JP2007056824A - Stationary blade and moving blade for axial flow turbine, and axial flow turbine provided with same - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a stationary blade, a rotor blade for an axial flow turbine capable of improving the whole performance (turbine efficiency) of the axial flow turbine, and the axial flow turbine provided with the same. <P>SOLUTION: In the stationary blade of the axial flow turbine, each stationary blade 3 is formed with being bent to project a trailing edge 3c out of a circumference direction blade face side surface as indicated in Fig. 2. A shape of each stationary nozzle 3 is established in such a manner that ratio δc<SB>s</SB>/t<SB>s</SB>of circumference direction projection quantity δc<SB>s</SB>of the trailing edge 3c to circumference direction blade root pitch t<SB>s</SB>reduces as ratio H<SB>s</SB>/t<SB>s</SB>of blade height H<SB>s</SB>to circumference direction blade root pitch t<SB>s</SB>increases. Each rotor blade 5 is formed with being bent to project a trailing edge 5c out of circumference direction blade face side surface. A shape of each rotor blade 5 is established in such a manner that ratio δc<SB>r</SB>/t<SB>r</SB>of circumference direction projection quantity δc<SB>r</SB>of the trailing edge 5c to circumference direction blade root pitch t<SB>r</SB>reduces as ratio H<SB>r</SB>/t<SB>r</SB>of blade height H<SB>r</SB>to circumference direction blade root pitch t<SB>r</SB>increases. <P>COPYRIGHT: (C)2007,JPO&INPIT

Description

  The present invention relates to an axial flow turbine including a stationary blade and a moving blade, and more particularly to an axial flow turbine capable of improving turbine efficiency.

  In recent years, with respect to axial turbines used in power plants, ensuring reliability and increasing efficiency have become important issues from the viewpoint of environmental problems and energy saving.

  A general axial turbine will be described by taking the steam turbine shown in FIG. 1 as an example. FIG. 1 schematically shows an axial section of a steam turbine.

  The rotating shaft (turbine rotor) 4 is provided with a plurality of disc-shaped rotor disks 4a integrally with the shaft at predetermined intervals in the axial direction. In addition, several tens of moving blades 5 are usually planted at predetermined intervals in the circumferential direction of the rotor disks 4a. A shroud 6 is provided at the tip of the rotor blade 5 to prevent vibrations.

  On the other hand, between each rotor disk 4a and the moving blade 5, a stationary blade 3 held between a stationary blade outer ring 1 and a stationary blade inner ring 2 supported by a turbine casing 9 is disposed. A pair of these stationary blades 3 and moving blades 5 forms a turbine stage, and the steam turbine equipment is configured by providing these turbine stages in a plurality of stages in the axial direction of the rotating shaft 4.

  The steam as the working fluid flows (from left to right in the drawing) in the annular flow path P formed by these paragraphs (the stationary blade 3 and the moving blade 5). Then, the steam whose direction of flow is changed by the stationary blades flows in the moving blades, whereby the kinetic energy is converted into the rotational energy of the moving blades and then circulates in the paragraphs. The rotating shaft 4 having obtained the rotational energy rotates a generator (not shown) to generate electric power, and the lost steam is exhausted from the final stage.

  As described above, the working fluid is an energy source for generating electric power, and ideally, all the energy of the working fluid is converted into the rotational energy of the rotating shaft 4. However, a physical gap is required between the rotating rotating shaft 4 and the moving blade 5 and the stationary stationary blade 3 (and the turbine casing 9), but the working fluid flowing therethrough is the moving blade. Therefore, it does not contribute to the rotational energy of the rotating shaft 4. For this reason, it is essential to reduce the working fluid passing through the gap as much as possible. For the purpose of reducing the working fluid, the stationary blade inner ring 2 and the rotating shaft are disposed between the tip of the moving blade (the shroud 6) and the stationary blade outer ring 1. The labyrinth fins 7 and 8 are provided between each of them 4 and 4 to reduce the working fluid passing therethrough.

  Next, the loss which generate | occur | produces in the channel | path part of the axial flow turbine 50 is demonstrated. The loss generated in the turbine passage part is the blade type loss due to the blade shape of the stationary blade 53 or the moving blade 55 and the inner and outer end wall portions (53a, 53b, 55a, 55b vicinity) of the stationary blade 53 or the moving blade 55. Secondary loss generated in the wall), chip leakage loss caused by steam leaking from the gap between the labyrinth fin 7 and the shroud 6, and steam from the gap between the labyrinth fin 8 and the rotating shaft 4 Is classified as labyrinth leakage loss caused by leakage.

Conventionally, in order to reduce the secondary loss generated in the inner and outer end walls of the stationary blade 53 and the moving blade 55, the stationary blade 53 and the moving blade 55 are curved so that the trailing edges 53c and 55c protrude in the circumferential direction. As shown in FIG. 8, when the circumferential blade root pitch is t and the circumferential protrusion amount of the trailing edge is δc, as shown in FIG. A method of designing the stationary blade 53 and the moving blade 55 so that the ratio δc / t is constant regardless of the height H is known (for example, see Patent Document 3).
By making the ratio δc / t constant in the stationary blade 53 and the moving blade 55, the fluid outlet angle distribution at the blade outlet can be made equal regardless of the number of blades and the blade height H. The loss reduction effect of the design blade can be fully exhibited.

