JPS6147285B2 - - Google Patents

Description

[Detailed description of the invention] The present invention relates to a turbine stage structure.

In axial flow fluid machines such as turbines and compressors, one stage is constituted by a pair of stationary blades (hereinafter referred to as stationary blades) and rotary blades (hereinafter referred to as moving blades). The stator blades and rotor blades that make up this paragraph are designed based on fluid dynamic design to ensure that the flow conditions (flow pattern) in the flow path and the blade shape are designed to effectively convert fluid energy into rotational energy. It is determined. However, regarding the movement of the working fluid in an actual axial fluid machine, there are a) effects of fluid viscosity, especially losses due to the velocity boundary layer near the channel walls, and b) loss due to the boundary layer on the blade surfaces of the stator blades and rotor blades. The development causes a velocity deficit condition in the wake of the wing;
Due to these technical problems, it is still not possible to effectively convert fluid energy into rotational energy with current technology.

The present invention aims to reduce loss and improve efficiency by improving the shape of stator blades in the turbine stage. In order to achieve the above object, the present invention starts from actually measuring and analyzing the condition of the working fluid in the turbine stage as described below.

Conventionally, in an axial fluid machine, as shown in FIG. 1, a stator blade 1 and a rotor blade 2 are provided in a working fluid flow path, and the stator blade 1 is provided with side walls 3, which constitute the flow path.
4, and the rotor blade 2 is attached to a rotating disk 5. In this axial fluid machine, three planes AA, B- perpendicular to the radius passing through the center of the stationary blade 1
The cross-sectional shapes taken along lines B and CC and the relative positions of the stationary blades 1 and the rotor blades 2 are as shown in FIG.

FIG. 3 shows the state of flow between the stator blades and rotor blades in the turbine. Stator blade 8 and moving blade 9
are provided at equal intervals over the entire circumference, and in this figure, five stationary blades 8, 8...8 and one rotor blade 9 are illustrated.

Arrows V ₁ and V ₂ indicate the flow direction and velocity of the fluid after passing through the flow path between the stationary vanes 8, 8, . . . 8. As shown in the figure, the flow velocity V ₂ in the wake of the stator blades 8, 8... ₈ is smaller than the flow velocity V 1 in the center of the flow path between the stator blades 8, 8...8. The flow state causes periodic changes in flow velocity over the entire circumference, the number of which is equivalent to the number of stationary blades.

The state in which the flow downstream of the stationary blade, which has a speed varying depending on its position on the circumference, flows into the rotating rotor blade 9 is shown by a speed triangle as shown in b in the same figure. Arrow U represents the circumferential speed of the rotor blade 9. When fluid flows into the rotor blade 9 moving as shown by the arrow U at the speed of the arrow _V1 , the velocity of the fluid when viewed from the rotor blade 9 as a reference becomes the arrow _W1 . Similarly, when the fluid flows in at the speed of arrow V ₂ , the velocity of the fluid with respect to the rotor blade 9 becomes arrow W ₂ .

As mentioned above, the velocity of the fluid with respect to the stationary blade is V ₁ ,
When the velocity with respect to the rotor blade is W ₁ , the inlet angle with respect to the rotor blade is β ₁ . Similarly, the speed relative to the stationary blade is
When the velocity relative to the rotor blade is W ₂ at V ₂ , the inlet angle to the rotor blade is β ₂ .

Therefore, when the rotor blades are rotating, looking at any one rotor blade, the velocity of the fluid flowing into this rotor blade is the same number of times as the number of stationary blades 8, 8...8 during one rotation. It fluctuates periodically between V ₁ and V ₂ .

Along with this, the inflow angle also changes periodically between β ₁ and β ₂ .

In the above state of flow inside a paragraph,
Periodic fluctuations in the inflow angle and flow velocity of the fluid flowing into the rotor blade have a large effect on efficiency, and when there is this periodic fluctuation, the performance of the rotor blade decreases compared to when there is no periodic fluctuation. It has been known. The rotor blade airfoil loss coefficient when there is no periodic variation is ζ _PB ,
It is known that the following relational expression holds true when the rotor blade airfoil loss coefficient when there is periodic variation is ζ _PU . (LAShubenko-Shubin:
Energomashinostroenie, 1912-1) ζ _PU = (1 + K) ζ _PB ...(1) In the above equation, the loss addition coefficient K is a function of the stator blade wake parameters B, U/V ₁ and α, and for B, There is a relationship between Here, Cf is a constant, ζ _PN is the airfoil loss coefficient of the stator blade, ΔZ is the distance in the turbine axial direction from the trailing edge of the stator blade to the front and rear ends of the rotor blade, t is the distance between adjacent stator blades, and α is the stator blade is the outflow angle at .

