JP3095210B2 - Fluid machinery with variable guide vanes - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a fluid machine such as a centrifugal or diagonal fluid pump, a blower for gas, a compressor, and the like, and more particularly to a fluid machine having inlet guide vanes and diffuser vanes. In the present specification, the fluid machines are collectively referred to as pumps.

[0002]

2. Description of the Related Art Conventionally, a diffuser has been attached to these pumps as a device for efficiently converting kinetic energy flowing out of an impeller into static pressure. In that case, a bladeless diffuser or a bladed diffuser is used, but in the case of a bladed diffuser, the channel formed between the blades and the blades is configured to serve as a diffuser as an enlarged flow channel. There was almost nothing.

However, recently, it has been reported that the performance is improved by using a diffuser blade whose blade length is shorter than the pitch obtained by dividing the circumferential length by the number of blades by widening the space between the blades (see "Small chord ratio circular shape"). Cascade diffuser ”(Transactions of the Japan Society of Mechanical Engineers, Vol. 45, No. 396, Showa 54-8, etc.) However, in these reports, the blades were fixed. "Experimental Resul
ts on a Rotatable LowSolidity Vaned Diffuser '' ASME
paper 92-GT-19).

When the conventional centrifugal and mixed flow pumps described above are operated in a low flow rate region other than the design point, the impeller,
Separation of flow 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 or the like) with the piping system occurs, resulting in a drawback that operation becomes impossible.

The flow instability phenomenon will be further considered below. The velocity vector of the flow flowing out of 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 that the flow is incompressible, the product r ₂ Vθ ₂ of the diffuser inlet radius r ₂ and the circumferential velocity component Vθ ₂ flows to the outlet according to the law of conservation of angular momentum, so the diffuser outlet circumferential velocity component V [theta] ₃ of, expressed as _{Vθ 3 = Vθ 2 · (r} 2 / r 3), is decelerated by the radius ratio portion of the diffuser.

On the other hand, if the diffuser width is b ₂ , the area A _{2 of the} diffuser inlet is expressed as A ₂ = 2πb ₂ r _2, and similarly, the area A _{3 of the} diffuser outlet is A ₃ = 2πb ₃ r _3. .

If the diffuser is a vaneless parallel wall diffuser, the area ratio A ₂ / A ₃ is the radius ratio r ₂ / r _3.
Becomes equal to No loss in the diffuser, the flow is assumed to be incompressible, the law of conservation of mass, the radial velocity component Vr ₃ of the diffuser exit becomes _{V r3 = V r2 · (r} 2 / r 3). Accordingly, in that the radial velocity component is also decelerated by the radius ratio portion of the diffuser, the inlet flow angle alpha ₂ and outlet flow angle alpha ₃ are equal, the flow of equiangular spiral.

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

[0009] Since there is a boundary layer on the diffuser wall surface and the velocity and the energy are smaller than the main flow velocity, even if the radial velocity component is positive at the flow rate considered at the above average velocity, the boundary layer does not exist. Separation occurs in the flow in the bed, and a flow of a negative velocity component occurs, which develops into a large-scale reverse flow. It is becoming clear that this backflow region is a turning stall that periodically fluctuates, and that this triggers large-scale system-wide vibrations such as surging.

In a conventional pump having a fixed diffuser flow path, it was impossible to suppress separation and backflow in the diffuser caused by a decrease in flow rate. In order to improve this, patents for varying the diffuser width (US Pat. No. 4,378,194, US Pat. No. 3,426,964)
JP-A-58-594, JP-A-58-122400
Issues). Japanese Patent Application Laid-Open No. 53-11330 discloses a technique for changing the angle of a diffuser blade.
No. 8, JP-A-54-119111, JP-A-54-13
No. 3611, JP-A-55-123399, JP-A-55-123399
-125400, JP-A-57-56699 and JP-A-3-37397).

As another method for preventing the flow from becoming unstable, a bypass (in the case of an air blower or a compressor, air blow) piping is provided in addition to the piping connected to the pump, so that the pump is not damaged. A method of opening the valve of the bypass pipe at a flow rate at which the stabilizing phenomenon occurs, so that the operation state of the pump does not change even if the flow rate on the device side is reduced is adopted.

