U.S. patent application number 13/396798 was filed with the patent office on 2012-08-23 for turbine nozzle blade and steam turbine equipment using same.
This patent application is currently assigned to Hitachi, Ltd.. Invention is credited to Seiichi Kimura, Tadaharu Kishibe, Goingwon Lee, Kiyoshi Segawa, Takanori Shibata.
Application Number | 20120210715 13/396798 |
Document ID | / |
Family ID | 45656180 |
Filed Date | 2012-08-23 |
United States Patent
Application |
20120210715 |
Kind Code |
A1 |
Shibata; Takanori ; et
al. |
August 23, 2012 |
Turbine Nozzle Blade and Steam Turbine Equipment Using Same
Abstract
Disclosed is a highly efficient turbine nozzle blade that
reduces the number of blades in an axial-flow turbine while
reducing secondary-flow loss. In the nozzle blade, when a
differential pressure between a pressure side and a suction side of
each blade, at the same axial chord position of the blade, is
defined as a load of the blade, and a ratio between axial chord
length "Cx" of the blade and an axial distance "xp" from a leading
edge of the blade at a maximum load position that maximizes the
blade load is defined as a maximum load relative position, Cx is
greater at a hub and tip than at an intermediate vertical portion,
and simultaneously a maximum load relative position at the hub and
tip is set to be nearer to a trailing edge thereof than a maximum
load relative position of the intermediate vertical portion of the
blade.
Inventors: |
Shibata; Takanori;
(Hitachinaka, JP) ; Segawa; Kiyoshi; (Hitachi,
JP) ; Kishibe; Tadaharu; (Hitachinaka, JP) ;
Kimura; Seiichi; (Hitachi, JP) ; Lee; Goingwon;
(Hitachi, JP) |
Assignee: |
Hitachi, Ltd.
Tokyo
JP
|
Family ID: |
45656180 |
Appl. No.: |
13/396798 |
Filed: |
February 15, 2012 |
Current U.S.
Class: |
60/670 ;
416/223A |
Current CPC
Class: |
Y02T 50/60 20130101;
F01D 5/141 20130101; F05D 2240/122 20130101; F01D 9/041 20130101;
F05D 2240/12 20130101; Y02T 50/673 20130101; F05D 2220/31
20130101 |
Class at
Publication: |
60/670 ;
416/223.A |
International
Class: |
F01D 5/14 20060101
F01D005/14; F01K 13/00 20060101 F01K013/00 |
Foreign Application Data
Date |
Code |
Application Number |
Feb 22, 2011 |
JP |
2011-035305 |
Claims
1. A nozzle blade for an axial-flow turbine, wherein, when a
differential pressure between a pressure side and a suction side of
each of blades, at one same axial chord position of the blade, is
defined as a load of the blade, and a ratio between axial chord
length of the blade and an axial distance from a leading edge
thereof at one same vertical position of the blade where the blade
load becomes a maximum is defined as a maximum load relative
position, the axial chord length of the blade is greater at a hub
and tip thereof than at an intermediate vertical portion thereof,
and a maximum load relative position at the hub and tip of the
blade is set to be nearer to a trailing edge thereof than a maximum
load relative position of the intermediate vertical portion of the
blade.
2. A nozzle blade for an axial-flow turbine, wherein: when a point,
except at a leading edge and trailing edge of each of blades, at
which a static pressure on a suction surface of the blade becomes a
minimum, is defined as a suction-side minimum pressure position,
axial chord length of the blade is greater at a hub and tip thereof
than at an intermediate vertical portion thereof, and a
suction-side minimum pressure position at the hub and tip of the
blade is set to be nearer to a trailing-edge side than a
suction-side minimum pressure position of the intermediate vertical
portion of the blade.
3. A nozzle blade for an axial-flow turbine, wherein: each of
blades is formed into a concave shape on a suction surface of a
trailing edge of the blade, the blade being further formed into a
convex shape on a pressure surface of the blade trailing edge; and
axial chord length of the blade is greater at a hub and tip of the
blade than at an intermediate vertical portion of the blade; and
circumferential chord length is smaller at the hub and tip of the
blade than at the intermediate vertical portion of the blade.