When trying to improve the stage efficiency of a conventional axial flow machine based on the above equation, there are limited degrees of freedom for t, α and U/V ₁ due to hydrodynamic design constraints. Therefore B
In order to make it small, it is necessary to make ΔZ large and ζ _PN small. However, increasing ΔZ increases the size of the entire turbine, which causes various problems, so the most effective means is to reduce the vane shape loss coefficient ζ _PN of the stationary blade.

The above is a general theory regarding the additional loss generated in the rotor blades due to the mutual flow conditions between the stator blades and the rotor blades, and its reduction. Shown in Figure 4. Figures a and b show locations where the flow velocity was measured.

The arrow x in FIG. 4a is the stator blade wake axis, and point 0 is a point on the wake axis where the distance in the turbine axial direction from the stator blade rear end is ΔZ ₁ . The measurement of the flow velocity is performed at each point in the +y direction and the -y direction at right angles to the stator blade wake axis with this point 0 as the center. As shown in FIG. 4b, the a-a plane passing through the center of the blade length of the stationary blade 10,
Measurements are taken at three locations: the bb plane near the side wall of the stator blade tip and the C-C plane closer to the side wall of the stator blade tip. Figure 4c is a graph showing the measured values, where the horizontal axis represents the measurement position on the straight line +y--y in Figure 4a, and the vertical axis represents the flow velocity at the center between the stationary blades of the measured flow velocity V. (V ₁ above). The three curves a, b, and c are shown in Figure 4b, respectively.
Measured values on the aa plane, bb plane, and cc plane are shown. As shown in this chart,
Compared to the wake generated on the a-a plane at the central portion in the blade span direction, the wake generated on the bb plane and further on the cc plane as it approaches the side wall 11 has a wider range and the flow velocity decreases significantly.

Next, if we change the measurement position for a certain plane, for example the b-b plane mentioned above, and show the change, the fourth
As shown in Figure d. This chart shows measurements similar to those in FIG. 4c for points at distances ΔZ ₁ , ΔZ ₂ , and ΔZ ₃ from the rear end of the stationary blade 10 in FIG. 4a. In this way, the flow velocity becomes more uniform as it moves away from the stationary blade 10, and the fluctuation in the velocity flowing into the rotor blade decreases.

In this way, the wake is large in the vicinity of the side wall 11, the airfoil loss coefficient of the stationary blade increases, and the velocity at the center of the wake becomes smaller than the velocity V ₁ at the center of the flow path between the blades. This increases the fluctuation in the velocity flowing into the rotor, resulting in an increase in additional loss in the rotor blades. It is widely known that this phenomenon near the sidewall of stationary blades is caused by the secondary flow that occurs in the flow path between the blades. Based on the above facts, it is important to increase the size of the equipment in order to improve the stage efficiency. To increase the above-mentioned ΔZ dimension only in the vicinity of the side wall where wake generation is significant, without changing the axial distance ΔZ between blades that would affect is considered to be an effective method.

In this case, it may be possible to improve the stage efficiency by increasing the ΔZ dimension not only near the side wall but also over the entire length of the blade, but increasing the ΔZ dimension over the entire length may increase the overall geometry of the turbine device. Therefore, in the present invention, consideration is given to increasing the ΔZ dimension only near the side wall so as not to affect the shape and dimensions of the entire turbine device, and further, by increasing the ΔZ dimension near the side wall. It is also expected that this will have the effect of improving the flow distribution in the blade span direction.

The present invention focuses on the above point, and reduces the airfoil loss of the stator blade by appropriately setting the chord length of the stator blade and the axial distance between the stator blade and the rotor blade at each position in the blade span direction. It is also characterized by reducing additional loss of the rotor blades.

Next, one embodiment of the present invention will be described with reference to FIG. Three planes A-A, B-B, C- shown in figure a
A cross section taken along line C is shown in figure b.

The chord length of the stationary blade 13 is B-, which is the center part in the blade span direction.
It is maximum on plane B, and therefore the axial distance between the blades ΔZ
is the minimum distance ΔZ ₀ on this B-B plane. The chord length of the stationary blade 13 becomes smaller as it approaches the side wall 15 of the blade tip from the B-B plane, and therefore ΔZ increases, and the axial distance between the blades increases in the c-c plane near the side wall 15 of the blade tip. It is becoming ΔZ _t .