[0012]

However, in the conventional method as described above, in the method of reducing the width of the diffuser, the friction loss on the diffuser wall surface is increased, and the diffuser performance is greatly reduced. The mechanism has the disadvantage that the usable flow range is limited.

Further, in the conventional method of changing the diffuser blade angle, the purpose is to optimally control the flow near the design point. Therefore, it is not possible to control the flow rate at which surging occurs or to a flow rate lower than that. Further, a method of determining the diffuser blade angle has not been disclosed.

As a method for determining the diffuser blade angle, there is Japanese Patent Laid-Open Publication No. Hei 4-81598, but this method also merely shows a method for conceptually determining the diffuser blade angle for the flow near the design point. A specific method for determining the diffuser blade angle in the range up to has not been disclosed.

US Pat. No. 3,957, in which the diffuser blades are divided so that only the front diffuser blades are variable.
No. 392, however, it was not possible to control the flow rate to the cutoff even with this method.

In the method of providing a bypass in a pipe,
In order to avoid instability on the device side, the flow rate at the point where the instability occurs in the pump must be determined in advance, and when the flow rate is reached, control must be performed to open the valve of the bypass pipe. Therefore, in this method, unless the flow rate at which the instability of the pump occurs can not be accurately grasped, the control of the entire apparatus cannot be performed accurately, and further, the characteristics when the rotation speed of the pump is changed must be accurately grasped. For example, since the control of the entire apparatus cannot be accurately performed, there is a disadvantage that the operation cannot continuously follow the operation in which the flow rate of the pump is continuously changed.

[0017] Further, even though the point of occurrence of the unstable phenomenon can be avoided by opening the valve of the bypass pipe, the pump itself is operated under a high load, and the pump is operated unnecessarily. Therefore, there were many problems from the viewpoint of energy saving. Further, there is a disadvantage that the cost of the bypass piping and the valve increases.

SUMMARY OF THE INVENTION The present invention has been made in view of the above-mentioned circumstances, and it is possible to operate a fluid machine in a wide flow rate range by avoiding an unstable phenomenon that occurs when the fluid machine is operated in a flow rate area below a design point flow rate. An object of the present invention is to provide a fluid machine with variable guide vanes.

[0019]

In order to achieve the above-mentioned object, the present invention suppresses the fluctuation of the flow generated in the diffuser by the following three means. Operation up to the small flow rate range.

According to a first aspect of the present invention, in a fluid machine having a diffuser blade, a detecting device for detecting a suction flow rate of the fluid machine, and the detecting device based on the following relationship between the suction flow rate and the diffuser blade angle: And a control device for controlling the diffuser blade angle from the suction flow rate detected by α, arctan (Q / (K ₁ N−K ₂ Q)) (1) It is as follows. α: diffuser blade angle Q: suction flow rate K ₁ = (πD ₂ ) ² σb ₂ B: constant N: rotation speed K ₂ = cotβ ₂ : constant σ: slip coefficient β ₂ : blade exit angle from the circumferential direction D ₂ : impeller outer diameter b _2: the impeller outlet width B: blockage factor wherein when the rotational speed N of the mechanical changes, provided a detection device for detecting a rotational speed is <br/> et al, on the detection value Based on this, the blade angle may be controlled by the above function.

According to a second aspect of the present invention, in a fluid machine having a diffuser blade, a detecting device for detecting a suction flow rate of the fluid machine, a detecting device for detecting a pressure ratio between an inlet and an outlet, and a suction flow rate and a pressure. And a control device for controlling the diffuser blade angle from the suction flow rate and the pressure ratio detected by the detection device based on the following relationship between the ratio and the diffuser blade angle.

## EQU3 ## The symbols here are as follows. alpha: diffuser blades angle Q: intake flow rate ^{_{K 1 = (πD 2) 2}} σb 2 B: constant N: rotational speed K ₂ = cotβ _2: Constant P _r: compressor inlet outlet pressure ratio kappa: specific heat ratio sigma: Slip Coefficient β ₂ : impeller exit angle from the circumferential direction D ₂ : impeller outer diameter b ₂ : impeller exit width B: blockage coefficient Here, when the number of revolutions N of the machine changes , it further rotates. A detection device for detecting the number may be provided, and the blade angle may be controlled by the function based on the detection value. Further, control of the may be performed from the maximum flow rate of the fluid machine deadline flow.