JP-A-3-189304 JP-A-6-81603 Japanese Patent No. 3397599

The loss generated in the passage portion of the axial turbine 50 is classified into the airfoil loss, the secondary loss, and the leakage loss as described above. Of these, the leakage loss is very small compared to other losses, and can be ignored. Among the above losses, the blade type loss due to the blade shape affects the overall blade row performance. Therefore, the performance of the entire axial flow turbine can be improved by using a blade row group having a small blade type loss.
FIGS. 9A and 9B show the relationship between the airfoil geometry of the stationary blade 53 and the airfoil loss obtained by the cascade wind tunnel test. Large parameters that determine the airfoil loss include the circumferential distance pitch t between adjacent stationary blades 53, the blade length code l, and the fluid outflow angle α. Further, a ratio t / l of the circumferential distance pitch t to the blade length code l (hereinafter also referred to as pitch / code ratio) is used as a dimensionless parameter of the airfoil geometry of the stationary blade 53. Further, the fluid outflow angle α is used as a parameter of the graph shown in FIG. The fluid outflow angle α is closely related to the blade load, and the blade load increases as the fluid outflow angle α decreases as shown in FIG. 9B. Here, as shown in FIG. 9B, the characteristics of the blade row of the stationary blade 53 increase the blade loss as the blade load increases (the fluid outflow angle α decreases). When the load is the same, that is, when the fluid outflow angle α is the same, there exists a unique pitch / code ratio (t / l) at which the airfoil loss becomes a minimum value. However, the airfoil loss increases whether the actual pitch / code ratio is larger or smaller than this inherent pitch / code ratio.

On the other hand, the secondary loss is a loss generated in the blade inner and outer end wall portions (wall portions near 53a, 53b, 55a, 55b) of the stationary blade 53 and the moving blade 55. As a large geometric parameter for determining the secondary loss, the ratio of the blade height H to the blade length code l, that is, the blade height / blade length code ratio (H / l, also referred to as aspect ratio) can be mentioned. 10A and 10B are graphs showing the relationship between the aspect ratio and the magnitude of the secondary loss. The solid line in the graph of FIG. 10B represents the relationship between the aspect ratio and the secondary loss when the blade height code l is constant and the blade height H is changed to change the aspect ratio. On the other hand, the broken line portion in the graph of FIG. 10B represents the relationship between the aspect ratio and the secondary loss when the blade height code L is changed and the aspect ratio is changed by changing the blade length code l. Yes. 10A and 10B, the unit of the blade height H and the blade length code l is inch (inch).
As shown in the graph of FIG. 10B, it can be seen that the secondary loss changes more greatly when the blade height H is changed than when the blade length code l is changed. In addition, when the aspect ratio is small, the increase or decrease of the secondary loss becomes significant.

In a general axial turbine, the pressure on the inlet side is set to several hundred kg / cm ² and the pressure on the exhaust side is reduced to a state close to vacuum. The size is several thousand times larger. This indicates that the blade height H is configured to increase from the inlet side toward the exhaust side. In a general axial turbine, the blade height on the inlet side is several tens of millimeters. On the other hand, the blade height on the exhaust side is about 800 to 1000 mm.

The relationship between the airfoil loss and the secondary loss in the stationary blade row will be described with reference to the graph of FIG. In the graph of FIG. 11, the horizontal axis represents the blade height ratio (h / H), and the vertical axis represents the loss factor. The broken line portion represents the airfoil loss, the shaded area represents the secondary loss, and the solid line portion represents the total cascade loss, which is the sum of the airfoil loss and the secondary loss.
As described above, the airfoil loss increases as the ratio s / t of the throat s to the circumferential distance pitch t of the stationary blade 53 decreases due to the correlation between the geometric shape of the stationary blade 53 and each loss. On the other hand, as shown in FIG. 10, the secondary loss increases as the aspect ratio (H / l) decreases. The blades constituting the axial flow turbine have a blade height of several tens of millimeters on the inlet side, whereas the blade height on the exhaust side varies greatly from about 800 to 1000 mm, and the aspect ratio also ranges from about 1 to 7-8. By changing to the extent, the secondary loss decreases from the inlet side to the exhaust side.

  For example, in the axial turbine shown in Patent Document 3 described above, by performing a three-dimensional design of the stationary blade and the moving blade, the secondary loss generated in the blade inner and outer end walls can be reduced by changing the load in the blade radial direction. Is going. Furthermore, in order to make the blade radial load distribution (fluid outflow angle distribution) of each turbine stage group consisting of a plurality of stages in the axial direction constant, the circumferential blade root regardless of the blade height or blade pitch size. The ratio δc / t of the circumferential protrusion amount δc of the trailing edge of the blade with respect to the pitch t is constant. FIGS. 12 and 13 are graphs showing the distribution of the fluid outflow angle α with the blade height direction (Span) of the conventional three-dimensional design blade shown in Patent Document 3 as the horizontal axis. As shown in the graphs of FIGS. 12 and 13, by making the ratio δc / t of the circumferential protrusion amount δc of the trailing edge of the blade with respect to the circumferential blade root pitch t constant, it depends on the size of the blade height H. The fluid outflow angle α has a substantially equal distribution.

