JP3383023B2 - Centrifugal fluid machine - Google Patents

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a centrifugal fluid machine such as a pump and a compressor used for supplying water and blowing air, and more particularly to a fluid machine requiring low noise and low pressure pulsation in a wide operating range.

[0002]

2. Description of the Related Art Conventionally, for example, in a centrifugal fluid machine such as a boiler feed pump shown in FIG. 8, which is continuously operated for a long time, in order to improve operation efficiency, a dynamic pressure component of a flow coming out of an impeller 1 is used. In order to effectively convert the static pressure component into a static pressure component, the vaned diffuser 2 and the volute casing are used.
However, as shown in FIG. 9, a velocity distribution V2 and a pressure distribution are generated in the circumferential direction between the blades 11 at the impeller outlet. As a result, the flow coming out of the impeller periodically hits the diffuser blade 21 and the volume casing tongue, and the pressure of the frequency component of (the number of impeller blades Z × the number of impeller rotations N) and its harmonics. Pulsations and noise were generated, which had a negative effect on the reliability of the equipment and the operating environment.

As a method for reducing such fluid noise, there is known a technique for canceling pulsation by adjusting the length of a passage through which fluid flows, as disclosed in Japanese Patent Laid-Open No. 206206/1982. Further, as disclosed in Japanese Patent Laid-Open No. 58-135397, a diffuser blade inlet, a volume casing tongue portion, and an impeller blade outlet are three-dimensionally skewed to provide a passage width direction. There have been proposed techniques for mitigating interference between the flow coming out of the impeller and the flow of the impeller and the diffuser vane inlet and the volute casing tongue.

However, the first prior art described above has a drawback that it is effective only in the case of a special double volume casing and cannot be applied to a general diffuser or volute, and the second prior art has a flow passage width. When the width is narrow, the twist is small, so that the interference reducing effect is small. On the other hand, when the width of the flow path is wide, there are problems in manufacturing and strength, and it is often impractical.

By the way, such a velocity distribution and a pressure distribution in the circumferential direction between the impeller outlet blades are mixed and diffused to be uniform when the flow path length on the downstream side of the impeller is sufficiently long. Therefore, the impeller blade outlet and the diffuser blade inlet and volume
Even if the radial gap with the casing tongue is increased, pressure pulsation and noise are reduced. However, if this radial gap expansion is applied up to the radial gap between the impeller side plate and the diffuser side wall, the efficiency is reduced and the axial thrust in the small flow range where backflow occurs from the diffuser blade leading edge. Cause increase. Therefore, as shown in FIG. 10, conventionally, as shown in FIG. 10, the trailing edge 11 of the impeller blade, the diffuser blade inlet 21 and the volume.
Only the gap 41 with the casing tongue is widened, and the radial gap 42 between the impeller side plate 12 and the diffuser side wall 22 is provided.
Was made narrow.

However, if the length between the diffuser side wall and the diffuser blade inlet is increased in order to keep the radial gap 42 between the flow path side walls small, a new drawback occurs in the fluid machine of low specific speed. That is, FIG. 13 (P.159) of the Society of Mechanical Engineers, Volume B, Volume 57, No. 543 (November 1991).
As shown in FIG. 6, when the inflow angle is small in the vaneless diffuser, r / r2 becomes large (in FIG. 10 of this application, the length from the inner diameter end of the side wall 22 to the diffuser vane inlet 21). 1 becomes longer), 2 near the side wall of the flow path
In 1b, the flow stalls and a backflow 9 occurs. Therefore, the extension of the diffuser and the volume casing side wall to the inner diameter side is
There is a problem that causes an upward rising instability characteristic of the lift curve.

[0007]

SUMMARY OF THE INVENTION The present invention has been made based on the above-mentioned matters, and it is possible to reduce noise and pressure pulsation with the conventional manufacturability and strength without sacrificing the efficiency and the characteristics of the lift curve. The purpose is to do.