4. A nozzle blade for an axial-flow turbine, wherein: axial chord
length of each of blades is greater at a hub and tip thereof than
at an intermediate vertical portion of the blade; circumferential
chord length is smaller at the hub and tip of the blade than at the
intermediate vertical portion of the blade; and a stagger angle of
the blade is smaller at the hub and tip of the blade than at the
intermediate vertical portion thereof.
5. The turbine nozzle blade according to claim 1, wherein: when an
interblade pitch is expressed as "t," the axial chord length as
"Cx," an inflow angle when measured from an axial direction, as
".alpha.," and an outflow angle when measured from the axial
direction, as ".beta.," and a value .PSI. given by
.PSI.=2(t/Cx)cos.sup.2 .beta.|tan .alpha.-tan .beta.| (Expression
1) is defined as a load coefficient, the load coefficient at the
intermediate vertical position of the blade takes any value ranging
between 0.7 and 1.1 in Expression 1.
6. The turbine nozzle blade according to claim 5, wherein: as the
axial chord length heads from the blade hub towards the blade tip,
the axial chord length of the blade hub decreases at a rate
equivalent to or higher than a change rate of the axial chord
length in the vicinity of the intermediate vertical position of the
blade, and the axial chord length of the blade tip increases at a
rate equivalent to or higher than the change rate of the axial
chord length in the vicinity of the intermediate vertical position
of the blade.
7. Steam turbine equipment comprising: a steam generator for
heating water and generating steam; a steam turbine driven by the
steam generated by the steam generator; and a condenser for
condensing the steam that has driven the steam turbine; wherein the
steam turbine includes the turbine nozzle blade according to claim
1.
8. The turbine nozzle blade according to 2, wherein: when an
interblade pitch is expressed as "t," the axial chord length as
"Cx," an inflow angle when measured from an axial direction, as
".alpha.," and an outflow angle when measured from the axial
direction, as ".beta.," and a value .PSI. given by
.PSI.=2(t/Cx)cos.sup.2 .beta.|tan .alpha.-tan .beta.| (Expression
1) is defined as a load coefficient, the load coefficient at the
intermediate vertical position of the blade takes any value ranging
between 0.7 and 1.1 in Expression 1.
9. The turbine nozzle blade according to 3, wherein: when an
interblade pitch is expressed as "t," the axial chord length as
"Cx," an inflow angle when measured from an axial direction, as
".alpha.," and an outflow angle when measured from the axial
direction, as ".beta.," and a value .PSI. given by
.PSI.=2(t/Cx)cos.sup.2 .beta.|tan .alpha.-tan .beta.| (Expression
1) is defined as a load coefficient, the load coefficient at the
intermediate vertical position of the blade takes any value ranging
between 0.7 and 1.1 in Expression 1.
10. The turbine nozzle blade according to 4, wherein: when an
interblade pitch is expressed as "t," the axial chord length as
"Cx," an inflow angle when measured from an axial direction, as
".alpha.," and an outflow angle when measured from the axial
direction, as ".beta.," and a value .PSI. given by
.PSI.=2(t/Cx)cos.sup.2 .beta.|tan .alpha.-tan .beta.| (Expression
1) is defined as a load coefficient, the load coefficient at the
intermediate vertical position of the blade takes any value ranging
between 0.7 and 1.1 in Expression 1.
11. Steam turbine equipment comprising: a steam generator for
heating water and generating steam; a steam turbine driven by the
steam generated by the steam generator; and a condenser for
condensing the steam that has driven the steam turbine; wherein the
steam turbine includes the turbine nozzle blade according to claim
2.
12. Steam turbine equipment comprising: a steam generator for
heating water and generating steam; a steam turbine driven by the
steam generated by the steam generator; and a condenser for
condensing the steam that has driven the steam turbine; wherein the
steam turbine includes the turbine nozzle blade according to claim
3.
13. Steam turbine equipment comprising: a steam generator for
heating water and generating steam; a steam turbine driven by the
steam generated by the steam generator; and a condenser for
condensing the steam that has driven the steam turbine; wherein the
steam turbine includes the turbine nozzle blade according to claim
4.
14. Steam turbine equipment comprising: a steam generator for
heating water and generating steam; a steam turbine driven by the
steam generated by the steam generator; and a condenser for
condensing the steam that has driven the steam turbine; wherein the
steam turbine includes the turbine nozzle blade according to claim
5.
