JPH0580594B2 - - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION Field of Industrial Application The present invention applies generally to impellers of turbomachines, in particular centrifugal compressors and centrifugal blowers.

BACKGROUND ART FIGS. 3 and 4 are a front view of a centrifugal compressor and a sectional view taken along a plane including an axis.

In FIGS. 3 and 4, the impeller 1 consists of blades 11, 12, etc. Also, impeller 1
There is a differential user 2 formed of differential user blades 21 on the outer periphery of the differential user blade 21 . Note that the shaft for rotating the impeller 1 is designated by 4, and the casing is designated by 3. According to such a configuration, the blades 1 of the rotating impeller 1
1, 12, . . . , the velocity energy of the working fluid is converted into pressure by the deceleration passage formed by the diffuser vanes 21 of the diffuser 2.

In this way, the blades 11, 12 of the impeller 1...
The differential user blade 21 of the differential user 2 performs the above-mentioned energy conversion by finely adjusting the shape in the blade width direction and the blade surface shape. Among these, the most important adjustment item is the blade 1
1, 12... and the inlet angle of the diffuser blade 21.

Therefore, FIGS. 5a to 6b are diagrams for explaining the design method according to the prior art. Fifth
Figures a and 5b show the distribution of the meridian velocity (fluid velocity component projected onto a plane passing through the axis) in the blade width direction at the outlet portions of the blades 11 and 12 (see Figures 3 and 4). It is a conceptual diagram. Moreover, FIG. 6a and FIG. 6b are respectively 5a and 6b.
FIG. 5B is a vector diagram showing a velocity triangle between the shroud side s and the hub side h at the outlet portions of the vanes 11 and 12 corresponding to FIGS. Furthermore, Section 5a
Both figures and Fig. 6a show the rated flow point [Q/
Q ₀ = 1.0], and Figures 5b and 6b both show the point at which the flow rate is decreased at the same rotational speed [Q/
Q ₀ = 0.85].

The symbols shown in Figs. 5a to 6b are as follows: C _n2 : Meridian velocity at the vane exit C ₂ : Absolute velocity at the vane exit W ₂ : Relative velocity at the vane exit U ₂ : Peripheral speed at the vane exit α ₂ : Absolute outflow angle at the vane exit β ₂ : Indicates the relative outflow angle at the blade exit. The subscripts s and h added to these symbols indicate the shroud side and the hub side of the blade outlet.

At the rated flow rate [Q/Q ₀ = 1.0], as shown in Figures 5a and 6a, the meridian velocity C _n2 is approximately constant in the blade width direction (s→h) (macroscopically ignoring minute boundary layers, etc.). ). moreover,
Generally, the exit angles of the vanes 11 and 12 are constant, and the outflow angle of the fluid exiting the vanes 11 and 12 with respect to the vanes; the relative outflow angle β ₂ is approximately constant. That is,
The following relational expression holds true.

β _2s ≒β _2h ……(1) C _n2s ≒C _n2h ……(2) Therefore, the velocity triangle becomes as shown in Figure 6a,
Since the circumferential speed U ₂ is constant, the absolute outflow angle α ₂ at the blade outlet (inflow angle to the differential user inlet) is also constant.

α _2s ≒ α _2s ……(3) Due to the relationships shown in equations (1), (2), and (3), good performance can be achieved by appropriately combining the exit angle of the blade and the entrance angle of the differential user. be done.

Problems to be solved by the invention However, when the rotational speed remains constant (U ₂ constant),
As the flow rate decreases (for example, Q/Q ₀ = 0.85
), the distribution of the Meridian velocity in the blade width direction shown in Figure 5a does not decrease uniformly, but rather decreases on the hub side due to the centrifugal force balance of the large flow direction change in the Meridian cross section of the impeller. The ratio becomes larger than that on the shroud side, and the shroud is deformed as shown in FIG. 5b. Therefore, the speed triangle in this case becomes as shown in FIG. 6b, and the absolute outflow angle α ₂ changes to the increasing side, but the amount of increase is particularly large on the hub side.

That is, if the entrance angle of the differential user blade is fixed constant, the angle of incidence on the differential user blade on the hub side will become excessive, leading to stalling or, in extreme cases, surging at the differential user blade.