However, in the conventional axial flow turbine, the performance of the entire turbine is optimized by keeping the fluid outflow angle distribution (α) of each stage of the axial flow turbine composed of a plurality of stages arranged in series constant. It is difficult. This will be described in detail below.
The total blade loss is expressed by the sum of the blade loss and the secondary loss, but the conventional three-dimensional design blade increases the load at the center of the blade to reduce the secondary loss (fluid outflow angle α). The load is reduced (fluid outflow angle α is increased) at the blade inner and outer end wall portions. In the airfoil loss shown in the graph of FIG. 9, the fluid outflow angle α is small in the central part of the blade, so the airfoil loss is large, and the fluid outflow angle α is large in the blade inner and outer end walls. Thus, the airfoil loss is reduced.
In addition, the secondary loss is reduced because the load on the inner and outer end walls of the blade is reduced. However, in terms of the performance of the entire blade cascade, the blade loss increased at the center of the blade and the second loss reduced at the inner and outer end walls of the blade. It is determined by comparison with the next loss.

  In order to improve the performance of the entire blade cascade in the axial turbine, it is necessary to make the reduction amount of the secondary loss larger than the increase amount of the airfoil loss at the blade center portion. As described above, this secondary loss decreases in inverse proportion as the aspect ratio increases (see FIG. 10). In particular, the secondary loss decreases as the blade height increases. Since the aspect ratio of the turbine stage constituting the axial turbine is 7 to 8 times larger on the exhaust side than on the inlet side, the ratio of the secondary loss and the airfoil loss to the total loss in each stage varies greatly. .

  FIG. 7 is a graph showing the relationship between the specific aspect ratio and each loss in the axial turbine cascade. In the graph of FIG. 7, the horizontal axis indicates the specific aspect ratio based on the aspect ratio of the inlet-side blade row, and the vertical axis indicates the specific blade row loss, which is the loss ratio when the inlet-side blade row total loss is 1. In the graph of FIG. 7, the broken line portion indicates the secondary loss ratio, and the alternate long and short dash line indicates the airfoil loss ratio. As shown in FIG. 7, the ratio of the total loss of the exhaust side blade row to the inlet side blade row is about 30% on the exhaust side on the inlet side. The airfoil loss occurs at a constant magnitude in all cascades with little relation to the specific aspect ratio, but the secondary loss decreases as the specific aspect ratio increases (especially as the blade height increases). When the specific aspect ratio is small (the turbine stage close to the inlet side), the secondary loss amount exceeds the blade type loss amount, and the secondary loss amount greatly decreases toward the exhaust side. Mold loss.

  As described above, in the three-dimensional design blade of the conventional axial flow turbine, the load at the blade central portion is increased and the load at the inner and outer end wall portions of the blade is decreased, thereby reducing the secondary loss and the performance of the entire blade row. And the ratio δc / t of the circumferential protrusion amount δc of the trailing edge of the blade with respect to the circumferential blade root pitch t is made constant, so that the blade radial load distribution is made equal in every paragraph. However, as shown in the graph of FIG. 7, since the secondary loss decreases from the inlet side toward the exhaust side, if the blade radial load distribution is the same in any paragraph, the paragraph with a large aspect ratio However, there is a problem that the total blade loss increases due to the increase in the blade loss at the center of the blade.

  The present invention has been made in consideration of such points, and the stationary blades and moving blades of an axial-flow turbine capable of improving the overall performance (turbine efficiency) of the axial-flow turbine, and the stationary blades and the moving blades. An object is to provide an axial turbine having blades.

The present invention provides a stationary blade of an axial-flow turbine provided in a plurality of stages in an annular flow path, wherein each stationary blade is formed to be curved so that a trailing edge protrudes toward the circumferential blade side. When the circumferential blade root pitch is t _s , the amount of protrusion of the trailing edge of each stationary blade to the circumferential blade surface is δc _s , and the blade height of each stationary blade is H _s , the circumferential direction of the stationary blade as the proportion H _{s /} t _s of blade height H _s of the vanes relative to the blade root pitch t _s is increased in the circumferential direction blade pressure side of the trailing edge of vanes with respect to the circumferential direction blade base pitch t _s of the stator blade it vanes of an axial flow turbine, wherein the ratio δc _{s /} t s of the protrusion amount .delta.c _s is set the shape of each vane to reduce.
According to stator vanes such axial flow turbine, the circumferential wings of the vane as the proportion H _{s /} t _s of blade height H _s of the vanes relative to the circumferential direction blade base pitch t _s of the stationary blade is increased since the shape of each vane so that the ratio .delta.c s _/ t _s of the protruding amount .delta.c _s to the edge of the circumferential blade pressure side after the vanes relative to the root pitch t _s is reduced is set, static for each paragraph An optimum load distribution can be formed on the blades, and the overall performance of the axial turbine can be improved.

