JP2730268B2 - Centrifugal impeller - Google Patents

Description

Description: TECHNICAL FIELD The present invention relates to an improvement in an impeller used for a centrifugal compressor or a centrifugal blower that pumps a fluid by rotation.

<Prior Art> Conventionally, as a centrifugal impeller, as shown in FIG. 11, a hub (core plate) 92 attached to a rotating shaft 91 and a shroud (side plate) 93 provided opposite thereto are provided. Between them, an arrangement in which a plurality of two-dimensional or three-dimensional blades 94 are arranged is provided.

Since the shape of each component of the impeller has a significant effect on the impeller efficiency, various design methods have been proposed to minimize the occurrence of loss in each component.
However, in the above-mentioned impeller, it is difficult to obtain the performance as designed because many factors are intricately entangled and affect the impeller efficiency. The reality is that the desired performance is ensured.

In conventional impellers provided, for example,
As shown in JP-A-55-134797, the length Ls of the blade 94 along the shroud surface 93a is equal to the length along the hub surface 92a.
It is set to be shorter than Lh.

Further, the blade outlet angle is usually set at 45 ° to 60 °.

<Problems to be Solved by the Invention> However, according to the above-mentioned conventional centrifugal impeller, in a low flow rate region, swirling stall or backflow of fluid occurs in a region E on the shroud surface 93a near the leading edge 94a of the blade 94. However, there is a problem that efficiency is poor because the Mach number at the blade outlet increases, the relative Mach number of the blade 94 on the shroud surface 93a side near the leading edge 94a increases, and the like.

In particular, in a turbocompressor using Freon, the speed of sound is as low as several hundred tens of meters, and the area where the Mach number exceeds 1 is increased, which has been a bottleneck in improving efficiency. Further, since the range of airflow at the point of use is wide, it is required to exhibit high efficiency over a wide airflow range. However, in reality, there has not been provided anything satisfying such a demand.

The present invention has been made in view of the above problems,
An object of the present invention is to provide a centrifugal impeller that can ensure good efficiency over a wide flow rate range.

<Means for Solving the Problems> A centrifugal impeller according to the present invention for achieving the above object is applied to a turbo compressor using a refrigerant gas, and is provided between a shroud surface 3a and a hub surface 2a. In a centrifugal impeller in which a plurality of three-dimensional blades 4 are arranged, a parallel portion 3b parallel to the rotation axis of the impeller is provided on a suction side B of a shroud surface 3a.
Is formed, and the relative Mach number is set to the following and by setting the blade shape so as to satisfy the following conditions and the following conditions.

The tip of the blade 4 on the shroud side protrudes toward the inflow side, and the blade length Ls along the shroud surface 3a has a relationship of Ls / Lh = 1.05 to 1.1 with respect to the blade length Lh along the hub surface 2a.

The leading edge 4a of the blade 4 extends toward the hub surface 2a starting from the parallel portion 3b, and the intersection between the leading edge 4a of the blade 4 and the shroud surface 3a.
The inclination angle α of the line 1 connecting C1 and the intersection A1 between the leading edge 4a of the blade 4 and the hub surface 2a with respect to the rotation axis 1 is 25 to 30 °.

The blade outlet angle β is 30 ° to 40 °.

The relative Mach number Mw at the blade inlet increases from the hub side toward the shroud side while taking a value of 1 or less.

The value obtained by dividing the average value of the relative Mach number Mw of the pressure surface and the suction surface at the blade inlet by the average value of the relative Mach number Mw of the pressure surface and the suction surface at an intermediate position between the blade inlet and the outlet is 1.6 or less. Become.

The difference between the relative Mach number Mw of the pressure surface and the suction surface is smaller than the average value of the relative Mach number Mw of the pressure surface and the suction surface.

<Operation> The inventor of the present application has found that setting the blade shape based on the distribution state of the relative Mach number on the blade surface may lead to an improvement in efficiency. The present invention has been made based on the above findings, and by adjusting the blade shape to satisfy the above condition, to achieve a relative Mach number distribution that satisfies the above condition, as a whole, an impeller Was able to effectively increase the efficiency. This efficiency is much better than similar impellers currently on the market.

Each condition is considered to contribute to an improvement in efficiency for the following reasons. That is, under the above conditions and conditions, a large kinetic energy can be given to the flow on the shroud side where the radius of rotation is large, so that the occurrence of turning stall can be delayed and the efficiency can be improved.

Under the above conditions, the absolute Mach number at the blade outlet can be reduced to 1 or less, and the generation of a shock wave can be reliably suppressed. As a result, a decrease in efficiency due to the generation of a shock wave can be prevented.

