JPS5944482B2 - axial turbine - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION The present invention relates to an axial flow turbine, and more particularly to an axial flow turbine having a unique shape in its rotor blades.

Generally, in an axial flow turbine used as a steam turbine, etc., as shown in FIG.
2 and the nozzle outer ring 13, and the nozzle outer ring 13 is fixed to the casing 14 to constitute a stationary part, and the disk 1 is integrated with or fitted into the rotating shaft 15.
A rotor blade 17 is installed on the rotor 6 to constitute a rotating section, and the stationary section and the rotating section form one stage.

An axial flow turbine is constructed by combining one or more of these stages in the axial direction.

In such an axial flow turbine, a working fluid such as steam or combustion gas flows approximately axially through an annular passage formed by the outer surface of the nozzle inner ring 12, that is, the nozzle inner wall 18, and the inner surface of the nozzle outer ring 13, that is, the nozzle outer wall 19. After passing through the nozzle blades 11 and being given sufficient swirling force, it enters the rotor blades 17, where the swirling force of the working fluid is converted through the rotor blades 17 into the rotational force of the rotating shaft 15. Ru.

The working fluid then flows out from the rotor blades 17 and flows into the next stage, where the same operation is repeated and the energy of the working fluid is converted into power for the rotating shaft.

On the other hand, in a flow path having an arcuate curved surface such as a nozzle blade of an axial flow turbine, a pressure distribution occurs in the flow path based on the curvature of the streamline, and the pressure is The pressure is high, and the pressure is low on the nozzle blade dorsal side 11b.

This pressure distribution is controlled by the flow velocity of the mainstream and the curvature of the streamline, but the flow velocity near the nozzle inner wall 18 and nozzle outer wall 19 is higher than the flow velocity of the mainstream due to the viscosity of the working fluid, which causes a boundary layer flow. As a result, sufficient centrifugal force to counter the pressure difference cannot be obtained, and the force is generated from the nozzle blade vent side 11a to the nozzle blade back side 11b along both the inner and outer wall surfaces.
A secondary flow 20 directed toward is induced, and a secondary flow vortex 21 is generated.

This secondary flow vortex 21 causes a large deviation in the outflow angle distribution of the working fluid from the nozzle blade 11 from the designed value.

FIG. 3 is a diagram showing the outflow angle distribution of the working fluid from the nozzle blade, and the broken line A indicates the geometric outflow angle, that is, the design value,
Moreover, the solid line B shows the actually measured value.

As is clear from Fig. 3, the outflow angle of the fluid in the vicinity of the nozzle inner wall and nozzle outer wall is significantly deviated from the design value due to the influence of secondary flow, and the operation from this nozzle blade is The deviation of the fluid outflow angle from the design value results in a subsequent deviation of the inflow angle to the rotor blade 1γ from the design value, which causes a decrease in turbine efficiency.

FIG. 4 is a diagram showing the relationship between the outflow angle from the nozzle blade 11 and the inflow angle to the rotor blade 17 in the vicinity of the wall surface. At the time of design, as shown by the broken line, the nozzle outflow angle αA1 changes to the nozzle exit velocity. The working fluid that flows out of the nozzle blade 11 at a relative outflow speed ωA from the circumferential speed U of the blade flows into the rotor blade 17 at a relative inflow speed ωA with a relative entrance angle βA designed to be an optimal value.

However, as shown by the solid line, the actual flow of the working fluid near the wall has a nozzle outflow angle α3 that is increased with respect to the design value of the nozzle exit speed CB1, and flows out at a relative outflow speed ωB. With an inlet velocity ωB, it flows into the rotor blade 17 at a relative inlet angle βB which is far from the optimum relative inlet angle β force.

Incidentally, the blade cascade performance of the rotor blade is sensitive to changes in the inlet angle of the working fluid, especially when the turning angle is large such as at the blade root, and the experimental data on the blade cascade loss shown in Figure 5 shows that As is clear, when the inflow angle β of the working fluid deviates from the optimum inflow angle βA to βB, the blade row loss W increases significantly.

Analyzing the loss distribution of the turbine stage in this way, we find that
An increase in loss due to secondary flow near the nozzle inner wall 18 and nozzle outer wall 19 is clearly recognized, and the decrease in turbine efficiency due to this secondary flow accounts for a particularly large proportion of the turbine internal loss, and the energy Effective countermeasures are required from the perspective of effective utilization of

The present invention has been made in view of these points, and it is an object of the present invention to provide an axial flow turbine that can prevent an increase in loss due to secondary flow and improve turbine efficiency.

An embodiment of the present invention will be described below with reference to the drawings.

FIG. 6A is an enlarged perspective view of a rotor blade of an axial flow turbine according to the present invention, and FIG. 6B is a comparison view of its main cross section.

The rotor blade 27 according to the present invention has an inlet angle of the rotor blade at the root portion 2
It gradually increases or decreases from 7a to 27b at a predetermined height Hl, and then increases gradually, and from the position 27b to the position 27c, which is a predetermined height H2 lower than the tip 27d, corresponds to an increase in the rotational circumferential speed due to an increase in radius. It increases in size and is twisted so that it becomes even larger from the position 27c toward the tip 27d.

