JPH08232887A - Pump with variable guide blade - Google Patents

Abstract

PURPOSE: To operate fluid machinery in a wide flow range by setting the length of a diffuser blade to be substantially equal to or a little larger than the length obtained by dividing the circumferential length in the diffuser fitting radius by the number of blades of the diffuser. CONSTITUTION: The length L of a diffuser blade is set to be substantially equal to or a little larger than the length obtained by dividing the circumferential length in the diffuser blade fitting radius γv by the number of diffuser blades. Accordingly, when the diffuser blades are rotated in such a manner as to be arranged in the connecting direction of the circumference, the leading edge and the trailing edge of two adjacent diffuser blades overlap each other. The position in the radial direction of a fulcrum for varying the diffuser blade is within a range such that the impeller radius γ2 is from 1.08 times to 1.65 times. The length L1 from the leading edge of the diffuser blade to the fulcrum is set 20%-50% of the full length L of the diffuser blade, so that the force for moving the diffuser blade can be minimized against the torque generated by an outflow from the impeller.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to an inlet guide vane for a centrifugal or mixed flow type liquid pump, a blower for gas, a compressor, etc. (these are collectively referred to as "pump" in the present specification). And a pump with a diffuser vane.

[0002]

2. Description of the Related Art Conventionally, a diffuser is sometimes attached to these pumps as a device for efficiently converting kinetic energy flowing out from an impeller into static pressure. In that case, a vaneless diffuser or a vaned diffuser is used. In the case of a vaned diffuser, the flow passage formed between the vanes is configured as an enlarged flow passage to serve as a diffuser. There was almost everything.

However, recently, it has been reported that the performance is improved by widening the interval between the blades and using a diffuser blade whose blade length is shorter than the pitch obtained by dividing the circumferential length by the number of blades ("small chord ratio circle"). A blade cascade diffuser "(Mechanical Society Papers, Vol. 45, No. 396, Sho 54-8, etc.) has been made. However, in these reports, the vanes were fixed. "Experimental Resul
ts on a Rotatable LowSolidity Vaned Diffuser "(AS
ME paper 92-GT-19) is available.

When the conventional centrifugal and mixed flow pumps as described above are operated in a low flow rate range other than the design point, the impeller,
Flow separation etc. occurs in components such as diffuser,
Due to these causes, the rate of increase in pressure with respect to the flow rate that can be generated by the pump is reduced, and an unstable phenomenon (surging, etc.) with the piping system occurs, resulting in the inability to operate.

The instability phenomenon of this flow will be further considered below. The velocity vector of the flow flowing out from the impeller can be divided into a radial component and a circumferential component as shown in FIG. Assuming that there is no loss in the diffuser and the flow is incompressible, the product r ₂ Vθ ₂ of the radius r _{2 of the} diffuser inlet and the circumferential velocity component Vθ ₂ flows to the outlet according to the law of conservation of angular momentum. The circumferential velocity component Vθ ₃ of Vθ ₃ is expressed as Vθ ₃ = Vθ ₂ · r ₂ / r ₃ (1) and is decelerated by the radius ratio of the diffuser.

On the other hand, the area A _{2 of the} diffuser inlet is expressed as A ₂ = 2πb ₂ r ₂ (3) when the diffuser width is b _2, and similarly the area A _{3 of the} diffuser outlet is A ₃ = 2πb ₃ It becomes r ₃ (4).

Assuming that the diffuser is a vaneless parallel wall diffuser, the area ratio A ₂ / A ₃ is equal to the radius ratio r ₂ / r _3.
Becomes equal to Assuming that there is no loss in the diffuser and the flow is incompressible, from the law of mass conservation, the radial velocity component Vr _{3 at} the diffuser outlet is V _r3 = V _r2 · r ₂ / r ₃ (5) . Therefore, the radial velocity component is also decelerated by the radius ratio of the diffuser, the inlet flow angle α ₂ and the outlet flow angle α ₃ become equal, and the flow becomes an equiangular spiral.