Under the above conditions, it is possible to prevent the occurrence of an impact stall or the like at the inlet portion, thereby contributing to an improvement in efficiency.

Under the above conditions, the load ratio of each part of the blade from the inlet to the outlet can be optimized, which can contribute to improvement in efficiency.

Under the above conditions, the load on the blade can be suppressed to a predetermined level or less, and separation of the flow in the impeller can be prevented. As a result, efficiency can be improved.

When only the conditions are set, there is a possibility that the outlet angle becomes smaller and the pressure rise becomes worse as compared with a general turbo compressor using refrigerant gas. On the other hand, since the relative Mach number distribution described above is achieved by a combination of the conditions described above, it is possible to increase the pressure by reducing the loss in the blade. Therefore, both the efficiency and the head can be increased.

<Example> Hereinafter, an example will be described in detail with reference to the accompanying drawings.

FIG. 1 is a sectional view of a main part showing an embodiment of a centrifugal impeller according to the present invention. In FIG. 1, the impeller has a plurality of blades 4 formed between a hub 2 fixed to a rotating shaft 1 and a shroud 3 arranged to face the hub 2. The rotating shaft 1 is driven to rotate via a speed increasing mechanism (not shown).

The shroud surface 3a on the inner surface side of the shroud 3 is:
On the suction side B (upstream side), there is provided a parallel portion 3b parallel to the rotation axis of the impeller. The parallel portion 3b rises along a gentle curve from the downstream end to the blade outlet O. I have.

Referring to FIG. 2 as well, the blade 4 is of a three-dimensional shape whose shape changes in the blade width direction, and the leading edge 4a, which is the suction-side edge, is the upstream end of the parallel portion 3b. Starting near the part
As C1, the blade length Ls extending linearly to the hub surface 2a side and along the shroud surface 3a and the blade length Lh along the hub surface 2a are: Ls / Lh ≧ 1, more preferably Ls /Lh=1.05 to 1.1. That is, the blade 4 has a blade length Ls along the shroud surface 3a,
It is set to be equal to or longer than the blade length Lh along 2a. As a result, the blades 4
The inclination angle α of the line 1 connecting C1 and the intersection A1 between the leading edge 4a and the hub surface 2a with respect to the rotation axis 1 becomes gentle, for example, about 25 to 30 °, and the leading edge 4a of the blade 4 The shroud surface 3a side protrudes greatly to the suction side B. The exit angle β of the blade 4 (see FIG. 3) is set to 30 ° to 40 °. Note that at least several tens of mm near the blade outlet is set to the same angle as the outlet angle β, but when the outer diameter is reduced or expanded according to the increase or decrease of the required head, the characteristics of the impeller are similar. It is preferable because it can be changed to. Further, the outlet width W of the blade 4 is appropriately adjusted according to the design air volume.

With the above configuration, the inclination angle α of the front edge 4a of the blade 4 is considerably gentle, and the shroud surface 3a side of the blade 4 projects greatly to the suction side B.
In the shroud 3 side region E near the front edge 4a, it is possible to suppress the occurrence of the rotating stall and the backflow of the fluid. Further, the blade length Ls along the shroud surface 3a is longer than before, and the diameter D2 of the starting point C1 is minimized, and the relative Mach number at the leading edge 4a of the blade 4 can be reduced. The flow in the impeller in the low flow rate region can be improved. Further, in addition to adjusting the shape of the blade 4 mainly on the shroud surface 3a side, by setting the blade outlet angle β to 30 ° to 40 °, the blade outlet O
The relative and absolute Mach numbers of the wings can be reduced to 1 or less, and the relative Mach numbers on the wing surface of the hub surface 2a, the central flow surface and the shroud surface 3a are reduced to 1 or less except for a part such as the leading edge 4a. Can be. Therefore, as a whole, the efficiency of the impeller can be effectively increased.

[Test I] The impeller according to the present invention (Test Example 1) and the conventional impeller (Comparative Example 1) were mounted on the same turbo compressor, and a comparison test of adiabatic efficiency in each air volume range was performed. . However,
As the working fluid, Freon 11 was used. In addition, FIG. 4 shows the blade shapes of the test example 1 and the comparative example 1 on the meridian plane. In this figure, the blade entrance diameter D2 of each of Test Example 1 and Comparative Example 1 is 216 mm, the blade exit diameter D1 is 410 mm, and the blade exit angle β is 30 ° in Test Example 1 and Comparative Example 1 is
45 ゜.

