second turbine stage provided upstream in the working fluid flow direction of the first turbine stage;
    And an annular diverting plate provided in the working fluid flow path, supported by the stator vanes of the first turbine stage, and extending from the first turbine stage toward the working fluid flow direction,
    An axial flow turbine, wherein the second turbine stage is provided on the inner peripheral side of the flow dividing plate.
  8. 　請求項７記載の軸流タービンであって、
    　前記第２のタービン段落の動翼出口流路高さが、上流側段落の動翼出口流路高さと比較して小さく、かつ下流側段落の動翼出口流路高さと比較しても小さいことを特徴とする軸流タービン。     The axial flow turbine according to claim 7, wherein
    The blade outlet channel height of the second turbine stage is smaller compared to the blade outlet channel height of the upstream stage, and smaller than the blade outlet channel height of the downstream stage Axial flow turbine characterized by
  9. 　請求項７記載の軸流タービンであって、
    　前記第１のタービン段落は、前記分流板の下流側であって、周方向に隣接する動翼間に設けられた第２の分流板を備えていることを特徴とする軸流タービン。     The axial flow turbine according to claim 7, wherein
    The axial flow turbine according to claim 1, wherein the first turbine stage includes a second flow dividing plate provided between the moving blades adjacent to each other in the circumferential direction on the downstream side of the flow dividing plate.
  10. 　請求項９記載の軸流タービンであって、
    　前記分流板と前記第２の分流板は、互いに間隙を空けて並行する延長部をそれぞれ有し、
    　前記延長部のいずれか一方または両方にシール装置を設けたことを特徴とする軸流タービン。     The axial flow turbine according to claim 9, wherein
    The diverting plate and the second diverting plate have respective extensions parallel to and spaced from each other,
    An axial flow turbine provided with a seal device in one or both of the extensions.
  11. 　請求項７記載の軸流タービンであって、
    　前記タービン段落バイパス流路の外周側に、タービン外部に連通する抽気スリットを備えることを特徴とする軸流タービン。     The axial flow turbine according to claim 7, wherein
    An axial flow turbine comprising an extraction slit communicating with the outside of the turbine on an outer peripheral side of the turbine stage bypass flow passage.
  12. 　　静止体に固定された静翼と、タービンロータに固定された動翼とからなるタービン段落と、
    　前記タービン段落をタービン軸方向に複数有する作動流体流路とを備える軸流タービンであって、
    　第１のタービン段落と、
    　該第１のタービン段落の作動流体流れ方向上流側に設けられる複数のタービン段落と、
    　前記作動流体流路中に設けられ、前記第１のタービン段落の前記静翼に支持され、前記第１のタービン段落から作動流体流れ方向上流側に向かって延伸する円環状の複数の分流板とを備え、
    　前記複数の分流板は、タービン半径方向に一定間隔を置いて設置され、タービン半径方向内周側から外周側へ設置順に延伸距離が長くなるように構成されており、
    　前記複数の分流板の内周側には、それぞれ前記タービン段落が設けられていることを特徴とする軸流タービン。     A turbine stage comprising a stationary blade fixed to a stationary body and a moving blade fixed to a turbine rotor;
    An axial flow turbine comprising: a working fluid flow passage having a plurality of the turbine stages in the axial direction of the turbine;
    With the first turbine stage,
    A plurality of turbine stages provided upstream in the working fluid flow direction of the first turbine stage;
    A plurality of annular diverting plates provided in the working fluid flow path, supported by the vanes of the first turbine stage, and extending from the first turbine stage toward the working fluid flow direction upstream side; Equipped with
    The plurality of flow dividing plates are installed at regular intervals in the radial direction of the turbine, and are configured such that the extension distance becomes longer in the order of installation from the inner peripheral side to the outer peripheral side of the turbine radial direction,
    The axial flow turbine, wherein the turbine stage is provided on the inner circumferential side of the plurality of flow dividing plates.
  13. 　請求項７または１２記載の軸流タービンであって、
    　前記第１のタービン段落は、低圧タービンの最終段落であり、前記作動流体は蒸気であることを特徴とする軸流タービン。     An axial flow turbine according to claim 7 or 12, wherein
    The first turbine stage is the final stage of a low pressure turbine, and the working fluid is steam.
  14. 　静止体に固定された静翼と、タービンロータに固定された動翼とからなるタービン段落を備え、
    　前記タービン段落をタービン軸方向に複数設けた多段落構造を有する軸流タービンの設計方法であって、
    　前記タービン段落入口における、単位質量当たりのエンタルピーと作動流体流速の二乗を２で割った単位質量当たりの運動エネルギーとの和である比全エンタルピーが、タービン半径方向外周側に向かって大きくなるように前記多段落構造を設計することを特徴とする軸流タービンの設計方法。     A turbine stage comprising a stationary blade fixed to a stationary body and a moving blade fixed to a turbine rotor,
    A design method of an axial flow turbine having a multi-stage structure in which a plurality of the turbine stages are provided in the axial direction of the turbine,
    The specific total enthalpy, which is the sum of the enthalpy per unit mass and the kinetic energy per unit mass divided by 2 at the inlet of the turbine stage divided by the square of the working fluid flow velocity, increases toward the radially outer side of the turbine A design method of an axial flow turbine characterized by designing the multi-stage structure.
PCT/JP2010/069086 2009-11-06 2010-10-27 Axial flow turbine WO2011055666A1 (en)