15. Steam turbine equipment comprising: a steam generator for
heating water and generating steam; a steam turbine driven by the
steam generated by the steam generator; and a condenser for
condensing the steam that has driven the steam turbine; wherein the
steam turbine includes the turbine nozzle blade according to claim
6.
Description
BACKGROUND OF THE INVENTION
[0001] 1. Field of the Invention
[0002] The present invention relates to a turbine nozzle blade used
for axial-flow turbines, especially gas turbines, steam turbines,
and the like, at electric power plants.
[0003] 2. Description of the Related Art
[0004] In recent years, it is strongly requested to further enhance
turbine performance for the purpose of an improvement in
electric-power generating efficiency at power plants. Turbine
performance is involved in the stage loss, exhaust loss, and
mechanical loss of the turbine, and it is considered to be most
effective to reduce the stage loss, in particular, for turbine
performance improvement. There are various types of stage loss.
Such stage loss can be roughly categorized, namely, (a) profile
loss due to the blade shape itself, (b) secondary-flow loss due to
a working fluid flow not along a main flow of the working fluid,
and (c) leakage loss caused by leakage of the working fluid from
the main flow. The secondary-flow loss, more specifically caused by
a difference in vertical blade height and/or interference between
the blade and an endwall boundary layer, has had a nature that
simply optimizing a vertical sectional shape of a certain blade is
ineffective for reducing the loss.
[0005] In order to solve such a problem, JP-2008-202420-A proposes
a technique for reducing circumferential chord length of a central
portion of a blade relative to that of a tip and hub near an
endwall of the blade, while at the same time gradually reducing the
latter circumferential chord length as the chord draws nearer to
the central portion of the blade. According to JP-2008-202420-A,
forming the blade in that form slopes the blade surface with a
convex pressure side at an endwall-neighboring portion close to a
leading edge of the blade, thus enabling reduction in magnitude of
a secondary flow occurring near the endwall. JP-2008-202420-A also
describes that near a throat of the blade, the blade surface is
slightly sloped, which enables suppression of a flow-rate bias to
one side in a vertical direction of the blade, alleviating an
increase in rotor blade loss.
SUMMARY OF THE INVENTION
[0006] JP-2008-202420-A, however, does not describe on a stagger
angle of the blades and on a load distribution in an axial chord
direction, and a reduction effect against secondary-flow loss has
been unlikely to be fully brought about with the conventional
technique involved. The load distribution, in particular, has a
close relationship to the number of blades, so it is crucial to
optimize the load distribution according to the number of
blades.
[0007] The present invention is intended to provide a turbine
nozzle blade minimized in contact area between a blade and a flow
of fluid by reducing the number of blades to a minimum requirement
for a desired direction change of a fluid flow in order to reduce
blade profile loss. And at the same time, the turbine nozzle blade
is further improved in stage efficiency of the turbine by
optimizing a load distribution and axial chord length near an
endwall in order to reduce secondary-flow loss due to interference
between the blade and a boundary layer on the endwall. The
reduction in the number of blades also contributes to reduction in
manufacturing costs of the turbine and to supply of a turbine high
in efficiency and low in manufacturing costs.
[0008] In order to achieve the above object, in a turbine nozzle
blade of the present invention, when a differential pressure
between a pressure side and a suction side of a blade, at one same
axial chord position of the blade, is defined as a load of the
blade, and a ratio between axial chord length of the blade and an
axial distance from a leading edge thereof at one same vertical
position of the blade where the blade load becomes a maximum is
defined as a maximum load relative position, the axial chord length
of the blade is greater at a hub and tip thereof than at an
intermediate vertical portion thereof, and simultaneously a maximum
load relative position at the hub and tip of the blade is set to be
nearer to a trailing edge thereof than a maximum load relative
position of the intermediate vertical portion of the blade.
[0009] The above configuration enables supply of a turbine nozzle
blade in which reducing the number of blades is effective for
reducing secondary-flow loss in addition to profile loss, and hence
for improving turbine stage efficiency.
[0010] The reduction in the number of blades further contributes to
reduction in manufacturing costs of the turbine and to supply of a
turbine high in efficiency and low in manufacturing costs.