FIG. 5 shows the results of this comparative test. As is clear from the figure, the adiabatic efficiency of the test example 1 is higher than that of the comparative example 1 in each air volume range. In this figure, the maximum heat insulation efficiency of Test Example 1 is set to 100% (the same applies hereinafter).

[Test II] The intersection of the blade leading edge 4a and the shroud surface 3a in Test Example 1 was
The shroud surface 3a is set closer to the blade outlet O than the parallel portion 3b, that is, the shroud surface 3a side of the leading edge 4a is on the suction side B.
A test was performed under the same test conditions to compare the heat insulation efficiency of Test Example 1 with that of Test Example 1 (Comparative Example 2).

FIG. 6 shows the results of the comparative test. As is clear from the figure, the efficiency of the compressor of Comparative Example 2 is lower than that of Test Example 1 in each air volume range.

[Test III] Test Example 2 in which the blade outlet angle β of Test Example 1 was 40 °,
Comparative Example 3 having 45 ° was prepared, and a test for comparing the adiabatic efficiency in each air volume range of Test Example 1 with these was performed under the same test conditions. Further, the relative Mach number Mw from the leading edge 4a of the blade to the outlet O of the blade on the negative pressure surface and the pressure surface of the blade 4 was examined.

FIG. 7 shows the adiabatic efficiency in each air volume range. The relative Mach number Mw of Test Example 1 is shown in FIG. 8, the relative Mach number Mw of Test Example 2 is shown in FIG. 9, and the relative Mach number Mw of Comparative Example 3 is shown in FIG.
Each is shown in the figure. However, in each of the figures, (c) shows the relative Mach number Mw on the shroud surface 3a side of the blade 4, and (b) shows the relative Mach number Mw in the middle part between the shroud surface 3a and the hub surface 2a. (A) shows the relative Mach number Mw on the hub surface 2a side. Also, in the figure, the horizontal axis indicates the relative position with respect to the entire length of the blade, the left side is the blade front edge 4a side, and the right side is the blade outlet side.

As is clear from FIG. 7, Test Example 1 and Test Example 2
In each of the air volume ranges, the adiabatic efficiency is improved as compared with Comparative Example 3. Also, as is clear from FIGS. 8 and 9,
In Test Example 1 and Test Example 2, the relative Mach number Mw was 1 or less in any part of the blade 4, whereas in Comparative Example 3, the relative Mach number Mw was 1 on the shroud surface 3a side. Exists. Moreover, on the shroud surface 3a side, the relative Mach number Mw between the negative pressure surface and the pressure surface
Are reversed. Therefore, in order to increase the efficiency, the blade outlet angle β is set in the range of 30 ° to 40 °,
It is necessary to determine the blade shape so that the relative Mach number Mw between the negative pressure surface and the pressure surface does not reverse at any position.

8 and 9, the relative Mach number Mw at the blade inlet takes a value of 1 or less from the hub surface 2a side to the shroud surface 3a.
Increasing towards the side. As a result, it is considered that the occurrence of impact stall or the like at the inlet portion can be prevented, thereby contributing to an improvement in efficiency.

Also, the relative Mach number of the pressure surface and suction surface at the blade inlet
The value obtained by dividing the average value of Mw by the average value of the relative Mach numbers Mw of the pressure surface and the suction surface at an intermediate position between the blade inlet and the outlet is 1.6.
(In FIGS. 8 (a), (b), and (c), they are approximately 1.2, approximately 1.2, and approximately 1.3, respectively.
In the figures (a), (b), and (c), approximately 1.
3, approximately 1.3 and approximately 1.6. ). As a result, it is considered that the load ratio of each portion of the blade from the inlet to the outlet can be optimized, thereby contributing to an improvement in efficiency.

Further, the difference between the relative Mach number Mw of the pressure surface and the suction surface is smaller than the average value of the relative Mach number Mw of the pressure surface and the suction surface. This specifies that the load on the blades 4 is suppressed to a predetermined level or less, whereby it is considered that the separation of the flow in the impeller can be prevented, thereby contributing to an improvement in efficiency.

In addition, by setting the blade exit angle to 30 ° to 40 °,
The absolute Mach number at the blade exit can be reduced to 1 or less,
It is presumed that as a result of reliably suppressing the generation of shock waves, it was possible to prevent a decrease in efficiency due to the generation of shock waves. However, the exit angle of 30-40 °
Since it is smaller than the exit angle (45 ° to 70 °) of a general impeller applied to a turbo compressor using refrigerant gas, there is a possibility that the pressure rise may be deteriorated. On the other hand, by setting the relative Mach number distribution as described above, it is considered that the boosting can be achieved through the reduction of the loss in the blade 4, so that both the efficiency and the head can be increased. .