When the flow rate of the pump is reduced, assuming that the fluid slip in the impeller is almost constant depending on the flow rate,
Although the velocity component in the circumferential direction hardly changes,
Since the velocity component in the radial direction decreases almost in proportion to the flow rate, the flow angle decreases. When the flow rate of the pump is reduced, the flow that had a velocity component in the radial direction at the diffuser inlet also slows down due to the expansion of the diffuser area, and the law of conservation of mass is maintained.As a result, the velocity component in the radial direction becomes small at the diffuser outlet. turn into.

Since there is a boundary layer on the wall surface of the diffuser and both the velocity and energy are smaller than the velocity of the main flow, even if the velocity component in the radial direction is positive at the flow rate considered at the above-mentioned average velocity, Separation occurs in the flow in the layer, a flow of negative velocity component is generated, and it develops into a large-scale backflow. It is becoming clear that this backflow region becomes a rotating stall that periodically fluctuates, which triggers the generation of large-scale system vibrations such as surging.

It has been impossible for a conventional pump having a fixed diffuser flow passage to prevent separation and backflow in the diffuser caused by a decrease in flow rate. In order to improve this, a patent in which the diffuser width is variable (US Pat. No. 4,378,194, US Pat. No. 3,426,96)
4, JP-A-58-594, and JP-A-58-12240.
No. 0) has been issued. Further, as a means for changing the angle of the diffuser blade, Japanese Patent Laid-Open No. 53-1133.
08, JP-A-54-119111, JP-A-54-1
33611, JP-A-55-123399, JP-A-5
5-125400, JP-A-57-56699, and JP-A-3-37397).

Further, as another method for preventing the flow from becoming unstable, a pipe for bypass (blow in the case of a blower or a compressor) is provided in addition to the pipe connected to the pump, and the pump is not connected. There is also adopted a method in which the valve of the bypass pipe is opened at a flow rate at which a stable phenomenon occurs so that the operating state of the pump does not change even if the flow rate on the device side is reduced.

[0012]

However, in the conventional method as described above, in the method of narrowing the diffuser width, the friction loss on the wall surface of the diffuser becomes large, and the diffuser performance is significantly deteriorated. The mechanism has the disadvantage that the usable flow rate range is limited. Further, in the conventional method in which the angle of the diffuser blades is made variable, since the length of the diffuser blades is long, the adjacent diffuser blades come into contact with each other at a certain finite angle, and the deadline flow rate cannot be controlled. .

Further, in the method of providing the bypass in the pipe,
In order to avoid the unstable phenomenon on the device side, the flow rate at the unstable phenomenon occurrence point of the pump must be determined in advance and the valve of the bypass pipe must be controlled to open at that flow rate. Therefore, with this method, unless the flow rate at which the instability phenomenon occurs in the pump can be accurately grasped, it is not possible to control the entire device accurately, and also the characteristics when the rotational speed of the pump is changed must be grasped accurately. For example, since the control of the entire device cannot be performed accurately, there is a drawback in that it cannot follow the operation when the flow rate of the pump is continuously changed.

Further, although the point where the unstable phenomenon occurs can be avoided by opening the valve of the bypass pipe, the pump itself is in a state of being operated under a high load, which causes the pump to perform unnecessary operation. Therefore, there were many problems from the viewpoint of energy saving. Further, there is a drawback that the cost of the bypass piping and valve becomes high.

The present invention has been made in view of the above circumstances, and an instability phenomenon that occurs when a fluid machine is operated in a flow rate range below a design point flow rate can be avoided, and the fluid machine can be operated in a wide flow rate range. An object is to provide a fluid machine with variable guide vanes.

[0016]

The present invention has been made to solve these problems, and the invention of claim 1 is such that the diffuser vanes are rotatably mounted around a fulcrum and the angle thereof is variably controlled. In the pump made possible, the length of the diffuser blade is set to be approximately equal to or slightly longer than the length of the circumference at the diffuser mounting radius divided by the number of diffuser blades, and the blade is set in the tangential direction of the circumference. When the blades are rotated so as to be aligned, the front edge and the rear edge of two adjacent blades are overlapped with each other. The configuration of the invention of claim 1 is based on the following consideration.