BRIEF DESCRIPTION OF THE DRAWINGS
[0011] FIG. 1 is a cross-sectional view showing a structure of
essential elements of a steam turbine stage section according to a
first embodiment of the present invention;
[0012] FIG. 2 is a perspective view of the turbine nozzle blade
structure of the first embodiment;
[0013] FIG. 3 is a cross-sectional view showing a structure of
essential elements of a conventional steam turbine stage
section;
[0014] FIG. 4 is an explanatory diagram that schematically
represents vortex flows occurring between the blades of the cascade
structure shown in FIG. 3;
[0015] FIGS. 5A, 5B, and 5C are graphs that represent vertical
axial chord length, maximum load relative position, and loss
distribution of a blade in the turbine nozzle blade according to
the first embodiment;
[0016] FIG. 6 is a graph that represents a relationship between a
pressure distribution on the blade surface and a load
distribution;
[0017] FIG. 7 is an explanatory diagram that represents a
meridional shape and cross-sectional shape of the turbine nozzle
blade according to the first embodiment;
[0018] FIG. 8 is an explanatory diagram that represents a
meridional shape of a turbine nozzle blade according to a second
embodiment of the present invention;
[0019] FIG. 9 is an explanatory diagram that represents a
meridional shape of a turbine nozzle blade according to a third
embodiment of the present invention; and
[0020] FIG. 10 is a schematic system diagram of steam turbine
equipment according to the first embodiment.
DESCRIPTION OF THE PREFERRED EMBODIMENTS
[0021] Hereunder, embodiments of the present invention will be
described in detail referring to the accompanying drawings as
appropriate. The same reference number is assigned to each of
equivalent constituent elements throughout the drawings.
[0022] Each embodiment described below is an example in which the
present invention is applied to nozzle blades of a steam turbine.
While the invention is applied to a steam turbine for the sake of
convenience in the description, principles of operation are
substantially the same, even for gas turbines using a working fluid
different from that of the turbine according to the invention, and
the invention can be applied to practically all types of axial-flow
turbines.
First Embodiment
[0023] A first embodiment of the present invention is described
below. FIG. 1 is a cross-sectional view showing a structure of a
steam turbine stage section according to the present embodiment. As
shown in FIG. 1, the steam turbine stage section according to the
present embodiment includes a plurality of nozzle blades 3 each
disposed in a circumferential direction of the turbine, between a
diaphragm outer ring 1 and a diaphragm inner ring 2. The steam
turbine stage section also includes a plurality of rotor blades 5
each disposed in a circumferential direction of a turbine rotor 4,
at a downstream side of a flow direction of steam relative to one
of the nozzle blades 3. The downstream side of the flow direction
of steam relative to the nozzle blade 3 is hereinafter referred to
simply as the downstream side. The radially outer side of the
turbine is hereinafter referred to simply as the outer side. A
shroud 6 is provided at a tip of each rotor blade 5, at a radially
outer side of the turbine, and a sealing structure not shown is
provided between the shroud 6 and a stationary body opposed
thereto.
[0024] A main steam flow 7 that is the working fluid is supplied
from the upstream side of a nozzle leading edge 8, passing across
the blade, and then flows out into the downstream side of a nozzle
trailing edge 9. The main steam flow 7 that has flown out from the
nozzle blade 3 impinges upon the rotor blade 5 located at a
downstream position relative to the nozzle blade 3. The steam
turbine thus rotates the turbine rotor 4, converts rotational
energy into electrical energy by means of a power generator (not
shown) connected to an end portion of the turbine rotor 4, and
generates electricity.
[0025] For ease in understanding a three-dimensional shape of the
nozzle blade according to the present invention, FIG. 2 shows the
nozzle blade in perspective view. In FIG. 2, part of the cascade
structure including the plurality of nozzle blades in the
circumferential direction of the turbine is represented in
horizontally developed form for the sake of descriptive
convenience. The nozzle blades 3 are arranged in the
circumferential direction between the diaphragm outer ring 1 and
the diaphragm inner ring 2. Each nozzle blade 3 is fixed at its hub
side 10 to the diaphragm inner ring 2, and at its tip side 11 to
the diaphragm outer ring 1. The nozzle blades 3 are equally spaced
at horizontals interval (pitch) "t" between adjacent blades in the
circumferential direction of the turbine, the pitch "t" being
determined from the number of blades in the cascade. The shortest
distance "s" from a nozzle blade 3a to the trailing edge of an
adjacent nozzle blade 3b, in a vertical section of each blade, is
hereinafter referred to as throat length. The present embodiment is
characterized in that as in FIG. 1, since the nozzle blades are
arranged in sectionally stacked form in a radial direction so that
the leading edges of each blade are linear, the blade has a suction
surface of a concave shape and a pressure surface of a convex
shape, near the trailing edge.