The structure of the invention of claim 1 or 2 is based on the following considerations. FIG. 2 is a schematic diagram showing a state of an impeller outlet of a fluid machine (compressor). The flow directions of the fluid flowing out of 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 impinges on the suction surface side of the blade 3a of the diffuser 3 and at a small flow rate, the flow collides with the pressure surface side of the blade 3a of the diffuser 3, and the flow separates. As shown in (a diagram showing the relationship between the dimensionless flow rate and the diffuser loss), the loss at 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). In addition to this, surging occurs at a certain flow rate, pressure fluctuations in the piping 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 the angle is controlled so as to coincide with the direction of the flow flowing out of the impeller. The method is described below. Assuming that the flow rate at the impeller outlet is Q ₂ , the impeller outer diameter is D ₂ , the impeller exit width is b ₂ , and the blockage coefficient at the impeller exit portion is B, the radial velocity component Cm _{2 at the} impeller exit is Cm ₂ = Q ₂ / (πD ₂ b ₂ B) (3) Assuming that the fluid is incompressible, Cm ₂ = Q / (πD ₂ b ₂ B) (4) since Q ₂ is equal to the inlet flow rate Q.

When a fluid flows in the diffuser, the velocity near the wall surface in the actual flow is smaller than that in the main flow due to the boundary layer. Assuming that the velocity of the main flow is U and the velocity in the boundary layer is u, the flow rate that is insufficient because the velocity in the boundary layer is lower than the velocity of the main flow is

(Equation 4) It is represented by When the excluded thickness is δ ^* and a flow equal to the main flow velocity U flows through this thickness portion, the flow rate is Uδ
^* Is represented. Since both are equal, the excluded thickness of the boundary layer is

(Equation 5) It is expressed as (For example, “Fluid dynamics (2)” (Corona), “Internal flow dynamics” (Yokendo))

In general, the average velocity in the cross section of the flow path is calculated in consideration of the decrease in the flow path width due to the excluded thickness. In the case of a fluid machine, the flow out of the impeller flows in the flow path width direction. It is not uniform (for example, JSME Transactions 44, 38
No. 4, "Research on the relative velocity distribution and performance of centrifugal impellers" (FIG. 20), the region where the speed is lower than the speed of the mainstream is larger than the thickness due to the boundary layer. Therefore, the velocity in the cross section of the flow path is underestimated unless the flow path width is corrected by the distortion of the boundary layer and the velocity with respect to the geometric flow path width. Becomes larger. Therefore, in the present invention, the blockage coefficient is taken into consideration as the correction amount of the flow path width.

It is clear from the above literature that the blockage coefficient is not uniform with respect to the flow rate. Therefore, the flow angle at the exit of the impeller cannot be determined unless the blockage coefficient that changes with the flow rate is grasped to some extent. Therefore, a means to detect physical quantities such as pressure, temperature or vibration or sound is installed on the body or piping of a fluid machine, and the blade angle at the time of the least vibration is obtained while adjusting the flow rate and the diffuser blade angle. And the blockage coefficient was inversely calculated based on the equation of the present invention. According to this calculation method, if the above equation is correct, a certain physical meaningful relationship between the flow rate and the blockage coefficient can be obtained.

FIG. 5 shows the relationship between the blockage coefficient and the flow rate thus obtained. Here, in order to compare with the value of the above-mentioned literature, the data is arranged using (1-B), and the abscissa indicates the dimensionless flow with respect to the design flow. According to these results, it was clarified that the blockage coefficient according to the present invention had a tendency different from the result of the above-mentioned document and changed almost in proportion to the flow ratio.

Although the slope of this straight line varies depending on the type of the impeller, the tendency is considered to be the same. Therefore, if the slope of this straight line is determined for each model, the blockage coefficient is calculated from this straight line for a change in the flow rate. And the flow angle at the impeller outlet can be accurately determined from the suction flow rate.