FIG. 2 is a schematic diagram showing the state of the impeller outlet of the fluid machine (compressor). The flow directions of the fluid flowing out from the impeller 2 are indicated by arrows a (design flow rate), b (small flow rate), and c (large flow rate). As is clear from this figure, at a large flow rate other than the design point, the flow collides with the negative pressure surface side of the blade 3a of the diffuser 3 and at the small flow rate, the flow collides with the pressure surface side of the blade 3a of the diffuser 3, and the flow is separated, As shown in 3 (a diagram showing the relationship between the dimensionless flow rate and the diffuser loss), the loss in the diffuser increases. As a result, the overall performance of the compressor is as shown in FIG. 4 (a diagram showing the relationship between the dimensionless flow rate and the dimensionless head coefficient), and a characteristic (unstable) that rises to the right on the small flow rate side from the design point appears. In addition to the above, surging occurs at a certain flow rate, pressure fluctuations in the pipes increase, and operation becomes impossible.

In order to avoid these phenomena, it is conceivable that the angle of the diffuser blade is made variable and this angle is controlled so as to match the direction of the flow flowing out from the impeller. The method is shown below. When the flow rate at the impeller outlet is Q ₂ , the impeller outer diameter is D ₂ , the impeller outlet width is b ₂ , and the blockage coefficient of the impeller outlet is B, the radial velocity component Cm _{2 at the} impeller outlet is Cm ₂ = It can be expressed as Q ₂ / (πD ₂ b ₂ B) (6). Assuming the fluid is incompressible, Q ₂ is equal to the inlet flow rate Q, so Cm ₂ = Q / (πD ₂ b ₂ B) (7). On the other hand, the circumferential velocity component Cu _{2 at the} outlet of the impeller is
The slip coefficient is σ, the impeller exit angle from the circumferential direction is β ₂ ,
When the impeller peripheral velocity is U ₂ , it can be expressed as Cu ₂ = σU ₂ −Cm ₂ cotβ ₂ (8).

Therefore, the flow angle at the outlet of the impeller, that is, the blade angle α of the diffuser is α = arctan (Cm ₂ / Cu ₂ ) = arctan (Q / (πσD ₂ U ₂ b ₂ B-Qcotβ ₂ )) (9 ). Here, K ₁ = (πD ₂ ) ² σb ₂ b ₂ B, K ₂ = cotβ ₂ (10), and N is the number of revolutions, α = arctan (Q / (K ₁ N−K ₂ Q) ) (11) On the other hand, in the case of a compressible fluid, the impeller outlet flow rate Q _2
Let P _{r be} the pressure ratio at the compressor inlet / outlet and κ be the specific heat ratio.

[Equation 1] Can be expressed as Therefore,

[Equation 2] Therefore, from the equations (8) and (13), the impeller outlet flow angle α is α = arctan (Cm ₂ / Cu ₂ ) (14)

(Equation 3) Becomes

As described above, the diffuser flow is detected by the detector for detecting the suction flow rate and the rotation speed in the fluid machine handling the incompressible fluid, and the suction flow rate, the rotation speed and the pressure ratio of the inlet and the outlet in the machine handling the compressible fluid. By calculating the angle and controlling the diffuser vanes to match the flow angle, flow separation can be suppressed and surging can be avoided. In this case, when the flow rate is 0, the angle of the diffuser blade must also be 0 degree. In order to enable such a state, the length L of the diffuser blade is
As shown in FIG. 5, a configuration of the present invention is created in which the circumferential length at the diffuser mounting radius r _v is equal to or slightly longer than the length obtained by dividing by the number of diffuser blades.

According to a second aspect of the invention, the radial position of the fulcrum for making the diffuser blade variable is 1.
It is characterized in that it is set to be in the range of 08 times to 1.65 times. The configuration of the invention of claim 2 is based on the following consideration. 5 to 9 show diffuser blades having such a shape from a diffuser blade angle of 0 degree corresponding to a compressor flow rate of 0 to 10 degrees, 20 degrees, 40 degrees and 60 degrees. This vane can change up to 90 degrees due to some influence. Therefore, the radius r _v to which the diffuser vane fulcrum is attached needs to prevent the rotating impeller from coming into contact with the diffuser vane when the diffuser vane angle is set to 90 degrees.