Further, according to the present invention, in the rotor blade of the axial flow turbine provided in a plurality of stages in the annular flow path, each rotor blade is formed to be curved so that a trailing edge protrudes toward the circumferential blade side. circumferential blade root pitch t _r of the _blade, the .delta.c r a circumferential projection amount of the trailing edge of each rotor _blade, wherein when the blade height of the blades was set to H _r, blade circumferential blade root pitch projection amount δc in the circumferential direction blade pressure side of the blade trailing edge of the rotor blade with respect to the circumferential direction blade base pitch t _r as the proportion H _{r /} t _r of the blade of the blade height H _r for t _r increases it is moving blade of the axial flow turbine, characterized in that the shape of each blade is configured to proportion .delta.c r _/ t _r of _r is reduced.
According to the rotor blades of such axial flow turbine, the blades of the circumferential blade as the proportion H _{r /} t _r of the blade of the blade height H _r for the rotor blades in the circumferential direction the blade root pitch t _r increases since the shape of each blade such that the ratio δc _{r /} t r of the protrusion amount .delta.c _r in the circumferential direction blade pressure side of the blade trailing edge for the root pitch t _r decreases is set, the dynamic of each paragraph An optimum load distribution can be formed on the blades, and the overall performance of the axial turbine can be improved.

The present invention is an axial flow turbine comprising the stationary blade of the axial flow turbine and the moving blade of the axial flow turbine.
According to such an axial flow turbine, since the stationary blades and the moving blades having the shapes set as described above are provided, it is possible to form an optimum load distribution in both the stationary blades and the moving blades for each stage. Thus, the overall performance of the axial flow turbine can be improved more reliably.

In the stationary blade of the axial flow turbine of the present invention, it is preferable that the shape of each stationary blade is set so as to satisfy the following formula.
_{δc s / t s = C s} × (H s / t s) -βs
(However, C _s is a constant between 0 and 0.5, and β _s is a constant between 0.3 and 1.2.)
Moreover, in the moving blade of the axial-flow turbine of this invention, it is preferable that the shape of each moving blade is set so that the following formula may be satisfied.
_{δc r / t r = C r} × (H r / t r) -βr
(However, _Cr is a constant between 0 and 0.5, and _βr is a constant between 0.3 and 1.2.)
According to such a stationary blade or moving blade of an axial flow turbine, it is possible to form an optimum load distribution with higher accuracy for each stage, and it is possible to further improve the overall performance of the axial flow turbine. .

In the axial turbine according to the present invention, it is preferable that the stationary blade of the axial turbine satisfying the above equation and the moving blade of the axial turbine satisfying the above equation are provided.
According to such an axial flow turbine, since the stationary blades and the moving blades having the shapes set as described above are provided, an optimal load distribution is formed with higher accuracy in both the stationary blades and the moving blades in each stage. It is possible to improve the overall performance of the axial turbine.

According to vane of the axial flow turbine of the present invention, the circumferential wings of the vane as the proportion H _{s /} t _s of blade height H _s of the vanes relative to the circumferential direction blade base pitch t _s of the stationary blade is increased since the shape of each vane so that the ratio .delta.c s _/ t _s of the protruding amount .delta.c _s to the edge of the circumferential blade pressure side after the vanes relative to the root pitch t _s is reduced is set, static for each paragraph An optimum load distribution can be formed on the blades, and the overall performance of the axial turbine can be improved.

Further, according to the rotor blades of an axial flow turbine of the present invention, of the rotor blade as the proportion H _{r /} t _r of the blade of the blade height H _r for the rotor blades in the circumferential direction the blade root pitch t _r increases peripheral since the shape of each blade such that the ratio δc _{r /} t r of the protrusion amount .delta.c _r to the edge of the circumferential blade pressure side after the blades is decreased is set with respect to the direction the blade root pitch t _r, for each paragraph It is possible to form an optimum load distribution on the moving blades of the turbine and improve the overall performance of the axial turbine.

  Further, according to the axial flow turbine of the present invention, since the stationary blades and the moving blades having the shapes as described above are provided, an optimal load distribution is formed in both the stationary blades and the moving blades for each stage. This makes it possible to improve the overall performance of the axial turbine more reliably.

Embodiments of the present invention will be described below with reference to the drawings.
1 to 7 are diagrams showing an embodiment of an axial turbine according to the present invention. Among these, FIG. 1 is a cross-sectional view showing the configuration of the axial flow turbine, and FIG. 2 is a view taken along the line AA of FIG. 1 showing the configuration of the stationary blades of the axial flow turbine of the present embodiment. Figure 3 is a graph showing the relationship between the ratio δc _{s /} t s and the ratio H _{s /} t _s in the stationary blade of an axial flow turbine of FIG. 2, FIG. 4, the stator blade of the axial-flow turbine of FIG. 2 is a graph showing the relationship between the ratio δc _{s /} t s and axial deviation of operating efficiency of the turbine △ eta. Further, FIG. 5 is B-B in arrow view of FIG. 1 showing a blade structure of an axial flow turbine of the present embodiment, FIG. 6, the ratio of the rotor blades of an axial flow turbine of FIG. 5 .delta.c _r / t is a graph showing the relationship between _r and the ratio H _{r /} t _r, 7 is a graph showing the relationship between the ratio the aspect ratio and wing column losses in stator vanes of an axial flow turbine.