[0008]

In order to achieve the above object, according to the present invention, the radial position of the trailing edge of the blade of the impeller is made to substantially match the outer diameter of the side plate at the impeller flow path side plate portion, and A side plate in which the trailing edge of the blade of the roadside plate is formed on the outer diameter side of the trailing edge of the blade in the center of the flow path, and is on the outer diameter side of the trailing edge of the blade of the center of the flow path.
The height of the trailing edge of the blade in the width direction of the flow passage is set higher than the thickness of the trailing edge of the blade.
It is configured to be. Further, the blade shape of the side plate portion on the outer diameter side from the central portion of the trailing edge of the impeller blade is configured to be smaller than the blade angle defined from the circumferential direction at the central portion of the blade trailing edge. Further, in a fluid machine in which the blade inlet passage width of the diffuser is a passage narrower than the impeller outlet passage width, the passage width direction position of the impeller side plate portion blade trailing edge is
It is configured so as to be substantially on the extension line in the radial direction of the width of the flow path of the blade inlet of the diffuser or the like.

Still further, the radial position of the leading edge of the diffuser blade or the like is arranged on the inner diameter side of the flow passage side wall portion of the diffuser or the like from the center portion of the flow passage, and the flow passage width of the leading edge of the side wall blade is provided. The directional position is configured to be substantially on the extension line in the radial direction of the impeller exit width.

[0010]

In the bladeless portion, the average flow velocity is decelerated in the flow direction to restore the dynamic pressure to static pressure, but the velocity of the fluid near the wall surface becomes slow due to friction on the wall surface. Therefore, in the vaneless portion, the inflow angle α3 to the diffuser is as shown in FIG.
When the (angle formed with the circumferential direction of the flow) becomes smaller, the inertial force (ρ / 2) V ² m3 in the radial direction of the flow cannot overcome the pressure gradient ΔP in the radial direction and flows backward to the inner diameter side, and normal pressure recovery occurs. You won't get it. As a result, in a centrifugal fluid machine having a vaneless diffuser, the lift curve often has an upward-sloping unstable characteristic in a low flow rate region where the inflow angle is small. This tendency is more remarkable in a fluid machine having a low specific velocity with a small outflow angle from the impeller.

The above-mentioned backflow in the vaneless portion is determined by the balance between the pressure gradient in the radial direction of the flow and the boundary layer thickness of the side wall formed before reaching the diffuser vane, and therefore, it is necessary to prevent backflow. It is effective to increase the flow angle near the side wall of the bladeless portion. For this purpose, in order to keep the radial gap 42 between the flow path side walls small without increasing the length from the inner diameter end of the side wall of the vane to the diffuser vane inlet, only the impeller side plate is the outer diameter. May extend in the direction. However, in this case, the flow near the impeller outlet side plate is likely to decelerate in the radial direction and accelerate in the circumferential direction due to the effect of friction on the wall surface, rather reducing the outflow angle and preventing backflow in the bladeless part. Can not.

In the centrifugal fluid machine of the present invention, the radial position of the blade trailing edge of the impeller is made to substantially coincide with the outer diameter of the side plate of the impeller flow path side plate portion, and the blade trailing edge of the flow path side plate portion is flowed. Configured on the outer diameter side from the trailing edge of the central blade, and
Flow path width height of the trailing edge of the side plate blade that is on the outer diameter side of the edge
Is configured to be higher than the blade trailing edge thickness ,
Normally, the feathers are placed at the base of the feathers so that they do not cause strength problems.
Even if the radius is approximately the same as the root rear edge thickness t, the blade action can be given to the flow near the impeller flow path side plate portion. Therefore, the inflow angle to the diffuser becomes larger than that in the case where there is no blade trailing edge up to the outlet of the impeller side plate, and it becomes difficult for backflow to occur near the side wall of the bladeless portion. On the other hand, since the blade trailing edge in the central portion of the impeller flow path through which the main flow flows is separated from the diffuser blade leading edge, the speed in the circumferential direction between the impeller outlet blades,
Pressure distribution is less affected and pressure pulsation and noise are reduced.

Further, when only the blade of the side plate portion on the outer diameter side from the central portion of the trailing edge of the impeller blade is formed so as to be smaller than the blade angle defined from the circumferential direction at the central portion of the blade trailing edge, As shown in FIG. 5, the relative outflow angle β2 ′ can be reduced only near the impeller side plate. Therefore, the absolute outflow angle α2 in the vicinity of the impeller side plate is increased with almost no effect on the overall characteristics, so that the inflow angle is increased also in the vicinity of the side wall of the bladeless portion, and it is more difficult to backflow.