The leading edge 4a of the blade 4 may have a curved shape whose central portion is recessed downstream in addition to the straight shape shown in the figure. Further, the intersection (starting point C1) between the front edge 4a and the shroud surface 3a may be set further on the suction side B side than that shown in the figure.

<Effects of the Invention> As described above, according to the centrifugal impeller of the present invention, by achieving a predetermined relative Mach number distribution from the inlet to the outlet of the blade, the occurrence of turning stall or the like at the inlet is prevented. In addition, the load ratio of each part of the blade from the inlet to the outlet was optimized, and the load on the blade was suppressed to prevent flow separation. Thereby, efficiency can be improved.

Further, the outlet angle is set smaller than that of a general impeller applied to a turbo compressor using a refrigerant gas.
Thus, the absolute Mach number at the blade outlet can be reduced to 1 or less, and the generation of a shock wave can be reliably suppressed. As a result, a decrease in efficiency due to the generation of a shock wave can be prevented. On the other hand, although the pressure may be insufficient due to the small exit angle, the head can be improved by achieving a predetermined relative Mach number distribution from the entrance to the exit of the blade. So
There is a unique effect that both the efficiency and the head can be improved.

[Brief description of the drawings]

FIG. 1 is a sectional view of an essential part showing an embodiment of a centrifugal impeller of the present invention, and FIG. 2 is a schematic view showing a meridional shape of the impeller. FIG. 3 is a front view of the blade, FIG. 4 is a schematic diagram showing a blade according to Test Example 1 and a plane circular cascade of blades according to Comparative Example 1, and FIG. 5 is thermal insulation of Test Example 1 and Comparative Example 1. The graph which shows efficiency. FIG. 6 is a graph showing the heat insulation efficiency of Test Example 1 and Comparative Example 2. FIG. 7 is a graph showing the adiabatic efficiencies of Test Example 1, Test Example 2, and Comparative Example 3. FIGS. 8, 9, and 10 are Test Examples 1, 2, and 3, respectively.
It is a graph which shows the relative Mach number of a blade about the test example 2 and the comparative example 3, Among them, the figure (c) is a graph which shows the relative Mach number on the shroud surface side of a blade,
(B) is a graph showing a relative Mach number at an intermediate portion between a shroud surface and a hub surface, (a) is a graph showing a relative Mach number on a hub surface side, and FIG. 11 is a main part showing a conventional example. Sectional view. 1 ... rotating shaft, 2a ... hub surface, 3a ... shroud surface, 4 ... blade, 4a ... leading edge, β ... blade outlet angle, C1 ... starting point, B ... suction side, Ls ... Blade length along shroud surface, Lh… Blade length along hub surface.

Claims (1)

(57) [Claims] 1. A centrifugal compressor applied to a turbo compressor using refrigerant gas and having a plurality of three-dimensional blades (4) arranged between a shroud surface (3a) and a hub surface (2a). In the impeller, a parallel portion (3b) parallel to the rotation axis of the impeller is formed on the suction side (B) of the shroud surface (3a), and the shape of the impeller is set so as to satisfy the following conditions. The centrifugal impeller is characterized in that the relative Mach number is as follows: The tip of the shroud side of the blade (4) projects toward the inflow side,
And the blade length (Ls) along the shroud surface (3a) is Ls / Lh = 1.0 with respect to the blade length (Lh) along the hub surface (2a).
It has a relationship of 5 to 1.1. The leading edge (4a) of the blade (4) extends toward the hub surface (2a) starting from the parallel portion (3b), and the leading edge (4a) of the blade (4)
(1) of the line (l) connecting the intersection (C1) of the blade and the shroud surface (3a) and the intersection (A1) of the leading edge (4a) of the blade (4) and the hub surface (2a). Angle (α) to 25
30 ゜. The blade outlet angle (β) is 30 ° to 40 °. The relative Mach number (Mw) at the blade inlet increases from the hub side to the shroud side while taking a value of 1 or less. Relative Mach number of pressure surface and suction surface at blade inlet (Mw)
Divided by the average value of the relative Mach number (Mw) of the pressure surface and the suction surface at an intermediate position between the blade inlet and the outlet is 1.
6 or less. The difference between the relative Mach number (Mw) of the pressure surface and the suction surface is smaller than the average value of the relative Mach number (Mw) of the pressure surface and the suction surface.