This phenomenon will be explained more quantitatively. In other words, at the rated flow point (Q/Q ₀ = 1.0) in Figure 5a,
From the C _n2 distribution state, the flow rate point (Q/Q ₀ =
0.85), Meridian velocity C _n2s is 0.90 and C _n2h is 0.80 (both Q/
Q ₀ =1.0), the velocity triangle of both flow rates becomes as shown in Fig. 6b. The conditions for the trial calculation in Figure 6b are Q/Q ₀ = 1.0.
C _n2 /U ₂ = 0.25, β ₂ = 45°. That is, when Q/Q ₀ = 1.0, α _2s = α _2s = 71.6° When Q/Q ₀ = 0.85, α _2s = 73.8° α _2h = 76.0°, and α when Q/Q ₀ = 1.0 By adapting the differential user entrance angle to 71.6° for _2s and α _2h , Q/
When Q ₀ = 0.85, the differential user inflow angle on the shroud side deviates from the optimal value by (73.8° - 71.6°), or 2.2°, while on the hub side, it deviates from the optimal value by (76° - 71.6°), or 4.4°. become.

The deviation of this inflow angle from the optimal value, that is, the differential user incidence angle, is Δα ₂ = (α ₂ − α ₂ ^* ), (where α ₂ ^*
is the optimal inflow angle). This differential user incident angle and differential user loss coefficient ratio (optimum incident angle)
1.0) is approximately as shown in Figure 7, and according to this, the loss of the differential user for the streamline corresponding to the hub side of the blade is approximately twice as much as that at the optimum angle of incidence, effectively stalling. state. on the other hand,
For the streamline on the shroud side, Δα ₂ ≒2°,
Since the amount of increase in loss is very small, no stall occurs.

As mentioned above, in a centrifugal compressor designed so that the outlet outflow angle of the blades 11 and 12 (≒ blade outlet) and the diffuser blade inlet angle are constant, even if good performance is obtained near the rated flow rate, the On the flow rate side, the angle of incidence at the diffuser inlet corresponding to the hub side of the vane outlet becomes excessive. This has caused problems such as deterioration of compressor performance or early occurrence of surging (with a small amount of flow reduction).

Although there is a slight deviation between the above-mentioned outflow angle of the blade at the impeller outlet and the blade exit angle, there is no large difference in the blade width direction (s→h). This deviation is only related to the adjustment of α ₂ ^* (difference user setting angle), and has no special meaning for changing the exit angle of the impeller in the blade width direction, which is the aim of the present invention. This deviation will not be discussed herein.

Means for Solving the Problems The present invention aims to solve the above-mentioned problems. That is, the present invention provides a turbomachine including an impeller and a differential user (including one without blades), in which the blade exit angle β _2h on the hub side of the impeller is changed to the blade exit angle β _2s on the shroud side.
and the difference angle between the blade exit angle on the hub side and the blade exit angle on the shroud side (β _2h −
β _2s ) is set so that 2°≦β _2h −β _2s ≦12°.

Effect: According to the means described above, it is possible to reduce the deviation (within 3 degrees) of the diff user incident angle from the optimum value on the hub side when the flow rate is reduced. As a result, a centrifugal compressor that can operate stably even in the small flow range and has high performance in the rated flow range is obtained.

Embodiment The structure of the turbomachine according to the present invention is a conventional example (third example).
4), detailed explanation thereof will be omitted here.

The gist of the present invention is to define the relationship between the blade exit angle β _2s on the shroud side of the impeller and the blade exit angle β _2h on the hub side as described above, β _2s < β _2h ... 2゜≦β _2h - β _2s ≦12゜... The basis for determining the formula of the above condition will be explained as an example.

FIG. 2 shows the impeller according to the present invention, Q/
Q/Q is the condition under which the differential user incidence angle Δα _h on the hub side is 3° or less even when Q ₀ = 1.0 to Q/Q ₀ = 0.85.
When Q ₀ = 1.0, C _n2 /U ₂ is usually within the range of 0.3 to achieve good compressor or blower performance.
This is a graph obtained by changing the hub side exit angle β _2h of the impeller with respect to 0.25.

In Figure 2, for an impeller with a hub side outlet angle β _2h of 40°, C _n2 /U ₂ = at the rated flow rate (Q/Q ₀ = 1.0).
When operating the machine at 0.25, the small flow rate side (Q/Q ₀
= 0.85), in order for the differential user attack angle on the hub side to be within 3°, the shroud side exit angle β _2s
This shows that it is sufficient to select point A as shown in FIG. Furthermore, it is shown that even when the machine is operated with C _n2 /U ₂ =0.3, it is sufficient to determine the shroud side exit angle β _2s as shown in point B. Although the differential user setting angle α ₂ ^* is not described in Fig. 2, this α ₂ ^* is determined at the rated flow rate [Q/Q ₀ =
1.0] can be easily set by drawing a velocity triangle.

In addition, the range shown in Figure 2 shows that the diff user incidence angle on the hub side is when the flow rate at the operating point decreases by 15% from the rated point and the meridian speed C _n2h on the hub side at the impeller outlet decreases by 20%. This shows the conditions under which the angle does not exceed 3°. However, the above-mentioned rated flow point [Q/Q ₀ =1.0] is merely a standard of flow rate, and is not essentially fixed in many cases.