The above blockage coefficient is a function of the suction flow rate Q.
You may make it give as. In addition,
To have a linear relationship with the suction flow rate Q
May be.

On the other hand, the circumferential velocity component Cu _{2 at the} impeller outlet
Is the slip coefficient σ and the impeller exit angle from the circumferential direction is β
_{2. If} the impeller peripheral speed is U ₂ , it can be expressed as Cu ₂ = σU ₂ −Cm ₂ cot β ₂ (5). Accordingly, the flow angle at the impeller outlet, that is, the diffuser blade angle α, is α = arctan (Cm ₂ / Cu ₂ ) = arctan (Q / (πσD ₂ U ₂ b ₂ B-Qcot β ₂ )) (6) . Here, K ₁ = (πD ₂ ) ² σb ₂ B, K ₂ = cot β ₂ (7), and if N is the number of rotations, α = arctan (Q / (K ₁ N−K ₂ Q)) ( 8) On the other hand, in the case of compressible fluid, the impeller outlet flow rate Q ₂
Is simple, assuming that the pressure ratio at the inlet and outlet of the compressor is P _r and the specific heat ratio is κ,

(Equation 6) Can be expressed as Therefore,

(Equation 7) From formulas (5) and (10), the impeller outlet flow angle, that is, the diffuser blade angle α is

(Equation 8) Becomes

In this manner, the flow rate of the suction and the rotation speed of a fluid machine that handles incompressible fluid, and the flow rate of the suction and the rotation speed and the pressure ratio between the inlet and the outlet of the machine that handles compressible fluid are detected by the diffuser flow. By calculating the angle and controlling the diffuser blades to match the flow angle, it is possible to suppress flow separation and avoid surging. Since the determination of the diffuser blade angle is calculated using the parameters of the fluid machine as parameters, the invention can be applied uniformly to fluid machines having different dimensions and shapes. Therefore, an appropriate diffuser blade angle corresponding to the flow rate can be input in advance to, for example, a machine control device without repeating the experiment for each different model.

Further, in the fluid machine provided with the de Ifuyuza blades, a detection device for detecting the suction flow rate of the fluid machine, the suction flow and, based on the relationship between the area between the suction flow rate and the diffuser vane previously obtained Alternatively, the area between the diffuser blades may be controlled .

The basic concept of the present invention is as follows. Let A be the area formed between the diffuser blades. The speed passing through this area is represented by K ₁ C, where C is the absolute speed of the exit of the impeller, and the deceleration rate in the section from the impeller to the diffuser blade is K ₁ , where K ₁ is the deceleration coefficient. Now C is the impeller outlet radial velocity component Cm _2, a circumferential velocity component and Cu _2,

(Equation 9) It is expressed as The flow rate Q passing between the blades
₂ is expressed as Q ₂ = K ₁ CA (13). The circumferential velocity component Cu ₂ is expressed as Cu ₂ = σU ₂ −Cm ₂ cotβ ₂ (14), similarly to the equation (5).

Therefore, Q ₂ is

(Equation 10) Becomes On the other hand, since Q ₂ is expressed as Q ₂ = πD ₂ b ₂ B · Cm ₂ (16) from the equation (3), the impeller outlet radial velocity component Cm ₂ is Cm ₂ = Q / πD ₂ b ₂ B (17) Therefore,

[Equation 11] Becomes

[0035] _{Here, K 2 = πD 2 b 2} B (19) K 3 = (K 2 σπD 2) 2 (20) K 4 = 2K 2 σπD 2 cotβ 2 (21) K 5 = 1 + cot 2 β 2 ( 22), if the fluid is incompressible, the suction flow rate is Q, and the number of revolutions is N, the area A is

(Equation 12) It can be expressed as. In the case of a compressible fluid, if the pressure ratio at the compressor inlet / outlet is Pr and the specific heat ratio is κ, the flow rate at the impeller outlet is simply

(Equation 13) Becomes

Using this equation, the area calculated from the experimental results of the actual compressor shown in FIG. 5 is compared with FIG.
Comparing the area between 2 to 24 diffuser blades,
In FIG. 11, the horizontal axis shows the flow rate ratio with the design point flow rate, and the vertical axis shows the area ratio with the area at the diffuser mounting radius position. Here, a solid line indicates a calculated value, and a circle indicates an actually measured area. Both that match well.