FIG. 10 is an explanatory view for obtaining such a condition, where L is the total length of the diffuser blade, L1 is the length to the front edge of the diffuser blade,
If the condition of claim 1 is set to this, L1 becomes

[Equation 4] Given in. Further, in FIG. 10, when the condition for not contacting the tip of the impeller when the diffuser blade reaches 90 degrees is obtained,

(Equation 5) Is obtained. The invention of claim 3 is characterized in that the length from the front edge of the diffuser blade to the fulcrum is set to be 20% to 50% of the total length of the diffuser blade.

The torque around the shaft required to drive the diffuser blade must be larger than the torque generated by the pressure difference between the pressure surface side and the suction surface side of the diffuser blade shown in FIG. When the pressures acting on the leading edge side and the trailing edge side of the diffuser blade are similar, the torque required for rotation can be minimized by positioning the drive shaft in the center of the blade. However, since the pressure on the leading edge side of the diffuser blade is actually higher than that on the trailing edge side, it is necessary to position the drive shaft at 20 to 50% of the diffuser blade.

By the way, depending on the situation and purpose of use,
In some cases, the minimum flow rate to be controlled in the fluid machine does not have to be close to zero. In such a case, the length of the diffuser blade is made shorter than in the case of claim 1, and a gap is formed when the opening is most narrowed, that is, when the blade angle is 0 degree with respect to the circumferential direction. You can The invention as set forth in claim 4 is made from such a point of view. In a pump in which a diffuser blade is rotatably mounted around a fulcrum and its angle can be variably controlled, the length of the diffuser blade is A pump with variable guide vanes, characterized in that the pump is determined based on a minimum flow rate to be controlled.

The invention according to claim 5 further embodies this, and the length of the diffuser blade is such that the area of the opening (A ₄ between adjacent diffuser blades when the diffuser blade angle is most narrowed). ) And the area (A ₅ ) of the opening of the adjacent diffuser vanes in the design flow rate, the pump with variable guide vanes according to claim 4. A ₄ can be approximated to the area of the opening formed between the leading edge and the trailing edge of the diffuser blade when the diffuser blade angle is set to 0 degree, for example. Further, A ₅ can be calculated as an area on the circumference of the radius to which the diffuser blade is attached, excluding only the thickness in the circumferential direction of the diffuser blade at the angle of the diffuser blade at that time.

[0026]

In the invention of claim 1 as described above, when the diffuser blade is rotated about the fulcrum and its angle is set to 0, the length of the blade is divided by the number of diffuser blades in the circumferential length at the diffuser mounting radius. Since the length is set to be substantially equal to or slightly longer than the specified length, the leading edge and the trailing edge of two adjacent blades overlap each other, and the flow does not flow out to the downstream of the diffuser. As a result, when the flow rate of the compressor is 0, the instability phenomenon does not occur if the blade angle is set to 0 degree, and thus the pump can be stably operated in a wide flow rate range. However, it is necessary to avoid operating the diffuser blades in a fully closed state for a long time because the temperature of the entire compressor rises.

Further, in the invention of claim 2, when the diffuser blade is rotated around the fulcrum of the blade, the impeller and the front edge of the diffuser do not contact each other even if the diffuser angle becomes 90 degrees. Further, in the invention of claim 3, when the diffuser blade is a flat plate, if the length from the front edge of the diffuser blade to the fulcrum of rotation is 20% to 50% of the entire diffuser blade, this diffuser blade is provided. It is possible to minimize the force that moves the diffuser vanes against the torque generated by the flow out of the impeller when the vanes are rotated about the fulcrum of the vanes.

Further, according to the invention described in claim 4, by shortening the diffuser blade, energy loss due to friction with the diffuser blade can be prevented, and vibration of the diffuser blade and noise caused thereby can be prevented. it can. Further, for the same reason, it becomes unnecessary to guarantee the rigidity of the diffuser excessively.

[0029]

Embodiments of the fluid machine with variable guide vanes according to the present invention will be described below with reference to the drawings.