In addition, the chord length of the stationary blade 13 becomes smaller as it approaches the side wall 16 at the blade root from the B-B plane, and therefore ΔZ increases, and the axial distance between the blades increases in the A-A plane near the side wall 16 at the blade root. ΔZ _h . In this configuration, the BB cross-sectional shape 13B of the stator blade 13
It is determined that the AA cross-sectional shape 13A and the CC cross-sectional shape 13C near the side wall are each smaller than the above. Two specific examples of this shape are shown in FIGS. 6a and 6b, respectively. In the embodiment shown in FIG. 6a, the arrow height line 19 of the A-section is moved to the rear side compared to the airfoil height line 18 of the B-B section. Further, in this embodiment, the arrow height line of the CC cross section coincides with the arrow height line 19 of the A-A cross section. And cross-sectional shape 13A and cross-sectional shape 13
C is made to have a similar shape to the cross-sectional shape 13B, sharing most of the contour line on the back surface. however,
In this case, similar shapes are not limited to strictly geometrically similar relationships, but mean shapes drawn using the same fluid-dynamic method but with slightly different applied factors. .

In the embodiment shown in FIG. 6B, the cross-sectional shapes 13A' and 13C' near the side walls are similar to the cross-sectional shape 13B' at the center so that they share the arrow height line 20.

In the above two embodiments FIGS. 6a and b,
The narrowest width of the inter-blade flow path is the minimum value S ₀ at the center section B--B plane in the blade span direction. Then, as one approaches the A-A plane or the C-C plane near the side wall from the B-B plane, the chord length of the stationary blade decreases from _C0 to C, and the narrowest width of the inter-blade flow path becomes S It increases gradually from ₀ to S.

In the turbine having the stage structure configured as described above, the following effects are produced.

Reduction of stator vane airfoil loss In the configuration of the present invention, as explained with reference to FIG. 6, the chord length C of the stator blade near the side wall is made smaller than the chord length C ₀ of the central portion of the stator blade. Further, the flow passage width S at the rear end of the stator blade near the side wall is made larger than the flow passage width S ₀ at the rear end of the stator blade in the central portion of the stator blade.

It is known that stator blades have blade cascade characteristics as shown in FIG. The vertical axis of this chart represents the airfoil loss coefficient ζ _PN of the stator blade, and the horizontal axis represents S/t
represents the value of The three curves show the cases where the circumferential spacing/chord length of the stationary blades is (t/C) ₁ , (t/C) ₂ , and (t/C) ₃ , respectively. (t/C) ₃ > (t/C) ₂ >
(t/C) is ₁ .

The following two trends can be recognized from this chart.

(1) The number of stator blade airfoil losses decreases as S/t increases.

(2) Stator vane airfoil loss coefficient decreases with increase in t/C.

Therefore, since the stage structure according to the present invention does not make any particular change to the value of t, by applying the present invention and increasing the channel width S near the side wall of the stator blade, S/t
The value of increases. Further, by reducing the chord length C near the side wall, the value of t/C increases. From the above relationship, the vane shape loss coefficient decreases.

Equalizing the velocity of the working fluid flowing into the rotor blades As explained in Figures 4c and d, the velocity reduction rate V/V ₁ in the wake of the stator blades is extremely small near the side wall, and from the rear end of the stator blade. As the axial distance ΔZ increases, it becomes uniform and approaches 1.

Increasing the value of ΔZ causes the problem of increasing the size of the entire device, but in the present invention, the value of ΔZ is increased only near the side wall where the wake is significant, so the overall size of the device does not become large. Moreover, the flow velocity fluctuations are made uniform only for the significant wake near the side wall. As mentioned above, it is widely known that periodic fluctuations in the velocity of the working fluid flowing into the rotor blades increase the rotor blade airfoil loss, and by applying the present invention, significant portions of the wake can be made uniform. As a result, the periodic fluctuations in the rotor blade inlet velocity are reduced, and the rotor blade airfoil losses are reduced as a result.

Improving the inflow angle and flow distribution of the rotor blades In general, the outflow angle from the stator blades is given by sin ^-1 (S/t) (see Figure 6), but in reality, secondary flows occur near the sidewalls and It changes depending on the influence of the velocity boundary layer that develops during the process. Figure 8 shows a comparison of the design expected value and the actual flow situation.

