WO1999036701A1

WO1999036701A1 - Centrifugal turbomachinery

Info

Publication number: WO1999036701A1
Application number: PCT/JP1999/000077
Authority: WO
Inventors: Hideomi Harada; Shin Konomi
Original assignee: Ebara Corporation
Priority date: 1998-01-14
Filing date: 1999-01-13
Publication date: 1999-07-22

Abstract

A centrifugal turbomachinery capable of preventing an excessive increase in the manufacturing cost, adapted to effectively reduce a secondary flow in a vane wheel flow passage and reduce a loss, which is ascribed to the secondary flow, to a minimum level, and having a high efficiency, comprising a vane wheel provided with a plurality of vanes (3) between a central inlet (6a) and an outer circumferential outlet (6b), and a flow passage formed between these vanes and used to send a fluid from the inlet to the outlet by the rotation of the vane wheel, the portions of the vanes (3) which are on the side of a hub (2) having a circumferential inclination preceding in the rotational direction of the vane wheel with respect to the portions of the vanes which are on the side of a shroud (4), the angle of inclination of the vanes which is defined as an angle of the vanes formed in a plane viewed from the outlet of the flow passage with respect to a plane perpendicular to a hub surface having a tendency to decrease from the inlet toward the outlet, whereby the center lines of a shroud-side portion and a hub-side portion viewed from a front part of a vane inlet cross each other in a position where a ratio of dimensionless radius to a radius of the outlet is 0.8-0.95.

Description

Description Centrifugal turbomachinery Technical field

The present invention relates to an improvement of an impeller of a machine generally called a "turbo machine" such as a centrifugal liquid pump for pumping a liquid and a blower and a compressor for pumping a gas. Background art

FIGS. 9A to 10B show a typical turbomachine, in which a casing (not shown) having required piping is provided with a hub 2, a shroud 4, and an impeller 6 including a plurality of blades 3 therebetween. And a rotary shaft 1 connected to a rotary drive source. In such an impeller, the upper end 3a of the blade 3 is covered with a shroud surface 4a, and a space surrounded by two adjacent blades 3, a hub surface 2a and a shroud surface 4a flows. The road is constructed.

As a result, the impeller 6 rotates at the rotational angular velocity ω about the rotation axis 1, so that the fluid flowing into the flow path from the impeller inlet 6 a via the suction pipe and the like to the impeller outlet 6 b Transported to the outside of the machine via a discharge pipe. The surface facing the rotating direction of the blade 3 is the pressure surface 3b. The opposite surface is the suction surface 3c.

In these figures, as an example of the blade, the three-dimensional shape of the closed impeller is schematically shown with most of the shroud surface broken. In the case of open impellers, a separate part for forming the shoulder surface 4 Although there is no material, the casing outside the figure surrounding the impeller 6 mechanically also serves as the shroud surface 4, and there is no difference in the basic structure of the fluid dynamics from the closed impeller. The following description proceeds with an example of a closed impeller.

In the flow in the impeller flow path of such a centrifugal turbomachine, the low-energy fluid in the boundary layer on the wall moves due to the pressure gradient in the flow path, in addition to the main flow flowing substantially along the flow path. A secondary flow (a flow having a velocity component orthogonal to the main flow) generated due to this is generated. This secondary flow has a complicated effect on the main flow, and forms vortices and non-uniform velocity in the flow channel, which are not only in the impeller but also in the downstream part (diffuser, guide vane, etc.) Was causing significant losses in The total loss caused by this secondary flow is called secondary flow loss. In addition, the low-energy fluid in the boundary layer that is accumulated in a specific area of the flow channel due to the action of the secondary flow induces large-scale flow separation, causing a rising characteristic to the right, resulting in a stable turbomachine. It is known to cause inconvenience such as hindering driving. The secondary flow in the impeller is roughly classified into a secondary flow between blades generated along the shroud surface or hub surface, and a secondary flow in the meridian plane generated along the pressure surface or suction surface of the blade. It is known that secondary flow between blades can be suppressed by curving the blade shape backward. The other secondary flow on the meridian plane requires detailed optimization of the three-dimensional shape of the flow channel and cannot be easily weakened or canceled.

The mechanism of occurrence of the secondary flow on the meridian plane is explained as follows. As shown in FIG. 9 B, with respect to the relative flow in the blade channel, and the centrifugal force W ² / R by the curvature of the streamlines against the mainstream, co Rio Rica 2 omega W theta by rotation of the impeller The relative pressure field (reduce ds tat ic pres sure) p * (= p-0.5 pu ² ) is determined. Here, W is the relative velocity of the flow, R is the radius of curvature of the streamline, ω is the rotational angular velocity of the impeller, and \ Υ Θ is the velocity component of W in the circumferential direction with respect to the rotation axis 1. And p is the static pressure, p is the density of the fluid, and u is the peripheral velocity at a predetermined radial position from the rotation axis 1.