Therefore, the key point of the present invention is that within a certain range of flow rate change to a small flow rate side (within about 15%),
In particular, it is an object of the present invention to provide a general means for avoiding an excessive increase in loss and stall in the differential user section due to the generation of an excessive incident angle of the differential user on the hub side. In other words, if β _2h is larger than β _2s and the difference is at least 2° or more, a certain effect is recognized compared to the conventional technology (β _2h ≒ β _2s ).
This shows that the maximum value of (β _2h −β _2s ) does not exceed 12°.

Next, the effects of the present invention will be explained in Fig. 1a and Fig. 1.
This will be explained in detail with reference to figure b. Figures 1a and 1b
Both figures illustrate the action of the impeller according to the invention by means of a velocity triangle at the impeller outlet. However, the drawing shows C _n2 / at the rated flow point [Q/Q ₀ = 1.0].
This is the case when U ₂ =0.25 and β _2s =45°.

In Figure 1a, the relative outflow angles β _2s = 45°, β _2h =
53.7°, the absolute flow angle α ₂ at the rated flow point (Q/Q ₀ = 1.0) is α _2s = 71.565°, α _2h =
It becomes 69.247°. That is, the blade exit absolute outflow angle α _2h is smaller than α _2s by 2.318°, and if the optimal inflow angle α ₂ ^* of the differential user blade is set to about 71.6°, the differential user incident angle (Δα ₂ = α _2h − α ₂ ^* ) is -2.4°, and as seen from FIG. 7, the differential user loss is not so excessive as compared to when Δα ₂ =0, indicating good performance.

On the other hand, at the point of Q/Q ₀ = 0.85, C _n2s and C _n2h are each 0.9 from the case of rated flow rate [Q/Q ₀ = 1.0].
and 0.8 times, resulting in a velocity triangle as shown in Figure 1b. The flow rate in FIG. 1b is Q/Q ₀ =0.85. In this Fig. 1b, the absolute outflow angle α _2s =
Since the absolute outflow angle α _2h on the hub side is 73.8° and 74.6°,
The diff user incident angle Δα ₂ is Δα _2h =74.6°−71.6°=3.0° Δα _2s =73.8°−71.6°=2.2°. According to this, since any incident angle Δα ₂ of the differential users is approximately 3° or less, no excessive differential user loss occurs.

The above explanation was about a turbomachine with a differential user with blades, but for a differential user without blades, the optimum inflow angle α ₂ ^* for this part of the flow is as shown in Fig. 7. has a change in loss for Δα ₂ similar to
The functions and effects of the impeller according to the present invention are also exhibited here as well.

Effects of the Invention Since the present invention has the above-described configuration, it is possible to realize a centrifugal compressor or blower with good low flow rate characteristics (less reduction in efficiency, less surging flow rate).

[Brief explanation of the drawing]

1a and 1b are vector diagrams showing velocity triangles at the impeller outlet according to a preferred embodiment of the invention, and FIG. 2 is a diagram showing the hub-side vane exit angle β _2h and the shroud-side vane exit angle of the impeller according to the invention. _FIG. 3 is a front view showing the impeller and differential user of a centrifugal compressor according to an embodiment of the present invention and a conventional example, and FIG. 5a and 5b are conceptual diagrams illustrating the distribution state of radial meridian velocity at the impeller outlet according to the prior art, and FIGS. 6a and 6b are diagrams 5a and 5b, respectively. A vector diagram showing a velocity triangle at the corresponding impeller outlet, and FIG. 7 is a conceptual diagram showing the relationship between the differential user incident angle Δα ₂ and the differential user loss coefficient ratio. 1... impeller, 2... differential user, 11, 12...
Vane, 3...Casing, 21...Diffusion user vane, 4...Shaft, C _n2 ...Blade exit Meridian speed, C ₂
...Blade exit absolute speed, W ₂ ...Blade exit relative speed,
U ₂ ... Peripheral speed at the blade exit, α ₂ ... Absolute outflow angle at the blade exit (inflow angle to the differential user), β ₂ ... Relative outflow angle at the blade exit (blade exit angle), s... Shroud side, h... Hub side.

Claims (1)

[Claims] 1. In a turbomachine including an impeller and a differential user, the blade exit angle β _2h on the hub side of the impeller is larger than the blade exit angle β _2s on the shroud side, and the blade exit angle on the hub side and the The difference angle (β _2h − β _2s ) from the blade exit angle on the shroud side is 2°≦β _2h
A turbomachine characterized in that −β _2s ≦12°.