On the other hand, in a fluid machine such as a diffuser in which the blade inlet passage width is narrower than the impeller outlet passage width, the pressure pulsation is generated by utilizing the difference between the impeller and the diffuser passage width. It is possible to prevent the backflow near the side wall of the bladeless portion while reducing the noise. That is, if the position of the trailing edge of the blade of the impeller side plate portion in the flow passage width direction is configured to be substantially on the extension line in the radial direction of the width of the flow passage at the blade inlet of the diffuser, etc. The distance between the diameter and the leading edge of the blade of the diffuser, that is, the length of the flow path side wall portion of the diffuser becomes shorter. Therefore, the inflow angle can be more effectively increased in the vicinity of the side wall of the vane-free portion without affecting the diffuser or the like by the influence of the velocity and pressure distribution in the circumferential direction between the impeller outlet blades.

Further, in a fluid machine such as a diffuser having a blade inlet passage width wider than the impeller outlet passage width, a diffuser is used.
By utilizing the difference between the flow passage widths of the impeller and the diffuser on the side, it is possible to prevent backflow near the side wall of the bladeless portion while reducing pressure pulsation and noise. That is, the radial position of the leading edge of the diffuser vanes, etc. is configured on the inner diameter side of the central portion of the flow passage in the side wall portion of the flow passage of the diffuser, and the leading edge of the vane of the flow passage side wall portion of the diffuser etc. The position in the flow passage width direction is approximately on the extension line of the impeller outlet width in the radial direction. In this case, the flow near the side wall of the vaneless portion is caused by the diffuser vanes in the radial direction from the circumferential direction before being decelerated due to the influence of friction on the wall surface, that is, the flow angle becomes large and the backflow is unlikely to occur.

[0016]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS An embodiment of the present invention will be described below with reference to FIGS. FIG. 1 is a cross-sectional view of an impeller and its vicinity of an embodiment of a centrifugal diffuser pump of the present invention.

In FIG. 1, an impeller 1 is mounted on a rotary shaft 3 and driven by a motor (not shown) via the rotary shaft 3. Further, in order to recover the dynamic pressure of the flow from the impeller 1 to static pressure, a diffuser 2 with vanes is arranged on the outer diameter side of the impeller 1. The blade 11 of the impeller 1 is formed such that the trailing edge portion is on the flow path side plate portion 11b and is on the outer diameter side of the flow path central portion 11a.

The flow near the side wall of the vaneless portion is decelerated by the effect of friction on the wall surface. Due to this deceleration, the radial inertial force of the flow cannot overcome the pressure gradient in the radial direction and flows backward to the inner diameter side. Therefore, backflow can be prevented by increasing the flow angle near the side wall of the bladeless portion.

In the centrifugal diffuser pump of the present invention,
The blade 11 of the impeller 1 is located at the trailing edge radial position and is located at the flow path side plate portion 11
b is made to substantially match the outer diameter of the side plate 12, and the trailing edge of the blade 11 of the flow channel side plate portion 11b is arranged on the outer diameter side of the trailing edge of the blade 11 of the flow channel central portion 11a. Outer diameter
The width of the trailing edge of the blade of the side plate that is the
It is configured to be higher than the thickness . When only the side plate is extended in the outer diameter direction at the impeller outlet, the flow is likely to be decelerated in the radial direction and accelerated in the circumferential direction due to the effect of friction on the wall surface. However, in this embodiment, the blade action can be applied to the flow near the impeller outlet side plate. Therefore, as compared with the case where there is no blade trailing edge up to the outlet of the impeller side plate, the inflow angle to the diffuser can be increased and backflow near the side wall of the bladeless portion can be prevented. On the other hand, since the blade-to-blade radial gap 41 between the blade trailing edge 11a and the diffuser blade leading edge 21 at the center of the flow path where the main flow flows is sufficiently separated, the speed in the circumferential direction between the impeller outlet blades at the diffuser blade leading edge, Pressure distribution is less affected and pressure pulsation and noise are reduced. Further, since the radial gap 42 between the impeller side plate and the diffuser side wall is narrow as usual, even when operating in a small flow rate range where a reverse flow is generated from the diffuser front edge, the reverse flow causes the impeller side plate gap 4
It becomes difficult to enter the inner diameter side through 2 and the sudden change in axial thrust due to backflow does not occur.