By controlling the diffuser blade angle in accordance with the above-described function determined by the suction flow rate rotation speed and the pressure ratio between the inlet and the outlet by the above-described means, the operation can be performed while avoiding an unstable state. Further, when the rotation speed can be controlled, when the head does not satisfy a predetermined value when the diffuser blade is controlled, the operation can be performed while controlling the rotation speed to avoid an unstable state.
Further, in a fluid machine equipped with a diffuser blade,
A detection device for detecting the suction flow rate of the fluid machine, and the suction flow rate
Volume and the suction flow rate determined beforehand and between the diffuser blades
Area between the diffuser blades based on the relationship with the area of the diffuser
And a control device for controlling the, them and by controlling the area between the diffuser vanes is determined by the suction flow
Thus , the vehicle can be driven while avoiding an unstable state.

[0038] In the fluid machine provided with the de Ifuyuza blades, a detection device for detecting the suction flow rate of the fluid machine, parallel control area between the diffuser vanes angle and diffuser vanes from suction flow rate detected by the detecting device may be in so that a controller for, if thus controlling <br/> blade angle and area simultaneously, can be operated while avoiding instability.

For a fluid machine equipped with diffuser blades,
Detection device for detecting the suction flow rate and rotation speed of the fluid machine.
And a detector for detecting the pressure ratio between the inlet and outlet, and this
The suction flow rate, rotation speed and
From the pressure ratio, the diffuser blade angle and diffuser blade
With a control device that controls the area between them in parallel
May be. In addition, the above control is performed based on the maximum flow rate of the fluid machine.
It may be performed up to the cutoff flow rate.

[0040]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS An embodiment of a fluid machine with variable guide vanes according to the present invention will be described below with reference to the drawings.

FIG. 6 is a longitudinal 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 an actuator 5 via a plurality of gears 4. On the suction side of the compressor, a detection device 6 for detecting a suction flow rate, a detection device 8 for detecting a rotation speed, and a detection device 9 for detecting a pressure ratio between an inlet and an outlet are provided. The actuator 5 is connected to the control device 7,
The blade angle of the diffuser blade 3a is variable.
Reference numeral 10 denotes an inlet guide blade, which can be opened and closed to adjust the flow rate of the fluid machine without changing the rotation speed. As described above, in the present invention, even if the inlet guide vanes 10 are provided, they can be handled without any problem. When the motor is driven at a constant speed, the rotation detecting device 8 may not be provided.

The controller 7 outputs a drive signal to the actuator 5 based on the input signals from the detectors 6, 8, and 9 and the relational expression input in advance, and controls the angle of the diffuser blade. This relational expression is set as follows based on the analysis described above. That is, if the fluid to be processed by the fluid machine is incompressible, α = arctan (Q / (K ₁ N−K ₂ Q)) (1) is set, and if the fluid is compressible,

[Equation 14] Is set. The symbols here are as follows. alpha: diffuser blades angle Q: intake flow rate ^{_{K 1 = (πD 2) 2}} σb 2 B: constant N: rotational speed K ₂ = cotβ _2: constant sigma: Slip factor beta _2: feathers from the circumferential exit angle D _2: Impeller outer diameter b ₂ : impeller outlet width B: blockage coefficient Here, the blockage coefficient B changes depending on the flow rate Q, and is given based on the relationship in FIG. By controlling the blade angle in this manner, the loss at 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 stable characteristics can be obtained by avoiding an unstable phenomenon.

In the case where a control device capable of controlling the number of revolutions is provided, the diffuser blade 3 is controlled in accordance with the diffuser blade angle determined by the suction flow rate of the equation (1) or (2).
If a is controlled and the head does not satisfy the predetermined value, the operation can be performed while controlling the rotation speed to avoid an unstable state.
FIG. 7 shows the difference between the blade angle of the diffuser blade which can be operated by changing the flow rate experimentally using the compressor shown in FIG. 6 to avoid surging, and the angle obtained by the equation (2) under the condition. The flow coefficient is shown on the horizontal axis.
From this equation, it can be seen that the equation set in the present invention is effective.