The distribution of the relative pressure field p * is high on the hub side and low on the shroud side so as to balance the centrifugal force W ² ZR toward the hub side and the corerica 2 ω W 図 in FIG. 9B. Distribution. In the boundary layer along the blade surface, the relative velocity W decreases due to the effect of the wall surface, so the centrifugal force W ² / R and the corioliser 2 w W 6 acting on the fluid inside the boundary layer are small. Therefore, it cannot be balanced with the mainstream pressure field p * described above. As a result, the low-energy fluid in the boundary layer moves toward the region with a small relative pressure p *, and on the pressure surface 3 b or the suction surface 3 c of the blade 3, along the blade surface from the hub side to the shroud side. Meridian secondary flow toward. These are indicated in FIG. 9A by dashed arrows on pressure surface 3b of blade 3 and solid arrows on suction surface 3c.

The secondary flow on the meridional surface can occur on both the suction surface 3c and the pressure surface 3b of the blade 3, but since the boundary layer on the suction surface 3c is generally thicker, the secondary flow on the suction surface 3c It is known that the generation of the next flow greatly affects the performance characteristics of turbomachines.

In this way, when the low-energy fluid in the boundary layer moves from the hub side to the shroud side, the center part between the blades conversely moves from the shroud side to the hub side so as to compensate for the flow caused by the movement. There is a flow towards. As a result, as schematically shown in FIG. 10A, a pair of vortices having different swirling directions called secondary vortices are formed in the flow path between the blades. This vortex causes the low-energy fluid in the flow path to accumulate at a specific location in the impeller (region where the relative pressure P * is low). Mixing causes large losses.

In addition, when a non-uniform flow created by insufficient mixing of a fluid with a low relative velocity and low energy and a fluid with a high relative velocity and high energy is released into the flow path downstream of the blade, Can cause significant losses when mixed. Such non-uniform impeller outlet flow makes the velocity triangle at the diffuser inlet inappropriate, causing mismatching with downstream vaned diff users and backflow in vaneless diffusers. This may cause the performance of the entire turbomachine to be significantly reduced.

Therefore, as shown in Fig. 11A and Fig. 11B, in order to optimize the distribution of the relative pressure P * inside the impeller, the blade position (blade inlet) with dimensionless meridional plane length m = 0 and Between the dimensionless meridian length m == 1.0 and the blade position (blade exit), the circumferential direction is such that the hub side of the blade precedes the rotation direction of the impeller with respect to the shroud side of the blade. Defined as the angle formed by the center line of the blade cross section of the impeller with respect to the plane perpendicular to the hub surface on the cross section of the flow path of the impeller as the dimension of the dimensionless meridional plane m increases, forming the blade inclination It is conceivable to configure the blade inclination angle to show a decreasing tendency.

In the impeller configured as described above, a circumferential blade inclination is formed so as to precede the rotation direction of the impeller, so that a force having a component directed toward the shroud surface 4 acts on the fluid, and The relative pressure field in the road generates a high relative pressure P * on the shroud surface side and a low relative pressure p * on the hub surface 2 side in order to balance against the force component going to the shroud surface. Will occur. Also, as the blade inclination angle is configured to exhibit a small inclination as the dimensionless meridian length m increases, the blade inclination is simply shifted in the circumferential direction by the blade on the side of the circle. As compared with the case, the effect of the inclination of the blade can be made higher. However, in the conventional technology having such a configuration, as shown in FIG. 11A, when viewed from the blade exit direction, a line connecting the shroud side and the hub center of the blade is perpendicular to the hub surface. When the impeller is rotated, the blade deforms so that it rises, causing a large bending stress at the root of the blade. Occurs.

Also, as shown in the figure, the line connecting the center of the blade on the hub side and the center of the blade on the hub side at the blade tip and the line connecting the center of the blade on the hub side and the center of the impeller at the blade impeller entrance. (Lean angle δ) Due to the rotation, the blade rises at the entrance due to the rotation, and a large bending stress is generated at the root of the blade. Furthermore, in the case of a closed impeller in which a cover is attached to the shroud side of the impeller, complex stress is generated in each part of the blade depending on the lean angle and the rake angle.

The root of such a blade is a welded structure when the impeller is welded, and it is easy to cause welding problems if the blade is inclined, and if the blade is insufficient, it is rotated. This may cause cracks in that area, leading to destruction. Also, high stresses in this area affect the service life of the impeller, and therefore place high demands on welding techniques and materials, which in turn increases production costs. Also, in the case of manufacturing by machine cutting, complicated machining must be performed, resulting in an increase in manufacturing cost. Disclosure of the invention

In view of the above-mentioned problems, the present invention effectively reduces the secondary flow in the impeller flow path without causing an excessive increase in manufacturing cost, thereby reducing the loss. The objective is to provide an efficient centrifugal turbomachine by minimizing wastewater.