FIG. 11 is a vertical sectional view showing a single-stage centrifugal compressor to which the present invention is applied. As shown in this figure,
The diffuser blade 3 a of the diffuser 3 downstream of the impeller 2 is connected to the actuator 5 via a plurality of gears 4. Further, on the suction side of the compressor, a detection device 6 that detects the suction flow rate, a detection device 8 that detects the number of revolutions, and a detection device 9 that detects the pressure ratio between the inlet and the outlet are installed. The actuator 5 is connected to the control device 7,
The blade angle of the diffuser blade 3a is variable.
Reference numeral 10 is an inlet guide vane, which can be opened and closed to adjust the flow rate of the fluid machine without changing the rotational speed. As described above, according to the present invention, even if the inlet guide vane 10 is provided, it can be handled without any problem. Further, when the motor is driven at a constant rotation speed, the rotation detection device 8 may be omitted.

The shape of the diffuser vanes of this compressor is
As shown in FIGS. 5-9, the length of the diffuser vanes is approximately equal to or slightly longer than the circumferential length at the diffuser vane mounting radius divided by the number of diffuser vanes. Therefore, when the blades are rotated so as to be aligned in the tangential direction of the circumference, the leading edge and the trailing edge of two adjacent blades overlap. Also, the radial position of the fulcrum for making the diffuser blade variable is in the range of 1.08 times to 1.65 times the impeller radius, and even if the blade angle becomes 90 degrees, Does not interfere with the impeller.

The length from the front edge of the diffuser blade to the fulcrum is 20% to 5% of the total length of the diffuser blade.
Since it is set to 0%, when the diffuser blade is rotated around the fulcrum of the blade, the force that moves the diffuser blade against the torque generated by the flow flowing out from the impeller can be minimized. The control device 7
Outputs a drive signal to the actuator 5 based on the input signals from the detection devices 6, 8 and 9 and the relational expression input in advance to control the angle of the diffuser blade. This relational expression is set as follows, for example. That is, when the fluid to be processed by the fluid machine is incompressible, α = arctan (Q / (K ₁ N−K ₂ Q)) (11) is set, and when the fluid is compressible,

(Equation 6) Is set. By controlling the blade angle in this way, the loss in the diffuser blade 3a can be reduced as shown by the broken line in FIG. As a result, the overall performance of the compressor can be improved as shown by the broken line in FIG. 4, and an unstable phenomenon can be avoided and stable characteristics can be obtained.

FIG. 12 shows a comparison of the overall performance of the conventional device with the fixed diffuser vanes and the performance of the device according to the present invention. It is understood that the performance of the device according to the present invention can be stably operated up to near the shutoff flow rate as compared with the conventional one.

When a control device capable of controlling the number of revolutions is provided, the diffuser blade 3a is controlled according to the diffuser blade angle determined by the suction flow rate of the equation (11) or (15), and the head does not satisfy the predetermined value. If
It is possible to operate while avoiding an unstable state by controlling the rotation speed. 13 to 20 show another embodiment of the present invention, in which diffuser blades have different sectional shapes.
13 to 16 show the case where the airfoil is used, and FIGS. 17 to 20 show the case where the arcuate blade is used. Also in these examples, the dimensions of the diffuser blades are set in the same manner as in the above-mentioned examples, and the same effects as those in the above-mentioned examples are obtained.

FIG. 21 is a diagram showing the dimensional relationship of diffuser blades of a fluid machine according to another embodiment of the present invention, in which the length of the diffuser blades is the maximum when the diffuser blade angle is narrowed. It is determined based on the ratio of the area (A ₄ ) of the opening between the adjacent diffuser blades and the area (A ₅ ) of the opening of the adjacent diffuser blade at the design flow rate. The details will be described below.

The circumferential length L is represented by L = 2πr _v, where r _v is the distance from the pump center to the position where the diffuser blades are attached (center of rotation axis). The flow angle at the impeller outlet at the design point is generally about 25 degrees with respect to the circumferential direction. If the diffuser vanes were optimally matched to the flow at the design flow rate, this would be approximately this angle.
Therefore, _assuming that the vertical thickness at the mounting radius of the diffuser blade is T _n , the thickness T _t when it is inclined by 25 degrees in the circumferential direction.
Becomes T _t = T _n / sin (25 °).