[0026] A configuration and operation of the present embodiment are
described in detail below. FIG. 1 shows a shape of one nozzle blade
3 as projected on a meridional plane. For the sake of descriptive
convenience, an axial direction is expressed as "x," a span
direction (synonymous with a vertical direction) of the blade, as
"y," and height from the hub of the blade to the tip, as blade
height H, with a leading edge 8 of the blade hub taken as an
origin. In addition, an axial distance from the leading edge of a
blade section to the trailing edge thereof, at certain blade height
"y," is termed axial chord length and expressed as reference code
"Cx," whereas an axial distance from the leading edge to a maximum
load position (described later herein) is expressed as reference
code "xp." The load here refers to the blade load that is a
differential pressure between the pressure side and suction side of
the blade, at the same axial chord position. As can be seen from
FIG. 1, the nozzle blades of the present invention are
characterized in that the axial chord length "Cx" has a definite
radial region near an intermediate vertical position of the blade
and in that the "Cx" value is smaller than those of the hub, tip,
and other portions neighboring an endwall of the blade. The maximum
load position "xp" near the intermediate vertical position is
relatively close to the leading edge, compared with the maximum
load position near the endwall.
[0027] A meridional shape of a general turbine stage is shown in
FIG. 3 for comparison with that of the present invention. As can be
seen from FIG. 3, the axial chord lengths of conventional nozzle
blades have been constant over a substantially entire region from
the blade hub to the blade tip. Additionally, in each conventional
nozzle blade, the maximum load position, as it approaches the hub,
is nearer to the trailing edge, and as it approaches the tip, is
nearer to the leading edge. The maximum load position has
association with a load coefficient (described later herein) of the
blade. That is to say, the maximum load position means that at the
hub side of a small blade pitch "t," the maximum load position is
set to be closer to the trailing edge, because of a small load
coefficient, while at the tip side of a large blade tip "t," the
maximum load position is set to be more shifted in an upstream
direction, because of a large load coefficient.
[0028] FIG. 4 is a schematic diagram of a secondary flow occurring
in the turbine nozzle blade shown in FIG. 3. For the sake of
simplification, a diaphragm outer ring 1 and inner ring 2 are
omitted in the diagram. As shown in FIG. 4, each nozzle blade 3
includes a pressure surface 12 formed at a pressure side of the
blade, and a suction surface 13 formed at a suction side of the
blade. A main steam flow 7 is supplied from the upstream side of a
nozzle leading edge 8, passing through a blade passage 14 formed
between the pressure surface 12 and the suction surface 13, and
then flows into the downstream side of a nozzle trailing edge 9.
While the main steam flow 7 passes through the blade passage 14, a
pressure gradient occurs between the blades. This generates a
secondary flow 15 heading from the pressure surface 12 towards the
suction surface 13. In addition, the main steam flow 7 streams from
the leading edge 8 into the blade passage 14, generating vortices
on the pressure surface 12 and the suction surface 13. Among all
generated vortices, only those existing on the pressure surface 12
each form a passage vortex 16 while developing within the blade
passage. After that, the vortices on the pressure surface 12 are
influenced by the secondary flow 15 and move to the suction surface
13. The passage vortex 16 occurs at both the blade hub 10 and the
blade tip 11. In a cascade structure with nozzle blades of a small
aspect ratio, especially nozzle blades of an aspect ratio as low as
1.0 or less, the passage vortices 16 that have occurred at the
blade hub 10 and the blade tip 11 interfere with each other,
forming a flow significant in loss. The passage vortices 16 lowers
work efficiency of the turbine nozzle blades, becoming a major
cause of endwall loss.
[0029] Differences in "Cx" and "xp" between the present invention
and the conventional example are described below referring to FIGS.