A fluid force acts on the trailing edge of the impeller blade due to the interference between the flow of the impeller outlet and the leading edge of the diffuser blade. Therefore, in order to alleviate the stress concentration at the root of the blade, the blade is rounded to the same extent as the blade trailing edge thickness t so as not to cause a problem in strength. When the thickness b is equal to or less than the blade thickness t, the blade function is not exerted. However, if the height b in the width direction of the side plate blade trailing edge 11b, which is on the outer diameter side of the blade central portion, is made larger than the blade trailing edge thickness t as shown in FIG. 3b, a sufficient blade action can be exhibited.

FIG. 2 shows the meridional flow at the exit of the conventional impeller, and the flow has already been decelerated on the side plate side at the exit of the impeller due to the effect of wall friction. Therefore, the outflow angle of the flow in the vicinity of the side plate shown in FIG. 3b is smaller than that in the central part of the impeller flow passage shown in FIG. 3a. Where U2 is the outer peripheral velocity of the impeller, V2 is the absolute velocity of the flow, W2 is the relative velocity of the flow, α
2 is the absolute outflow angle of the flow, β2 is the relative outflow angle of the flow,
′ Indicates the flow near the side plate.

FIG. 4 is a sectional view of the blades of the impeller of another embodiment of the centrifugal diffuser pump of the present invention. In the present embodiment, considering the non-uniformity of the flow angle at the outlet of the impeller in the width direction of the flow path, the blade shape of the side plate portion 11b on the outer diameter side from the central portion of the trailing edge of the impeller blade is set to The blade trailing edge central portion 11a is configured to be smaller than the blade angle β2 defined in the circumferential direction. Therefore, as shown in FIG. 5, the relative outflow angle β2 ′ can be reduced only near the impeller side plate. As a result, the outflow angle α2 in the vicinity of the impeller side plate is increased with almost no effect on the overall characteristics, so that the flow in the vicinity of the bladeless part side wall is more difficult to flow backward.

FIG. 6 is a cross-sectional view of the vicinity of the impeller of another embodiment of the centrifugal diffuser pump of the present invention. In this embodiment, in consideration of the width of the central partition wall 13 of both suction impellers, the blade inlet passage width B3 of the vaned diffuser is narrower than the impeller outlet passage width B2. Therefore, by utilizing the difference in the flow passage width between the impeller and the diffuser, in this embodiment, the position b in the flow passage width direction of the trailing edge of the impeller side plate blade is set to the blade inlet portion flow passage width B3 of the diffuser. It is constructed almost on the extension line. Therefore, the flow angle in the vicinity of the vaneless diffuser side wall portion 21b can be reliably increased without affecting the velocity distribution and the pressure distribution in the circumferential direction between the impeller outlet blades.

In this embodiment, the case of a double suction type impeller has been described, but if the blade inlet passage width B3 of the vaned diffuser is narrower than the impeller outlet passage width B2, one-side suction is performed. It goes without saying that the impeller has the same effect.
Further, if the diffuser side wall shape is formed in a taper shape from the inlet end to the blade leading edge as shown in FIG. 6, the slow flow with the meridional flow velocity Vm2 near the impeller side plate portion is accelerated by the contraction. Therefore, the flow angle near the vaneless diffuser side wall portion 21b can be further increased.