  A general axial turbine will be described by taking the steam turbine shown in FIG. 1 as an example. FIG. 1 schematically shows an axial section of a steam turbine.

  The rotating shaft (turbine rotor) 4 is provided with a plurality of disc-shaped rotor disks 4a integrally with the shaft at predetermined intervals in the axial direction. In addition, several tens of moving blades 5 are usually planted at predetermined intervals in the circumferential direction of the rotor disks 4a. A shroud 6 is provided at the tip of the rotor blade 5 to prevent vibrations.

  On the other hand, between each rotor disk 4a and the moving blade 5, a stationary blade 3 held between a stationary blade outer ring 1 and a stationary blade inner ring 2 supported by a turbine casing 9 is disposed. A pair of these stationary blades 3 and moving blades 5 forms a turbine stage, and the steam turbine equipment is configured by providing these turbine stages in a plurality of stages in the axial direction of the rotating shaft 4.

  The steam as the working fluid flows (from left to right in the drawing) in the annular flow path P formed by these paragraphs (the stationary blade 3 and the moving blade 5). Then, the steam whose direction of flow is changed by the stationary blades flows in the moving blades, whereby the kinetic energy is converted into the rotational energy of the moving blades and then circulates in the paragraphs. The rotating shaft 4 that has obtained the rotational energy rotates a generator (not shown) to generate power, and the steam that has lost energy is exhausted from the final stage.

  As described above, the working fluid is an energy source for generating electric power, and ideally, all the energy of the working fluid is converted into the rotational energy of the rotating shaft 4. However, a physical gap is required between the rotating rotating shaft 4 and the moving blade 5 and the stationary stationary blade 3 (and the turbine casing 9), but the working fluid flowing therethrough is the moving blade. Therefore, it does not contribute to the rotational energy of the rotating shaft 4. For this reason, it is essential to reduce the working fluid passing through the gap as much as possible. For the purpose of reducing the working fluid, the stationary blade inner ring 2 and the rotating shaft are disposed between the tip of the moving blade (the shroud 6) and the stationary blade outer ring 1. The labyrinth fins 7 and 8 are provided between each of them 4 and 4 to reduce the working fluid passing therethrough.

Next, the stationary blade of the axial turbine 10 will be described in detail with reference to FIGS.
As shown in FIG. 2, each stationary blade 3 is formed to be curved so that the trailing edge 3c protrudes toward the blade side surface in the circumferential direction. Here, the protrusion amount of the trailing edge 3c of each stationary blade 3 to the circumferential blade surface is δc _s , and the blade height of each stationary blade 3 (the distance between the stationary blade outer ring 1 and the stationary blade inner ring 2). H _s, the two circumferential blade root pitch vanes 3 that adjoin the t _s.

And .delta.c _s the amount of protrusion of the circumferential blade pressure side surface of the edge 3c after each vane 3, the blade height H _s of the stationary blade 3, the relationship between the circumferential blade root pitch t _s of the stationary blade 3 This will be described with reference to FIGS.
In the graph of FIG. 3, the horizontal axis represents the percentage .delta.c s _/ t _s of the protruding amount .delta.c _s in the circumferential direction blade pressure side surface of the edge 3c after the stationary blade 3 with respect to the circumferential direction blade base pitch t _s of the stationary blade 3, the vertical axis represents the ratio _H s / _{t s} of blade height _{H s} of the stationary blade 3 with respect to the circumferential direction blade base pitch _{t s} of the stationary blade 3. Further, in the graph of FIG. 4, the horizontal axis represents the above ratio δc _{s /} t s, the vertical axis of the embodiment with respect to operating efficiency of the axial flow turbine three-dimensional design is not performed in the axial turbine 10 The deviation Δη in operating efficiency is shown. Here, the axial turbine in which the three-dimensional design is not performed refers to an axial turbine in which the trailing edge 3c does not protrude in the circumferential direction in each stationary blade 3. In the graph of FIG. 4, lines A, B, the largest value of the ratio _H s / _{t s} in line A of and C, the value of the ratio _H s / _{t s} is the smallest in the line C.

In the graph of FIG. 4, the axial flow turbine of the present embodiment is increased by increasing the deviation Δη of the operation efficiency of the axial flow turbine 10 of the present embodiment relative to the axial flow turbine when the three-dimensional design is not performed. A good operating efficiency of 10 can be obtained. Here, when the maximum value of the deviation △ eta operational efficiency of the axial-flow turbine 10, the graph of FIG. 4, the value of the ratio δc _{s /} t s according to decrease the value of the ratio H _{s /} t _s Gradually grows. Thus in the case of the maximum value of the deviation △ eta operational efficiency of the axial-flow turbine 10, the showed the relationship between the ratio H _{s /} t _s and percentage δc _{s /} t s is a graph of FIG.