5A, 5B, and 5C. Plottings of the axial chord length "Cx," maximum
load relative position "xp/Cx," and fluid loss, in that order from
above, against dimensionless blade height "y/H" on the horizontal
axis, are shown in FIGS. 5A, 5B, and 5C, respectively. The
dimensionless blade height "y/H" here is a value derived by making
it dimensionless with the blade height H so that the position
relative to the blade height H, the dimension from the hub to tip
of a blade of given blade height, becomes readily identifiable. The
maximum load relative position is a ratio between the axial chord
length of the blade and the axial distance from the leading edge at
the same vertical position of the blade where the blade load
becomes a maximum. The maximum load relative position is a value
derived by making it dimensionless with the axial chord length "Cx"
at a particular vertical position so that the position relative to
the axial chord length becomes readily identifiable. Solid lines in
FIGS. 5A-5C represent each distribution state in the present
embodiment, and dashed lines represent each distribution state in
the conventional example.
[0030] As can be seen from FIG. 5A, the axial chord lengths of
conventional nozzle blades have commonly exhibited a distribution
that is constant, or otherwise monotonically increases, in the
region from the blade hub to the blade tip, whereas the axial chord
length in the present embodiment is greater than the axial chord
length at the intermediate vertical position of the blade, in the
neighborhood of the blade hub or tip. As the axial chord heads from
the blade hub towards the blade tip, the axial chord length of the
blade hub decreases at a rate equivalent to or higher than a change
rate of the axial chord length in the vicinity of the intermediate
vertical position of the blade, and the axial chord length of the
blade tip increases at a rate equivalent to or higher than the
change rate of the axial chord length in the vicinity of the
intermediate vertical position of the blade. While the example in
which the axial chord lengths of the hub and tip are greater than
the axial chord length of the intermediate vertical position has
been described and shown as a typical example in the present
embodiment, only the axial chord length of either the hub or the
tip may be greater than the axial chord length of the intermediate
vertical position. In addition, in order to render the leading
edges of the nozzle blades linear in FIG. 1, the nozzle blades are
arranged in sectionally stacked form with the trailing edges
extended, however the nozzle blades do not always need to be
stacked in this fashion. For example, the nozzle blades may be
stacked so that the trailing edges are linear, as in FIG. 9.
Furthermore, while FIG. 1 shows a region in which the axial chord
length "Cx" at the intermediate vertical position of the blade is
constant in the radial direction, the region of constant "Cx" may
be unnecessary if the blade is too short. What is important in the
present invention is whether the blade hub, intermediate height,
and the axial chord length at the tip are maintained in the
relationship described above.
[0031] If such a constant "Cx" distribution as seen in the
conventional example is adopted as a distribution of axial chord
length, "t/Cx" that is the ratio between the interblade pitch "t"
and the axial chord length "Cx" will be greater at positions closer
to the tip of the blade. This characteristic is involved in the
fact that if the number of blades is considered to be fixed, blade
sections nearer to an outside-diametral side will be larger in
pitch "t." The "t/Cx" value is involved in an aerodynamic blade
load, and as "t/Cx" increases, the blade load for obtaining a turn
of a certain flow will be higher.
[0032] The following Zweifel load coefficient is generally used as
an index to denote the aerodynamic load exerted upon the blade:
.PSI.=2(t/Cx)cos.sup.2 .beta.|tan .alpha.-tan .beta.| (Expression
1)
where "t" is the pitch, "Cx" the axial chord length, ".alpha." an
inflow angle as measured from the axial direction, and ".beta." an
outflow angle as measured from the axial direction. The above
Expression indicates that the load becomes higher with increasing
turn of flow at the blade inlet or exit or with increasing cascade
pitch relative to the axial chord length. In general, cascades of
high load coefficients tend to increase in profile loss and are
known to depend greatly upon the load distribution of the blade, as
well as upon the load coefficient.
[0033] The relationship between profile loss and the load
distribution is described below referring to FIG. 6. The load
distribution here refers to a distribution of the blade load from
the leading edge of the blade to the trailing edge thereof, the
blade load being defined as a difference in blade surface static
pressure between upper and lower surfaces of the blade. As with
non-design points having an incidence angle, a region in which the
pressures at the leading edge and trailing edge of the blade suffer
abrupt changes is excluded from the load distribution in FIG. 6.