FIG. 7 is a cross-sectional view of an impeller of another embodiment applied to a centrifugal type diffuser pump in which the flow passage width B3 of the vane inlet portion of the diffuser with vanes is wider than the flow passage width B2 of the impeller outlet. Is. Considering the difference in the flow passage width between the impeller and the diffuser and the non-uniformity of the flow angle at the impeller outlet shown in FIG. 3 in the flow passage width direction, the diffuser blade central portion 21a is considered.
The side wall blade leading edge 21b on the inner diameter side needs to have a certain height or more. In this embodiment, the width direction position of the side wall blade front edge 21b on the inner diameter side of the diffuser blade center is approximately on the extension line of the impeller outlet width B in the radial direction. Therefore, the flow angle α3 in the vicinity of the vaneless diffuser side wall portion 21b can be reliably increased without being affected by the velocity distribution and the pressure distribution in the circumferential direction between the impeller outlet blades.

Although the embodiments have been described in the case of being applied to a centrifugal type diffuser pump, the present invention is effective for a gas machine such as a compressor or a blower as long as it is a centrifugal type fluid machine having a diffuser with vanes. is there. Further, the same effect can be obtained even when a volume casing is used in place of the diffuser with vanes on the outer diameter side of the impeller, and particularly in the case of a multiple volute having a plurality of tongue portions. The effect is obtained.

[0027]

As described above, according to the present invention, the radial position of the trailing edge of the impeller of the impeller is substantially matched with the outer diameter of the side plate of the impeller flow path side plate portion, and The edge is configured on the outer diameter side of the blade trailing edge in the center of the flow path,
Width direction of the trailing edge of the side plate blade that is on the outer diameter side of the blade trailing edge
Since the height of the blade is higher than the thickness of the trailing edge of the blade, it is usually attached to the blade so that problems with strength do not occur.
The root has a radius equal to the trailing edge thickness t
Also , the blade action can be applied to the flow near the impeller outlet side plate, which is easily decelerated in the radial direction and accelerated in the circumferential direction due to the influence of friction on the side surface. Therefore, the inflow angle to the diffuser becomes larger than that in the case where there is no blade trailing edge up to the outlet of the impeller side plate, and pressure pulsation and noise can be reduced while preventing backflow near the side wall of the bladeless portion.

Furthermore, since the side plate blade shape on the outer diameter side of the impeller blade central portion is configured to be smaller than the blade angle defined from the circumferential direction at the blade central portion trailing edge,
The relative outflow angle can be reduced only in the vicinity of the impeller side plate. Therefore, the inflow angle to the diffuser is increased without affecting the overall performance, and pressure pulsation and noise can be reduced while preventing backflow near the side wall of the bladeless portion.

Further, since the blade tip of the element having the wider flow passage is extended to the opposing element side along the flow passage extension line of the opposing element, the velocity distribution in the circumferential direction between the impeller outlet blades is increased. The flow near the bladeless side wall can be caused to flow before it is decelerated by the effect of wall friction without affecting the pressure distribution, and pressure pulsation and noise can be prevented while preventing backflow near the bladeless side wall. Can be reduced.

[0030]

[Brief description of drawings]

FIG. 1 is a vertical sectional view of an impeller and its vicinity of an embodiment of a centrifugal diffuser pump of the present invention.

FIG. 2 is an explanatory diagram of the flow on the meridian surface of the centrifugal impeller outlet.

3A and 3B are explanatory diagrams of the relationship between the flow velocity components, FIG. 3A is an explanatory diagram of the relationship between the flow velocity components of the impeller blade I-I cross section of FIG. 2, and FIG. 3B is the impeller blade of FIG. Explanatory drawing of the relationship of the velocity components of the flow of II-II cross section.

FIG. 4 is a blade shape diagram of the impeller blade I-I cross section of FIG. 1;

5 is an explanatory diagram of a relationship between velocity components of an impeller outlet flow in FIG.

FIG. 6 is a vertical cross-sectional view of an impeller and its vicinity of another embodiment of the centrifugal diffuser pump of the present invention.

FIG. 7 is a vertical cross-sectional view of an impeller and its vicinity of another embodiment of the centrifugal diffuser pump of the present invention.

FIG. 8 is a vertical sectional view of a conventional centrifugal diffuser pump.

FIG. 9 is an explanatory diagram of the flow between the blades at the outlet of the centrifugal impeller.

10 is a vertical cross-sectional view of the main part near the impeller of FIG.

FIG. 11 is an explanatory diagram of a flow near the side wall of the vaneless diffuser.