According to the present invention, a plurality of blades are provided between an inlet on the center side and an outlet on the outer peripheral side, and a flow path for sending a fluid from the inlet to the outlet is formed between the blades by rotation of the impeller. A wheel having a circumferential blade inclination such that a hub side precedes a rotating direction of the impeller with respect to a blade side of the blade, and viewed from an outlet side of the flow path. Surface, the blade inclination angle, defined as the angle the blade makes with respect to the plane perpendicular to the hub surface, has a decreasing tendency from the inlet to the outlet, and therefore, when viewed from the front of the blade inlet. An impeller characterized in that the center lines of the blades on the shroud side and the hub side intersect in a dimensionless radius position indicated by a ratio with respect to the exit radius of the impeller in a range of 0.8 to 0.95. is there.

Further, an impeller rotatably accommodated in the casing is provided, and a plurality of blades are provided between an inlet on the center side and an outlet on the outer peripheral side of the impeller, and the impeller rotates between these blades. In the centrifugal turbomachine, a flow path for sending fluid from the inlet to the outlet is formed by the blade, the blade has a circumferential blade inclination such that the hub side precedes the shroud side of the blade in the rotation direction of the impeller. And, on the surface viewed from the outlet side of the flow passage, the blade inclination angle defined as the angle formed by the blade with respect to the plane perpendicular to the hub surface shows a decreasing tendency from the inlet to the outlet. Thus, the center line of the blades on the shroud side and the hub side when viewed from the front of the blade entrance is 0.8 to 0.95 at the dimensionless radius position indicated by the ratio to the exit radius of the impeller. Intersecting in a range It is a centrifugal turbo machine. BRIEF DESCRIPTION OF THE FIGURES 1A and 1B are diagrams schematically showing the shapes of the blades of the turbomachine according to the embodiment of the present invention. FIG. 1A is a meridional view, and FIG. 1B is a front view. 2A and 2B are diagrams schematically showing the shapes of the blades of a turbo machine according to another embodiment of the present invention. FIG. 2A is a meridional view, and FIG. 2B is a front view.

3A and 3B are diagrams schematically showing the shapes of the blades of a turbomachine according to another embodiment of the present invention. FIG. 3A is a meridional view, and FIG. 3B is a front view.

4A and 4B are diagrams schematically showing the shapes of the blades of a turbomachine according to another embodiment of the present invention, wherein FIG. 4A is a meridional view and FIG. 4B is a front view.

Fig. 5 is a diagram showing the relationship between the lean angle δ at the tip of the blade at the entrance of the closed impeller and the stress generated at the root of the blade on the exit side of the impeller. Fig. 6 shows the relationship between the rake angle y of the closed impeller and the stress at the root of the blade at the entrance of the impeller.

7A and 7B are diagrams showing the shape of an impeller as a simulation model for advancing the analysis. FIG. 7A is a meridional view, and FIG. 7B is a front view.

FIG. 8 is a graph showing a result of a test in which an impeller having a shape according to the present invention was attached to a stage of a compressor.

9A and 9B are views showing the shape of an impeller of a conventional centrifugal turbomachine, FIG. 9A is a perspective view, and FIG. 9B is a meridional view.

1OA and 1OB are diagrams showing the shape of the blades of the impeller of the conventional centrifugal turbomachine, FIG. 1OA is a cross-sectional view, and FIG. 10B is a front view.

Figures 11A and 11B also show other impellers of a conventional centrifugal turbomachine. It is a figure which shows the shape of a blade | wing, FIG. 11A is sectional drawing, FIG. 11B is a front view.

FIGS. 12A and 12B are views showing the shape of the blades of another impeller of the conventional centrifugal turbomachine. FIG. 12A is a sectional view, and FIG. 12B is a front view. It is. BEST MODE FOR CARRYING OUT THE INVENTION

Embodiments of the impeller having such a shape are shown in FIGS. 1A to 4B. In these figures, FIGS. 1A and 1B show a specific speed of 500, FIGS. 2A and 2B show a specific speed of 400, and FIGS. 3A and 3B show a specific speed of 350. 4A and FIG. 4B show a specific speed of 250. These impellers are designed based on the following ideas.

The inventor of the present invention has a circumferential blade inclination such that the hub side of the blade precedes the rotating direction of the impeller with respect to the blade side of the blade as shown in FIGS. 11A and 11B, As the dimensionless meridian length m increases, the blade inclination angle, defined as the angle formed by the center line of the blade cross section with respect to the plane perpendicular to the hub surface, decreases on the impeller flow path cross section. Based on the impeller with a tendency to show a shape, we simulated it by changing several parameters with the aim of suppressing excessive inclination. As a guideline for such a maximum value of the inclination angle, 110% of the stress acting when there is no inclination was considered appropriate.