FIG. 7 is a distribution diagram showing changes in the rotor blade inlet angle β corresponding to the height of the rotor blade 27, where the rotor blade inlet angle distribution according to the present invention is shown by a solid line Y1, and the conventional rotor blade inlet angle distribution is shown by a broken line X. ing.

The conventional rotor blade inlet angle distribution X corresponds to an increase in rotational circumferential speed due to an increase in radius, whereas the rotor blade inlet angle distribution Y according to the present invention changes the inlet angle at the rotor blade tip to the center of the rotor blade. The shape changes from the part to the end.

That is, the inflow angle at the root of the rotor blade is made to continuously and gradually increase from a position at a predetermined height Hl from the root of the rotor blade toward the root.

The change in the inflow angle may be such that the rate of decrease continuously and gradually decreases from the position of the predetermined height Hl toward the root portion.

Further, the inflow angle at the tip of the rotor blade is made to continuously and gradually increase from a position a predetermined height H2 below the tip of the rotor blade toward the tip.

Therefore, as shown in FIG. 7, the rotor blade inflow angles at the root and tip of the rotor blade are 1 or 2, respectively, from the rotor blade inflow angles at positions Hl and H2 away from the root and tip, respectively. It is formed so that it changes by 2.

The values of Hl and H2 are based on the thickness of the area dominated by the secondary flow on the wall surface, that is, the distance at which the actually measured nozzle outflow angle distribution begins to change rapidly near the wall surface, as shown in Figure 3. In addition, ■1. (2) The value of 2 can be determined based on the amount of change in the rotor blade inflow angle, which can be compensated for by the deviation in the nozzle outflow angle of the wall surface portion due to the secondary flow.

In this way, H and H2 shown in Fig. 7 correspond to the thickness of the area dominated by the secondary flow on the wall surface, so the outflow angle of the nozzle blade, the height of the nozzle blade, and the size of the nozzle blade It is appropriately selected depending on the shape and the like.

FIG. 8 shows an example of this relationship, where the vertical axis represents the aspect ratio Z1 of the nozzle V, that is, the ratio of the nozzle blade height to the nozzle blade chord length, and the horizontal axis represents Hl and H2.

In other words, since the area of secondary flow expands in inverse proportion to the aspect ratio Z, as the aspect ratio Z decreases, Hl and H2
This shows that the increase in

Change in rotor blade inlet angle shown in Figure 7 ■1. According to the experimental results, the values of (2) are -2° to 10°, respectively.

The angle is preferably about 10° to 22°, and is appropriately selected depending on the turning angle of the nozzle blade, the height of the nozzle blade, the chord length of the nozzle blade, the shape of the nozzle blade, etc.

According to experiments, the amount of increase in the nozzle outflow angle due to the secondary flow near the nozzle inner and outer walls shown in Figure 3 is in the range of about 2° to 7° in the turning angle range of nozzle blades used in practical use. However, the relationship between the deviation amount Δβ of the rotor blade inlet angle and the deviation amount Δα of the outflow angle is shown in FIG.

According to this Figure 9, △ for a change of Δα = 2° to 7°
The amount of change in β was obtained, and the above-mentioned (1) 0. ■Converting to 2, the change in rotor blade inflow angle ■1°■2 is -2° to 10°,
The angle is 10° to 22°, and within this range, it is possible to always set the inflow angle to the rotor blade to the optimum value.

As described above, according to the present invention, it is possible to always set the inflow angle to the rotor blade at an optimal value, it is possible to prevent a decrease in turbine efficiency due to secondary flow, and it is possible to effectively utilize energy. can.

[Brief explanation of the drawing]

FIG. 1 is a sectional view showing the configuration of one stage part of an axial flow turbine;
Figure 2 is an explanatory diagram showing the secondary flow in the nozzle blade cascade flow path, Figure 3 is a diagram showing the nozzle outflow angle distribution due to the secondary flow, and Figure 4 is the flow of working fluid from the nozzle blade to the rotor blade. FIG. 5 is a diagram showing the blade cascade loss of the rotor blade, FIG. 6A is an enlarged perspective view of the rotor blade in the present invention, and FIG. 6B is the main part of the rotor blade in the present invention. cross section,
FIG. 7 is a diagram showing the rotor blade inlet angle distribution in the present invention, and FIG.
The figure is an explanatory diagram of the start position of the change in the inflow angle of the rotor blade, and FIG. 9 is a diagram showing the relationship between the deviation of the nozzle blade outflow angle and the deviation of the inflow angle of the rotor blade. 11... Nozzle blade, 12... Nozzle inner ring, 13... Nozzle outer ring, 18... Nozzle inner wall, 19... Nozzle outer wall, 27...
... Moving blades.

Claims (1)

[Claims] 1 The inflow angle of the rotor blade root in the rotor blade is determined as follows from a position above the rotor blade root by a predetermined height toward the rotor blade root.
An axial flow turbine characterized in that it has a continuous and progressively larger configuration.