That is, when the maximum value of the deviation △ eta operational efficiency of the axial-flow turbine 10, as shown in FIG. 3, blade height H _s of the stationary blade 3 with respect to the circumferential direction blade base pitch t _s of the stationary blade 3 ratio _H s / ratio of protrusion amount .delta.c _s of as _{t s} is increased in the circumferential direction blade pressure side surface of the edge 3c after the stationary blade 3 with respect to the circumferential direction blade base pitch _{t s} of the stationary blade 3 .delta.c _s / _{t s} of Will decrease.
Furthermore, as shown in FIG. 2, the ratio _H s / _{t s} and the ratio δc _s / _{t s} is to satisfy the following relational expression (1).
δc _s / t _s = C _s × (H _s / t _s ) ^−βs (1)
In the formula (1), C _s is a constant of 0 or more and 0.5 or less, and β _s is a constant of 0.3 or more and 1.2 or less.

Next, the moving blades of the axial turbine 10 will be described in detail with reference to FIGS. 5 and 6.
As shown in FIG. 5, each rotor blade 5 is formed to be curved so that the trailing edge 5 c protrudes toward the blade side surface in the circumferential direction. Here, the protruding amount of the trailing edge 5c of each moving blade 5 to the blade side surface in the circumferential direction is δc _r , the blade height of each moving blade 5 is H _r , and the circumferential blade root pitch of two adjacent moving blades 5 _{Let tr} be.

And .delta.c _r a protrusion amount in the circumferential direction of the blade ventral side edges 5c after moving blades 5, the blades 5 wings and the height H _r, the circumferential blade root pitch t _r of the rotor blade 5 The relationship will be described with reference to FIG.
In the graph of FIG. 6, the horizontal axis represents the percentage δc _{r /} t r of the protrusion amount .delta.c _r in the circumferential direction blade pressure side surface of the edge 5c after moving blade 5 with respect to the circumferential direction blade base pitch t _r of the rotor blade 5, the vertical axis represents the ratio H _{r /} t _r of the blade height H _r of the rotor blade 5 with respect to the circumferential direction blade base pitch t _r of the rotor blade 5.

As in the case of the stationary blade, when the deviation Δη of the operation efficiency of the axial turbine 10 of the present embodiment relative to the axial turbine that is not three-dimensionally designed is a maximum value, the ratio H _r / t percentage value .delta.c _r / _{t r} according to decreasing the value of _r is gradually increased. It is shown the relationship between the ratio H _{r /} t _r and percentage δc _{r /} t r in the case where the maximum value of the deviation △ eta operational efficiency of the axial-flow turbine 10 is a graph of Figure 6.

As shown in FIG. 6, the circumferential blade root pitch t of the rotor blade 5 as the proportion H _{r /} t _r of the blade height H _r of the rotor blade 5 with respect to the circumferential direction blade base pitch t _r of the rotor blade 5 is increased ratio δc _{r /} t r of the protrusion amount .delta.c _r in the circumferential direction blade pressure side side edge 5c after moving blade 5 comes to decrease for _r.
Furthermore, as shown in FIG. 6, the ratio _H r / _{t r} and the ratio δc _r / _{t r} will satisfy the following relational expression (2).
_{δc r / t r = C r} × (H r / t r) -βr ··· formula (2)
In the equation (2), the _{C r} 0 or more and 0.5 or less constant, the beta _r is a constant of 0.3 to 1.2.

Next, the operation of the present embodiment having such a configuration will be described.
In the axial turbine 10 shown in FIG. 1, when the working fluid flows in the right direction in the annular flow path P, the working fluid passes through the stationary blade 3 and the moving blade 5 to rotate the rotating shaft 4 and the rotor disk 4a. As a result, the energy of the working fluid is recovered and power is obtained.

At this time, each vane 3, the ratio H _{s /} t _s of blade height H _s in the circumferential direction blade base pitch t _s is reduced, the trailing edge of the stationary blade 3 with respect to the circumferential direction blade base pitch t _s ratio δc _{s /} t s of the protruding amount .delta.c _s in the circumferential direction the blade pressure side is increased 3c, axial flow turbine of the present embodiment with respect to axial flow turbine three-dimensional design is not performed as shown in FIG. 4 The deviation Δη in operating efficiency of 10 is also increased.
This will be described in detail with reference to FIG. FIG. 7 is a graph showing the relationship between the specific aspect ratio and specific blade row loss in the stationary blades of the axial flow turbine 10. Here, the specific aspect ratio refers to the aspect ratio of the stationary blade 3 for each paragraph (blade height H _s / blade length code l _s , see FIG. 10A) as the aspect ratio of the stationary blade on the inlet side. It means the value divided.