The opening area A ₅ between the diffuser vanes on the circumference of the diffuser blades mounting radius, the number of blades of the diffuser Z, the width of the pump (inner size) and W, A ₅
= A (L-Tt) W = ( 2πr v -T t Z) W. Here, when the flow rate Q _s that can be controlled by the diffuser blade is shown by the ratio S of the design point flow rate Q _d , S = Q _s / Q _d . Therefore, the minimum required diffuser blade length L
_{v is, L v = (2πr v -} (2πr v -ZT n / sin (25 °)) (Q _s
/ Q _d )) / Z.

The diffuser vane thickness is made dimensionless by the length of the diffuser vane, which is determined as 10% of the vane length, divided by the diffuser mounting radius by the number of vanes, and the flow rate ratio that requires control is plotted along the horizontal axis. This is as shown in FIG. In practice, the diffuser blade length ratio, that is, the diffuser blade length, should be set to be equal to or greater than the value shown in the figure with respect to the flow rate ratio that needs to be controlled.

[0039]

As described above, according to the first aspect of the present invention, the unstable phenomenon such as surging which occurs when the fluid machine is operated in the flow rate range below the design point flow rate is avoided,
The fluid machine can be operated in a wide flow rate range from the cutoff flow rate to the maximum flow rate. According to the invention of claim 2, when the diffuser blade is rotated around the fulcrum of the blade, even if the diffuser angle becomes 90 degrees, the impeller and the front edge of the diffuser do not come into contact with each other, and the control in a wide range is possible. It will be possible.
According to the invention of claim 3, when the diffuser blade is rotated around the fulcrum of the blade, it is possible to minimize the force for moving the diffuser blade against the torque generated by the flow flowing out from the impeller. Therefore, it is possible to provide a fluid machine having good control characteristics. According to the invention of claim 4 or 5, it is possible to prevent energy loss due to friction with the diffuser blade, to prevent vibration of the diffuser blade and noise resulting therefrom, and also to ensure excessive rigidity of the diffuser. Disappear.

[Brief description of drawings]

FIG. 1 is an explanatory diagram of a flow in a vaneless diffuser.

FIG. 2 is a schematic diagram showing a state of fluid at an outlet of an impeller.

FIG. 3 is a diagram showing a relationship between a dimensionless flow rate and a diffuser loss.

FIG. 4 is a diagram showing a relationship between a dimensionless flow rate and a dimensionless head coefficient.

FIG. 5 is a diagram showing a blade shape when a flat diffuser blade has an angle of 0 degree and a shape of an opening formed between the blades.

FIG. 6 is a diagram showing a blade shape when a flat diffuser blade has an angle of 10 degrees, and a shape of an opening formed between the blades.

FIG. 7 is a diagram showing a blade shape when a flat diffuser blade has an angle of 20 degrees, and a shape of an opening formed between the blades.

FIG. 8 is a diagram showing a blade shape when a flat diffuser blade has an angle of 40 degrees, and a shape of an opening formed between the blades.

FIG. 9 is a diagram showing a blade shape when a flat diffuser blade has an angle of 60 degrees, and a shape of an opening formed between the blades.

FIG. 10: When the diffuser blade angle is set to 0 degree,
It is explanatory drawing for obtaining the conditions which a rotating impeller and a diffuser blade do not contact.

FIG. 11 is a vertical sectional view showing a single-stage centrifugal compressor which is an example of a fluid machine with variable guide vanes according to the present invention.

FIG. 12 is a diagram showing a comparison between the overall performance of a conventional device with a fixed diffuser vane and the performance of a device according to the present invention.

FIG. 13 is a diagram showing a blade shape when the angle of the blade-shaped diffuser blade is 10 degrees, and a shape of an opening formed between the blades.

FIG. 14 is a diagram showing a blade shape when the angle of the blade-shaped diffuser blade is 20 degrees, and a shape of an opening formed between the blades.

FIG. 15 is a diagram showing a blade shape when the angle of the blade-shaped diffuser blade is 40 degrees, and a shape of an opening formed between the blades.

FIG. 16 is a diagram showing a blade shape when a blade-shaped diffuser blade has an angle of 60 degrees, and a shape of an opening formed between the blades.

FIG. 17 is a diagram showing a blade shape when an angle of the arc-shaped diffuser blade is 10 degrees, and a shape of an opening formed between the blades.