The horizontal axis in Fig. 8a is the outflow angle α from the stationary blade mentioned above.
sin ⁻¹ s/t, and the vertical axis indicates the position of the stationary blade in the blade span direction. The horizontal axis in Figure b shows the ratio of each partial flow rate to the working gas flow rate at the center in the blade span direction of the stationary blade as 1, and the vertical axis represents the ratio of each partial flow rate to this figure.
Similarly, it shows the position in the stator blade span direction.

In the conventional paragraph structure, as shown in Fig. a, the mechanical outflow angle is kept almost constant at each position in the blade span direction as shown by line 21, but the actual outflow angle is as shown by line 30.
It is noticeably smaller near the side wall, as in Along with this, the flow rate near the side wall also decreases. Line 3 in this figure b
1 is the expected flow rate distribution based on design, but compared to this, the actual flow rate distribution is significantly smaller near the side wall as shown by line 23.

As described above, when the outflow angle α from the stationary blade decreases, the inflow angle β to the rotor blade also decreases. FIG. 8c shows this situation in a vector diagram. In the figure, α ₁ is the expected outflow angle in terms of design, and α ₂ is the actual outflow angle near the side wall. As the outflow angle changes in this way, the inflow angle β ₁ expected from the design perspective becomes β ₂ near the side wall. Since this inlet angle _β2 is not the best condition for the rotor blade, the rotor blade performance deteriorates.

In the embodiment of the present invention, the value of the mechanical outflow angle α=sin ^-1 (S/t) of the stationary blade is made larger near the side wall. Line 22 in FIG. 8a shows the mechanical outflow angle in an embodiment of the invention. As a result, the reduction in partial flow rate near the side wall is alleviated. Line 24 in Figure b shows the flow rate distribution in the embodiment of the invention.

In this way, line 3 where the flow rate distribution is the design expected value
By approaching β1, the inflow angle to the rotor blade also approaches the design expectation value _β1 , improving rotor blade performance.

Although the improvement in turbine performance obtained by the present invention has been qualitatively explained above, the configuration and effects when the present invention is actually applied will be quantitatively explained based on the experimental results conducted by the inventor. It is as follows.

The distance l shown in FIG. 8a indicates the range affected by the secondary flow and velocity boundary layer near the side wall. It has been experimentally confirmed that the value of l is approximately given by the relational expression l/H=1/3 (C ₀ /H). However, C ₀ is the chord length at the center of the stationary blade.
(See Figure 6).

Based on this, the mechanical flow angle of the stator vane as shown by line 22 in the figure a can be calculated from the distance l from the side wall as follows:
It increases in a parabolic manner as it approaches the side wall, and the channel width S at the position touching the side wall is equal to the channel width at the center.
Make it about 1.3 to 1.5 times compared to S ₀ . As a result, the distance indicating the influence range is reduced to l' as shown in Figure b. It has been experimentally confirmed that the following relational expression l'/H=1/6(C ₀ /H) holds true.

As mentioned above, when the value of S is set to 1.3 to 1.5 times S ₀ , in the case of a normal airfoil, the chord length C of the portion in contact with the side wall is approximately 0.6 to 0.8 of the chord length C ₀ of the central portion. In this way, the axial distance ΔZ between the stationary blade and the rotor blade increases near the side wall, but since α increases, this is not a favorable trend when looking only at equations (1) and (2) shown above. However, the present inventors have experimentally confirmed that the effects described above improve the stage efficiency of the turbine as a comprehensive result.

Figure 10 is a chart comparing the total pressure loss distribution of the stator blades for the embodiment of the present invention and the conventional example in a model turbine that is an actual three-dimensional flow field, and shows the stator blade aspect ratio (blade length/ The graph shows the results of an experiment using a single-stage turbine with chord length) = 0.6 and rotor blade aspect ratio = 1.0.

FIG. 10a shows an embodiment of the present invention, in which the outflow angles at the blade root and the blade tip were set to be approximately 5° larger than the outflow angle near the blade center.

Figure 10b shows a conventional example in which the outflow angle is almost constant (13°±1) from the blade root to the blade tip.
゜).

The illustrated curve is a series of points where the total pressure loss per pitch is equal, and the numbers 1 to 30 are a
This is an index added to compare the curve in the figure and the curve in figure b. As is clear from this figure, the loss reduction at the blade tip and blade root is particularly remarkable in this example.

FIG. 9 is a diagram analyzing the loss coefficient of the turbine, in which the vertical axis represents the position in the blade span direction, and the horizontal axis represents the loss coefficient.