Figure 5 shows the angle (lean angle) between the line connecting the center of the blade on the shroud side and the hub, and the line connecting the center of the blade on the hub side and the center of the impeller at the tip of the blade at the entrance of the closed impeller. δ) is plotted on the horizontal axis, and the stress generated at the root of the blade on the exit side of the impeller is calculated. It is shown on the basis of the time. From this figure, it can be seen that the stress increases as the lean angle increases. In this figure, if the allowable stress of the slat is defined as 110% of the stress when the lean angle is 0 degree, the limit angle at that time is 25 degrees.

Fig. 6 shows the angle between the line connecting the shroud side of the closed impeller and the center of the blade on the hub side and the plane perpendicular to the hub surface (rake angle γ) on the horizontal axis. The ordinate shows the stress at the root of the blade, and it can be seen that the stress increases as the rake angle increases. In this figure, if the allowable stress of the blade is defined as 110% of the stress at a rake angle of 0 degree, the limit angle at that time is 20 degrees.

Thus, when the rake angle and the lean angle of the blade are determined, the general shape of the blade is determined. Figures 7 7 and 7Β show the shape of the impeller as a simulation model for further analysis. Fig. 7A shows a meridional view and Fig. 7B shows a front view. For simplicity in the front view, a straight line is connected between the inlet and the outlet on the hub side and the chassis side of the impeller. Actually, the shape of the blade is slightly different from this shape because it is composed of curves.

As is clear from this figure, in the impeller where the hub has a shape that precedes the shaft side in the rotation direction at the blade outlet, the line connecting the inlet and outlet blade center lines on the hub side and the shaft side. Meet at one place.

From the above description, it is presumed that the position of this intersection is closer to the entrance from the exit of the impeller if the angle is large. The inventors have

δ <25, y <20

As a prerequisite, several impellers with different specific velocities were created, and the shape and dimensions of those with high efficiency were measured and analyzed. 1A to 4B show a front view and a meridional view of impellers developed by the inventors, each having a different specific speed. As is clear from this figure, in the front view of the impeller, the center lines of the blades on the shroud side and the hub side intersect near the exit of the impeller, and the intersection is defined by the ratio to the exit radius of the impeller. It can be seen that the dimensional radius position is in the range of 0.8 to 0.95. Fig. 8 shows an example of the result of a test in which the impeller according to the present invention was attached to a stage of a compressor, and it was found that the performance was significantly better than that of a stage having a conventional shape of an impeller. I will.

As described above, according to the present invention, the secondary flow in the impeller flow path is effectively reduced without causing an excessive increase in the manufacturing cost, and the loss due to the secondary flow is minimized. An efficient centrifugal turbomachine can be provided. Industrial applicability

INDUSTRIAL APPLICABILITY The present invention has industrial application by being used for an impeller of a machine generally called a "turbo machine" such as a centrifugal liquid pump and a blower and a compressor for pumping gas.

Claims

The scope of the claims

1. An impeller in which a plurality of blades are provided between the center-side inlet and the outer-peripheral outlet, and a flow path is formed between these blades for sending fluid from the inlet to the outlet by rotation of the impeller. And

The blade has a circumferential blade inclination such that the hub side precedes the shroud side of the blade in the rotation direction of the impeller,

And, on the surface viewed from the outlet side of the flow passage, the blade inclination angle defined as the angle formed by the blade with respect to the plane perpendicular to the hub surface shows a decreasing tendency from the inlet to the outlet, and Therefore, the center line of the blades on the shroud side and the hub side when viewed from the front of the blade inlet is in the range of 0.8 to 0.95 at the dimensionless radius position indicated by the ratio of the exit radius of the impeller. An impeller characterized by crossing.

2. It has an impeller rotatably housed in a casing, and a plurality of impellers are provided between an inlet on the center side of the impeller and an outlet on the outer peripheral side, and the impeller is disposed between these impellers. In a centrifugal turbomachine having a flow path for sending a fluid from an inlet to an outlet by rotation,

And, on the surface viewed from the outlet side of the flow passage, the blade inclination angle defined as the angle formed by the blade with respect to the plane perpendicular to the hub surface shows a decreasing tendency from the inlet to the outlet, and Therefore, when viewed from the front of the blade entrance, the center line of the blades on the shaft side and the hub side is in a dimensionless radius position shown by a ratio of the exit radius of the impeller in a range of 0.8 to 0.95. A centrifugal turbomachine characterized by crossing.