Profile loss is composed of factors such as skin friction loss,
deceleration loss, and trailing edge loss. Skin friction loss
increases with increasing fluid velocity level of the blade, while
deceleration loss increases with increasing rate of reduction in
the velocity along the blade surface. The fluid velocity
distribution of the blade has a strong correlation to the pressure
distribution, with a decrease in blade surface pressure reducing a
density of the fluid and increasing a velocity of the fluid.
Accordingly, the blade surface also suffers more deceleration loss
with an increase in adverse pressure gradient.
[0034] An ideal distribution of blade surface pressure that allows
for the above is described below. First, conditions under which the
load coefficient is fixed are assumed. For reduced skin friction
loss, aft loading with suppressed fore loading of the blade is
desirable in order to suppress an increase in velocity. If a load
peak is brought too close to the trailing edge side, however, the
adverse pressure gradient will commonly be steep near a trailing
edge portion of the suction side, resulting in augmented
deceleration loss. Therefore, loading that generates a load
distribution peak near the trailing edge without causing too
significant deceleration loss creates an ideal load
distribution.
[0035] A load distribution by reducing the number of blades for
reduced trailing-edge loss is considered below. Reducing the number
of blades while a load peak position remains fixed will change the
blade suction-side pressure distribution from a state of a solid
line in FIG. 6 to that of a dashed line. At this time, the adverse
pressure gradient from a minimum pressure at the suction side to
the trailing edge will be steep and deceleration loss will
increase. Besides, since the suction-side pressure distribution
whose influence upon performance is dominant will descend as a
whole, velocity will increase and skin friction loss will also
increase. Reducing the number of blades with the same load
distribution maintained, therefore, will actually increase overall
element performance of the blade even more significantly. In the
present embodiment, as with the load distribution represented by
the dashed line, the suction-side minimum pressure position at
which the static pressure on the suction surface of the blade
becomes a minimum is shifted more to the trailing-edge side, at the
blade hub and the blade tip, than to the intermediate position of
the blade, and the load peak position is shifted in an upstream
direction. These measures relax the suction-side adverse pressure
gradient in the neighborhood of the trailing edge, thus suppressing
deceleration loss. Accordingly, the amount of reduction in trailing
edge loss due to the reduction in the number of blades surpasses
the increment in deceleration loss due to the increase in blade
load, resulting in the blade as a whole improving in element
performance. In the present embodiment, optimizing the load
distribution in a 0.7-1.1 data range of the load coefficient .PSI.
is effective for achieving both of the improvement of blade
performance and the reduction in the number of blades at the same
time.
[0036] The relaxation of the adverse pressure gradient can be
achieved by extension of the axial chord length "Cx" as well as by
correction/modification of the load distribution. Extending "Cx"
with the same load distribution retained will extend a substantial
distance and will create a correspondingly smaller pressure
gradient "dp/dx," whereby the reduction in deceleration loss will
be realized.
[0037] The relationship between the blade load distribution and the
secondary flow developing in the neighborhood of the endwall is
described below. As can be seen from FIG. 4, a slow fluid inside
the endwall boundary layer flows from the pressure side of the
blade, towards the suction side thereof, because of the
differential pressure between both sides. This secondary flow tends
to be stronger as the differential pressure between the suction
side and the pressure side (i.e., the blade load) increases, and to
be much stronger as the differential pressure occurs at positions
shifted in the upstream direction. This means that the blades of
the aft-loading type that are smaller in upstream differential
pressure (blade load) will suffer less curling/swirling of the
endwall boundary layer, decreasing in secondary-flow loss. For the
reduction in secondary-flow loss, an aft-loading distribution is
advantageous in performance.
[0038] The relationship between the blade shape and the load
distribution is described below referring to FIG. 7. FIG. 7 shows
the meridional shape of the nozzle blade according to the present
invention, and the cross-sectional shape of the blade hub and
intermediate vertical portions. As set forth above, the axial chord
length of each nozzle blade according to the present invention is
greater at the hub and tip than at the intermediate vertical
portion. On the other hand, at the hub and the tip, the maximum
load relative position "xp/Cx" of the blade load is shifted more to
the trailing edge, and at the intermediate vertical portion, rather
exists in the upstream direction (in the present invention, near
the intermediate chord). The peak position of the load has a close
relationship to a camber line that is a center line of blade
thickness, so that as the peak position shifts more to the upstream
side, the camber line turns more in the upstream direction and so
that in consideration of the fact that inlet angle and exit angle
of the blade are determined from specifications, a large profile
stagger angle is needed to obtain predetermined axial chord length.