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Cited By (2)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
EP2540967A3 (en) * 2011-06-29 2017-06-21 Mitsubishi Hitachi Power Systems, Ltd. Supersonic turbine moving blade and axial-flow turbine
CN115003898A (en) * 2020-01-31 2022-09-02 三菱重工业株式会社 Turbine engine

Families Citing this family (3)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
WO2013027239A1 (en) * 2011-08-24 2013-02-28 株式会社 日立製作所 Axial flow turbine
JPWO2013027239A1 (en) * 2011-08-24 2015-03-05 三菱日立パワーシステムズ株式会社 Axial flow turbine
JP5677332B2 (en) * 2012-01-23 2015-02-25 株式会社東芝 Steam turbine

Citations (4)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US1493266A (en) * 1922-12-05 1924-05-06 Gen Electric Elastic-fluid turbine
FR731766A (en) * 1931-03-24 1932-09-08 Ljungstroms Angturbin Ab Device applicable to sets of axial vanes of gas or steam turbines
JPS5322904A (en) * 1976-08-13 1978-03-02 Guruupu Uuropeen Puuru Ra Tech Turbine for use in compressible fluid
JPH0419303A (en) * 1990-05-10 1992-01-23 Toshiba Corp Steam turbine stage structure

Patent Citations (4)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US1493266A (en) * 1922-12-05 1924-05-06 Gen Electric Elastic-fluid turbine
FR731766A (en) * 1931-03-24 1932-09-08 Ljungstroms Angturbin Ab Device applicable to sets of axial vanes of gas or steam turbines
JPS5322904A (en) * 1976-08-13 1978-03-02 Guruupu Uuropeen Puuru Ra Tech Turbine for use in compressible fluid
JPH0419303A (en) * 1990-05-10 1992-01-23 Toshiba Corp Steam turbine stage structure

Cited By (4)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
EP2540967A3 (en) * 2011-06-29 2017-06-21 Mitsubishi Hitachi Power Systems, Ltd. Supersonic turbine moving blade and axial-flow turbine
EP3828387A1 (en) * 2011-06-29 2021-06-02 Mitsubishi Power, Ltd. Turbine moving blade and axial-flow turbine
EP3832068A1 (en) * 2011-06-29 2021-06-09 Mitsubishi Power, Ltd. Turbine moving blade and axial-flow turbine
CN115003898A (en) * 2020-01-31 2022-09-02 三菱重工业株式会社 Turbine engine

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