The total loss of the cascade of the stationary blades of the axial turbine 10 is determined by the sum of the airfoil loss and the secondary loss. As shown in FIG. 7, the ratio in the case blade height H _s of the stationary blade 3 is small in the entire blade cascade loss of the aspect ratio is small becomes secondary loss ratio is increased. Therefore, by increasing the radial blade loading distribution at the center (i.e., by increasing the protrusion amount .delta.c _s in the circumferential direction blade pressure side surface of the edge 3c after stationary blade 3), the wing inner and outer endwall Loss can be reduced and secondary loss can be reduced to improve the overall performance of the cascade.
On the other hand, proportions of blade height H _s of the stationary blade 3 is large in the entire blade cascade loss of the aspect ratio become large secondary loss ratio becomes smaller. Thus, is possible to reduce the circumferential blade pressure overhang .delta.c _s the increasing substantial secondary losses by reducing the loss at blade inner and outer end wall portion of the side surface of the edge 3c after stationary blade 3 Conversely, the increased load at the blade center increases the blade loss and reduces the overall performance of the blade row.
Thus, the by blade height H _s of the stationary blade 3, a change in protrusion amount .delta.c _s in the circumferential direction blade pressure side surface of the edge 3c after vanes 3 that the overall performance of the blade row with each paragraph is optimum It becomes. Therefore, as shown in the graph of FIG. 3, as the proportion H _{s /} t _s of blade height H _s in the circumferential direction blade base pitch t _s is increased, after the stationary blade 3 with respect to the circumferential direction blade base pitch t _s by percentage δc _{s /} t s of the protruding amount .delta.c _s in the circumferential direction blade pressure side surface of the edge 3c to set the shape of each vane 3 to reduce, the optimal load distribution in the vane 3 for each paragraph Therefore, the overall performance of the axial flow turbine 10 can be improved.

Similarly, in each blade 5, when the ratio H _{r /} t _r of the blade height H _r with respect to the circumferential direction blade base pitch t _r decreases, the edge 5c after the rotor blade 5 with respect to the circumferential direction blade base pitch t _r circumferential wings proportion δc _{r /} t r of the protrusion amount .delta.c _r to the ventral aspect is increased, the three-dimensional design of the operating efficiency of the axial flow turbine 10 of this embodiment with respect to axial flow turbine not done deviation △ η is also increased.
This will be described in detail as in the case of the stationary blade group.

The total loss of the blade cascade of the axial turbine 10 is determined by the sum of the airfoil loss and the secondary loss. As with the stationary blade, percentage of total blade cascade loss of the secondary loss becomes smaller ratio aspect ratio when blade height H _r of the rotor blade 5 is small becomes large. Therefore, by increasing the radial blade loading distribution at the center (i.e., by increasing the protrusion amount .delta.c _r in the circumferential direction blade pressure side surface of the edge 5c after moving blade 5), the wing inner and outer endwall Loss can be reduced and secondary loss can be reduced to improve the overall performance of the cascade.
Meanwhile, the ratio of total blade cascade loss of the secondary loss becomes larger the ratio the aspect ratio when blade height H _r of the rotor blade 5 is greater is reduced. Thus, is possible to reduce the circumferential blade pressure overhang .delta.c _r a is increased two major primary loss to reduce the loss at the blade inner and outer end wall portion of the side surface of the edge 5c after moving blade 5 Conversely, the increased load at the blade center increases the blade loss and reduces the overall performance of the blade row.
Thus, the blade height H _r of the rotor blade 5, a change in protrusion amount .delta.c _r in the circumferential direction blade pressure side surface of the edge 5c after moving blades 5 to be optimal overall performance of the blade row with each paragraph It will be. Therefore, as shown in the graph of FIG. 6, as the proportion H _{r /} t _r of the blade height H _r with respect to the circumferential direction blade base pitch t _r increases, after the rotor blade 5 with respect to the circumferential direction blade base pitch t _r by percentage δc _{r /} t r of the protrusion amount .delta.c _r in the circumferential direction blade pressure side surface of the edge 5c sets the shape of each blade 5 so as to reduce, the optimal load distribution in the rotor blades 5 of each paragraph Therefore, the overall performance of the axial flow turbine 10 can be improved.

According to the axial-flow turbine 10 of the present embodiment as described above, the as the proportion H _{s /} t _s of blade height H _s of the stationary blade 3 with respect to the circumferential direction blade base pitch t _s of the stationary blade 3 is increased setting the shape of each vane 3 as the ratio .delta.c s _/ t _s of the protruding amount .delta.c _s in the circumferential direction blade pressure side surface of the edge 3c after the stationary blade 3 with respect to the circumferential direction blade base pitch t _s of the stationary blade 3 is decreased As a result, an optimum load distribution can be formed in the stationary blade 3 for each paragraph, and the overall performance of the axial turbine 10 can be improved.

In addition, since the shape of each stationary blade 3 is set so as to satisfy the following formula (1), it is possible to form an optimum load distribution with higher accuracy for each paragraph, and the entire axial turbine 10 can be formed. The performance can be further improved.
δc _s / t _s = C _s × (H _s / t _s ) ^−βs (1)
(In Formula (1), C _s is a constant of 0 or more and 0.5 or less, and β _s is a constant of 0.3 or more and 1.2 or less.)

Furthermore, the rotor blade as according to the axial turbine 10 of this embodiment the proportion H _{r /} t _r the blade height H _r of the rotor blade 5 with respect to the circumferential direction blade base pitch t _r of the rotor blade 5 is increased for 5 circumferential blade root pitch t _r is set the shape of each blade 5 so that the ratio δc _{r /} t r of the protrusion amount .delta.c _r in the circumferential direction blade pressure side surface of the edge 5c after moving blade 5 is reduced As a result, an optimum load distribution can be formed in the rotor blade 5 for each paragraph, and the overall performance of the axial turbine 10 can be improved.