FIG. 18 is a diagram showing a blade shape when an angle of the arc-shaped diffuser blade is 20 degrees, and a shape of an opening formed between the blades.

FIG. 19 is a diagram showing a blade shape when the angle of the arc-shaped diffuser blade is 40 degrees, and a shape of an opening formed between the blades.

FIG. 20 is a diagram showing a blade shape when an angle of the arc blade diffuser blade is 60 degrees, and a shape of an opening formed between the blades.

FIG. 21 is a view showing a diffuser blade according to another embodiment of the present invention, and is a diagram showing a dimensional relationship when the angle is 25 degrees which corresponds approximately to the design flow rate.

FIG. 22 is a graph showing the relationship between the flow rate to be controlled and the diffuser length.

[Explanation of symbols]

 2 Impeller 3 Diffuser part 3a Diffuser blade 4 Gear 5 Actuator 6 Detection device 7 Control device

─────────────────────────────────────────────────── ───

[Procedure amendment]

[Submission date] January 22, 1996

[Procedure Amendment 1]

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[0012]

However, in the conventional method as described above, in the method of narrowing the diffuser width, the friction loss on the wall surface of the diffuser becomes large, and the diffuser performance is significantly deteriorated. The mechanism has the disadvantage that the usable flow rate range is limited. Further, in the conventional method in which the angles of the diffuser blades are variable, since the diffuser blades are long, adjacent diffuser blades come into contact with each other at a certain finite angle, and the deadline flow rate cannot be controlled. Di
The diffuser blade is divided into the diffuser blades in the front part.
There is U.S. Pat. No. 3,957,392 with variable displacement.
However, even with this method, it is impossible to control the deadline flow rate.
It was

[Procedure Amendment 2]

[Document name to be amended] Statement

[Correction target item name] 0019

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Therefore, the flow angle at the outlet of the impeller, that is, the blade angle α of the diffuser is α = arctan (Cm ₂ / CU ₂ ) = arctan (Q / (πσD ₂ U ₂ b ₂ B-Qcotβ ₂ )) (9 ). Here, K ₁ = (πD ₂ ) ² σ b ₂ B , K ₂ = cot β ₂ (10), and N is the number of revolutions, α = arctan (Q / (K ₁ N−K ₂ Q)) (11) On the other hand, in the case of a compressible fluid, the impeller outlet flow rate Q _2
Let P _{r be} the pressure ratio at the compressor inlet / outlet and κ be the specific heat ratio.

It can be expressed as Therefore,

Therefore, from the equations (8) and (13), the impeller outlet flow angle α is α = arctan (Cm ₂ / Cu ₂ ) (14)

[Equation 3]

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[Equation 4] When the intersection (x, y) of ▲ b ▼ is calculated The length of L ₁ is ─────────────────────────────────────────────────── ───

[Procedure amendment]

[Submission date] February 1, 1996

[Procedure Amendment 1]

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[Equation 4]

Claims (5)

[Claims] 1. A pump in which diffuser blades are rotatably mounted around a fulcrum and whose angle can be variably controlled, wherein the length of the diffuser blades is defined by the circumferential length at the diffuser blade mounting radius. The length is almost equal to or slightly longer than the length divided by, and when the blades are rotated so as to be aligned in the tangential direction of the circumference, the leading edge and the trailing edge of two adjacent blades overlap. A pump with variable guide vanes. 2. The radial position of the fulcrum for varying the diffuser blade is 1.08 times the impeller radius to 1.65.
The pump with variable guide vanes according to claim 1, wherein the pump has a double range. 3. The length from the front edge of the diffuser blade to the fulcrum is 20% to 50% of the total length of the diffuser blade.
%, The pump with variable guide vanes according to claim 1 or 2. 4. A pump in which a diffuser vane is rotatably mounted around a fulcrum and whose angle can be variably controlled, wherein the length of the diffuser vane is determined based on a minimum flow rate to be controlled by the pump. A pump with variable guide vanes, which is characterized by 5. The length of a diffuser vane is based on the ratio of the area of the opening between adjacent diffuser vanes when the diffuser vane angle is most narrowed, and the area of the opening of the adjacent diffuser vanes at the design flow rate. The pump with variable guide vanes according to claim 4, which is determined.