Simultaneously with this, circumferential chord length that is a
circumferential component of chord length increases as the maximum
load relative position shifts more to the upstream side. It follows
from this that since the nozzle blade of the present invention
approaches aft-loading at the hub side, the stagger angle .gamma.h
of the blade becomes smaller than a stagger angle .gamma.m of the
intermediate vertical portion and at the same time, the
circumferential chord length also decreases. These also apply to
the tip portion omitted from FIG. 7; the tip portion also becomes
smaller than the intermediate vertical portion in stagger angle,
and the circumferential chord length decreases.
[0039] Advantageous effects of the present invention are described
below referring to FIG. 5C. For reduced profile loss near the
intermediate vertical portion, the nozzle blade of the present
invention uses a minimum necessary number of blades so that the
maximum load relative position near the intermediate vertical
portion is shifted more to the upstream side than at the hub and
the tip. This suppresses an increase in deceleration loss near the
intermediate vertical portion due to the reduction in the number of
blades, and reduces trailing edge loss through the reduction in the
number of blades. Any rates of loss near the intermediate vertical
portion are therefore reduced relative to those of the conventional
example. Additionally, the axial chord length is extended relative
to that of the conventional example, and this extension avoids an
increase in load coefficient due to the reduction in the number of
blades. Furthermore, the load peak position at the hub and tip is
shifted more to the trailing edge than at the intermediate vertical
portion, such that secondary-flow loss is also reduced.
Accordingly, secondary-flow loss in the vicinity of the hub and the
tip is reduced. The two advantageous effects, namely the reduction
in profile loss of the main flow and the reduction in
secondary-flow loss near the endwall, significantly improve the
overall performance of the blade.
[0040] Simultaneously with the improvement of performance, the
number of blades is also reduced in the present invention.
Manufacture of the blades form a main part of steam turbine
materials, machining, and assembly costs. Adoption of the present
invention also enables significant reduction in steam turbine
manufacturing costs.
[0041] Next, steam turbine equipment applying the turbine nozzle
blade of the present embodiment is described below. FIG. 10 shows
an example of a schematic system diagram of the steam turbine
equipment according to the present embodiment. The steam turbine
equipment 23 includes a boiler 18 that is a steam generator for
firing a fuel, heating feedwater, and generating steam, a steam
turbine that includes a high-pressure turbine 19,
intermediate-pressure turbine 20, and low-pressure turbine 21
driven by the boiler-generated steam, and a condenser 22 that
condenses the steam released from the low-pressure turbine 21. New
feedwater that has resulted from the condensation by the condenser
22 flows down a feedwater line not shown, and is supplied to the
boiler 18. The high-pressure turbine 19, the intermediate-pressure
turbine 20, and the low-pressure turbine 21 are interconnected via
a turbine rotor 4, with a power generator 24 being connected at one
end of the turbine rotor 4.
Second Embodiment
[0042] A second embodiment of the present invention is shown in
FIG. 8. Differences from the foregoing embodiment are mainly
described below.
[0043] The second embodiment relates to a blade of large H/Cx, a
ratio between blade height H and axial chord length "Cx." This
cascade is equivalent to long blades having a significant
difference in radius between the blade tip and hub. In this case,
in order to minimize the difference in "t/Cx" (i.e., the difference
in load coefficient) between the tip and hub of the blade, "Cx" has
been traditionally increased according to a particular radius R.
When the present invention is applied to such a case, the hub
includes a portion at which, as the radius increases, "d Cx/d R"
decreases relative to the distribution of "Cx" that originally has
a tendency to increase, and the tip includes a portion at which, as
the radius increases, "d Cx/d R" also increases more than at a
mid-span position. In other words, the present embodiment is
characterized in that "d Cx/d R" is larger at the tip than at the
mid-span position, and in that "d Cx/d R" is smaller at the hub
than at the mid-span position. Unlike the foregoing embodiment, the
present embodiment does not always make it necessary for the axial
chord length itself of the hub or tip to be greater than that of
the mid-span position.
[0044] The adoption of the present invention reduces any
secondary-flow loss near the endwall without causing an increase in
loss at the mid-span position, even in a cascade of relatively long
blades, thus enabling more efficient, less expensive axial-flow
turbines.
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