In addition, since the shape of each rotor blade 5 is set so as to satisfy the following formula (2), it is possible to form an optimum load distribution with higher accuracy for each paragraph, and the entire axial turbine 10 can be formed. The performance can be further improved.
_{δc r / t r = C r} × (H r / t r) -βr ··· formula (2)
(In the expression (2), the _{C r} 0 or more and 0.5 or less constant, beta _r is a constant of 0.3 to 1.2.)

  Moreover, since the axial flow turbine 10 of this Embodiment is equipped with the stationary blade and moving blade which were set as mentioned above, both optimal load distribution in the stationary blade 3 and the moving blade 5 for every paragraph. Can be formed, and the overall performance of the axial turbine 10 can be improved more reliably.

It is sectional drawing which shows the structure of an axial flow turbine. It is an AA arrow line view of Drawing 1 showing the composition of the stationary blade of the axial flow turbine of this embodiment. Is a graph showing the relationship between the ratio δc _s / _{t s} and the ratio _H s / _{t s} in the stationary blade of an axial flow turbine of FIG. Is a graph showing the relationship between the ratio δc _{s /} t s and axial deviation of operating efficiency of the turbine △ eta in the stationary blade of an axial flow turbine of FIG. It is a BB arrow line view of Drawing 1 showing the composition of the moving blade of the axial flow turbine of this embodiment. It is a graph showing the relationship between the ratio δc _r / _{t r} and the ratio _H r / _{t r} at the rotor blade axial flow turbine of FIG. It is a graph which shows the relationship between the specific aspect ratio and specific blade row loss in the stationary blade of an axial flow turbine. It is an AA arrow line view of Drawing 1 showing the composition of the stationary blade of the conventional axial flow turbine. (A) is a cross-sectional view showing the airfoil geometry of the stationary blade, and (b) is a graph showing the relationship between the airfoil geometry of the stationary blade and the airfoil loss obtained by the cascade wind tunnel test. It is. (A) is a perspective view which shows the structure of a stationary blade, (b) is a graph showing the relationship between an aspect-ratio and the magnitude | size of a secondary loss. It is a graph which shows the relationship between the blade height ratio and each loss factor in a stationary blade row. It is a graph which shows the change of the fluid outflow angle (alpha) when the ratio (delta) c / t is made constant and the blade height H is changed. It is a graph which shows the change of the fluid outflow angle (alpha) when the ratio (delta) c / t is made constant and a blade size and a blade pitch are changed.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 Stator blade outer ring 2 Stator blade inner ring 3 Stator blade 3c Trailing edge 4 Rotating shaft 4a Rotor disk 5 Rotor blade 5c Trailing edge 6 Shroud 7 Labyrinth fin 8 Labyrinth fin 9 Turbine casing 10 Axial flow turbine 50 Axial flow turbine 53 Stator blade 53c Rear Edge 55 blade 55c trailing edge P annular channel Q gap

Claims (5)

In a stationary blade of an axial flow turbine provided in a plurality of stages in an annular flow path,
Each stationary blade is curved and formed so that the trailing edge protrudes to the ventral side of the circumferential blade,
When the circumferential blade root pitch of each stationary blade is t _s , the amount of protrusion of the trailing edge of each stationary blade to the circumferential blade surface is δc _s , and the blade height of each stationary blade is H _s , circumferential direction of the trailing edge of the vanes relative to the circumferential direction blade base pitch t _s of the vane as the proportion H _{s /} t _s of blade height H _s of the vanes relative to the circumferential direction blade base pitch t _s of the stationary blade is increased overhang .delta.c _s ratio δc _{s /} t s is stator blade of the axial flow turbine, characterized in that it is configured as a shape of each vane to reduce to Tsubasahara side. The stator blade of an axial-flow turbine according to claim 1, wherein each stator blade has a shape that satisfies the following formula.
_{δc s / t s = C s} × (H s / t s) -βs
(However, C _s is a constant between 0 and 0.5, and β _s is a constant between 0.3 and 1.2.) In an axial flow turbine blade provided in multiple stages in an annular flow path,
Each blade is curved and formed so that the trailing edge protrudes to the ventral side of the circumferential blade,
When each blade circumferential blade root pitch t _r, the .delta.c r a projection amount of the edge of the circumferential blade pressure side after each rotor _blade, wherein the blade height of the blades was set to H _r, circumferential direction after the rotor blades of the rotor blade with respect to the circumferential direction blade base pitch t _r edge as the proportion H _{r /} t _r of the blade of the blade height H _r for the rotor blades in the circumferential direction the blade root pitch t _r increases rotor blades of an axial flow turbine, wherein the shape of each blade is configured to proportion δc _{r /} t r of the protrusion amount .delta.c _r to Tsubasahara side is reduced. 4. The rotor blade of an axial flow turbine according to claim 3, wherein the shape of each rotor blade is set so as to satisfy the following formula.
_{δc r / t r = C r} × (H r / t r) -βr
(However, _Cr is a constant between 0 and 0.5, and _βr is a constant between 0.3 and 1.2.)   An axial flow turbine comprising the stationary blade of the axial flow turbine according to claim 1 and the moving blade of the axial flow turbine according to claim 3.

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