The total loss coefficient ζ is the stator vane loss coefficient ζ _PN and the moving blade airfoil loss coefficient ζ _PB when there is no periodic variation in flow velocity.
is the sum of the increase in rotor blade airfoil loss coefficient due to periodic fluctuations in flow velocity (ζ _PU - ζ _PB ), and the secondary flow loss coefficient ζ _SC . The solid line in Figure 9a shows the above-mentioned total loss coefficient ζ in the conventional paragraph structure, analyzed into the above-mentioned ζ _PN , ζ _PB , (ζ _PU - ζ _PB ) and ζ _SC for each position in the blade span direction. It is something. And the dotted lines indicate the values of each loss coefficient as a result of implementing the present invention,
The shaded area indicates the loss that can be reduced by implementing the present invention. That is, as for the stator vane airfoil loss coefficient ζ _PN , t/C becomes smaller, and as S/t becomes larger, the amount of loss reduction becomes larger as the vane approaches the side wall. The rotor blade airfoil loss coefficient ζ _PB does not change. Increase in loss coefficient due to periodic fluctuations in flow velocity (ζ _PU −ζ _P
B ) is significantly reduced as the side wall approaches.
The secondary flow loss coefficient ζ _SC is reduced by reducing the vane shape loss and by increasing S/t near the sidewall and making the flow velocity distribution uniform.

Combining the above results, as shown in Figure b,
Line 25, representing the total loss factor in a conventionally constructed turbine stage, is reduced to line 26 by the practice of the present invention.

The present invention achieves the above-mentioned effects when applied to various turbine stages, but the effect is particularly remarkable in the turbine stage where the ratio of the distance l of the range affected by the side wall to the blade length H as shown in FIG. 8 is large. , and the effect is remarkable in a turbine stage where the chord length C ₀ is large relative to the blade length H.

In the case of a stage where 2l≧H, that is, a stage where l overlaps, and 0.3<H/C ₀ <1.5, the turbine stage efficiency can be improved by 2 to 3% compared to the conventional example by applying the present invention. is possible.

[Brief explanation of the drawing]

FIG. 1 is a partially sectional view showing an example of the configuration of a turbine stage commonly used in the past, and FIG.
C sectional view, Figure 3 is a diagram explaining the state of the flow between the stator blade and the rotor blade, a is the flow velocity distribution diagram downstream of the stator blade, b is a vector analysis diagram, and Figure 4 is the flow downstream of the stator blade. Figure 5 is a diagram explaining the loss occurrence situation, where a shows the flow velocity measurement point in the flow direction, b shows the flow velocity measurement point in the blade span direction, c and d are charts showing the flow velocity distribution, and Figure 5 is a diagram showing the flow velocity measurement point in the blade span direction. In the drawings for explaining the configuration of an embodiment of the invention, a is a partially sectional view corresponding to FIG. 1, b is a sectional view corresponding to FIG. 2, and FIG. 6 is a sectional view corresponding to FIG.
Figures a and b are cross-sectional views showing the shape of a stator vane in an embodiment of the present invention, Figure 7 is a chart showing the relationship between the blade row configuration of the stator blade and airfoil loss, and Figures 8 a and b are A chart showing the situation of performance improvement in the embodiment of the present invention, FIG. 9c is a vector analysis diagram of the same, FIGS. In order to explain the effect of the present invention, it is a chart showing the average total pressure loss in comparison with a conventional example. 13... Stator blade, 13A, 13C... Cross section near the side wall of the stator blade, 13B... Cross section at the center of the stator blade, 15... Side wall at the tip of the stator blade, 16... Side wall at the root of the stator blade, 18, 19, 20... Arrow height lines in the cross section of the stationary blade.

Claims (1)

[Claims] 1. In a step structure consisting of stator blades and rotor blades of an axial fluid machine, the chord length of the stator blade is maximized at the center of the blade span, and from the center of the blade span to the blade tip and The blade is configured to gradually contract as it approaches the side wall of the blade root, and the distance from the stator blade outlet end to the rotor blade inlet end is the minimum at the center of the blade length, and the distance from the blade center to the blade tip and blade root is A turbine stage structure configured to gradually increase in size as it approaches a side wall. 2. The stationary blade is characterized in that the stationary blade has a three-dimensional shape such that the projection of the cross-sectional shape by an arbitrary plane perpendicular to the radius passing through the stationary blade results in a similar shape that shares most of the arrow height line. A turbine stage structure according to claim 1. 3. The stator vane is characterized in that it has a three-dimensional shape such that the projection of the cross-sectional shape by an arbitrary plane perpendicular to the radius passing through the stator vane results in a similar shape that shares most of the rear contour line. A turbine stage structure according to claim 1.