[Explanation of symbols]

DESCRIPTION OF SYMBOLS 1 ... Impeller 11 ... Impeller vane trailing edge 11a ... Impeller vane trailing edge central part 11b ... Impeller vane trailing edge side plate part 12 ... Impeller side plate 13 ... Impeller central partition 2 ... Diffuser 21 ... Diffuser Blade leading edge 21a ... Diffuser blade leading edge central portion 21b ... Diffuser blade leading edge side wall 22 ... Diffuser sidewall 3 ... Rotating shaft 41 ... Radial gap 42 between impeller blade and diffuser blade. Radial gap between impeller side plate and diffuser side wall 5 Water return vane 6 Bearing 7 Shaft seal 8 Barrel casing 9 Backflow

─────────────────────────────────────────────────── ─── Continuation of the front page (72) Hiromi Kobayashi, 502 Jinrachicho, Tsuchiura-shi, Ibaraki Prefecture Machinery Research Institute, Hitachi, Ltd. In-plant (72) Inventor Shin Terashima 603 Kintate-cho, Tsuchiura-shi, Ibaraki Hitachi Ltd. Tsuchiura Plant (56) References JP-A-55-107099 (JP, A) (58) Fields investigated (Int.Cl . ⁷ , DB name) F04D 29/24

Claims (5)

(57) [Claims] 1. A centrifugal fluid machine having a diffuser with blades or a volute casing, wherein an impeller is mounted on a rotating shaft mounted in a casing, and the dynamic pressure of the flow flowing out from the impeller is restored to static pressure. , The blade trailing edge radial position of the impeller is made to substantially coincide with the outer diameter of the side plate at the impeller flow channel side plate portion, and the blade trailing edge of the flow channel side plate portion is located on the outer diameter side from the blade trailing edge of the flow channel central portion. After the blade in the center of the flow path
The height of the trailing edge of the side plate blade on the outer diameter side of the edge in the width direction of the flow path
A centrifugal fluid machine characterized in that the thickness is higher than the trailing edge thickness of the blade . 2. An impeller on a rotating shaft mounted inside a casing.
Mounted to reduce the dynamic pressure of the flow out of the impeller to static pressure.
Bladed diffuser or volute case for power recovery
In a centrifugal fluid machine having a blade
Set the root rear edge radius position to the outer diameter of the side plate at the impeller flow path side plate part.
Make sure that the blades on the flow path side plate are in the center of the flow path,
The outer diameter side of the blades at the rear edge of the
Set the blade shape of the side plate that is on the outer diameter side of the edge to the rear edge of the blade center.
In a centrifugal fluid machine characterized by being configured smaller Kunar so than the blade angle defined from a circumferential direction. 3. The blade of claim 1, wherein the blade shape of the side plate on the outer peripheral side of the blade trailing edge in the central portion of the flow passage is smaller than the blade angle defined from the circumferential direction at the blade central portion trailing edge. A centrifugal fluid machine characterized by the above. 4. A diffuser with blades or a volute casing, wherein an impeller is mounted on a rotating shaft mounted in a casing, and the dynamic pressure of the flow flowing out from the impeller is restored to static pressure. in the fluid machine channel width of the tongue portion of the blade inlet or volute casing of the diffuser is narrow flow passage from the impeller outlet channel width, the blade trailing edge radial position before Symbol impeller in the impeller flow path side plate portion It is arranged on the outer diameter side from the central part of the flow path, and the position of the trailing edge of the side plate blade in the flow path width direction is substantially on the extension line in the radial direction of the blade inlet flow path width of the diffuser or the like. Centrifugal fluid machine. 5. A centrifugal fluid machine having a vaned diffuser or a volute casing, wherein an impeller is mounted on a rotating shaft mounted in a casing, and the dynamic pressure of the flow flowing out from the impeller is restored to static pressure. The vane of the diffuser or the tongue of the volute casing is configured such that the leading edge radial position is on the inner diameter side of the flow passage side wall of the diffuser or volute casing from the center of the flow passage. A centrifugal fluid machine characterized in that its position is arranged on a line extending substantially in the radial direction of the impeller outlet width.