In the figure, a circle indicates that the ratio between the peripheral speed of the impeller and the sound speed at the compressor inlet (hereinafter referred to as "peripheral speed Mach number") is 0.87, and the angle of the inlet guide blade is 0. Degrees (full open), the results of the triangles with the peripheral speed Mach number of 0.87 and the angle of the inlet guide blades of 60 degrees, the results of the squares with the peripheral speed Mach number of 1.21, The results at the time when the angle of the entrance guide vane is 0 degree are shown. From these results, the expression of the present invention, regardless of the peripheral speed of the impeller, i.e., the rotation speed, and whether or not the flow is given to the inlet of the impeller by the inlet guide blade, the diffuser impeller according to the formula of the present invention. It can be seen that the angle can be determined. FIG. 8 shows (2) of the present invention.
The diffuser blade angle in the equation is plotted against the flow coefficient, and can be approximated by a quadratic curve.

FIG. 9 is a flowchart showing a processing procedure of the fluid machine with variable guide blades according to the present invention. As shown in FIG. 9, when controlling the rotation speed, the rotation speed is set in step 1. When the rotation speed control is not performed, the process proceeds to the next step. Next, in step 2, the suction flow rate and, if necessary, the pressure ratio of the inlet and outlet are measured,
In step 3, the diffuser blade angle is determined from the relationship shown in equation (1) or (2). Then, in step 4, the diffuser blade angle is changed. When the rotation speed control is performed, it is determined in step 5 whether or not the head satisfies a predetermined value.

FIG. 10 shows a comparison between the overall performance of the conventional apparatus having fixed diffuser blades and the performance of the apparatus according to the present invention. It can be seen that the performance of the apparatus according to the present invention can be operated more stably up to the vicinity of the cutoff flow rate than the conventional apparatus.

FIG. 11 is a diagram for explaining a control method according to another embodiment of the present invention. The target device is the same as that shown in FIG. 6, and a description thereof will be omitted. In this embodiment, the diffuser blade angle is controlled by the suction flow rate to change the area formed between the blades. The method of obtaining the graph of FIG. 11 is as described above. With such a control method, at a flow rate equal to or less than the design point flow rate, the speed in the radial direction does not become negative at the diffuser outlet, and the operation can be performed while avoiding an unstable state without causing excessive loss.

FIGS. 12 to 16 show the shape of the flat diffuser blade at various angles and the shape of the opening formed between the blades, and the size of the opening is indicated by a circle. Was. 17 to 20 show the case where an airfoil is used, and FIGS. 21 to 24 show the case where a circular arc blade is used. The area of the opening is related only to the thickness of the blade, and the flat plate, airfoil, and arc blade described above have almost the same tendency.
It does not depend on the shape of the wing.

In these embodiments, the length of the diffuser blade is substantially equal to or slightly longer than the length obtained by dividing the circumferential length of the diffuser blade mounting radius by the number of diffuser blades. When rotated so as to be aligned in the direction, the leading edge and the trailing edge of two adjacent blades overlap. Further, 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 radius of the impeller. The length from the leading edge to the fulcrum of the diffuser blade is set to 20% to 50% of the entire length of the diffuser blade.

FIG. 11 shows the diffuser inlet radius (r _V ).
12 shows the relationship between the area ratio 2πr _v b ₂ and the area ratio between the openings between the diffuser blades shown in FIGS. 12 to 24. It can be seen that the area ratio changes almost linearly with the flow rate. . This area ratio is related only to the thickness of the blade, and the flat plate, airfoil, and arc blade described above have almost the same tendency.
It does not depend on the wing shape. Based on the relationship between the area 2πr _v b ₂ at the diffuser inlet radius (r _v ) shown in FIG. 11 and the flow ratio between the area ratio of the opening between the diffuser blades and the design flow rate (Q _d ), the diffuser opening due to the flow rate Can be determined.

FIG. 24 shows the vector of the absolute velocity from the diffuser entrance (impeller exit) to the diffuser exit when the diffuser blade angle is at a certain angle, and the velocity component vectors in the radial and circumferential directions. . The broken line in the figure shows the state in the vaneless diffuser where the vanes are not attached to the diffuser.

At the blade inlet, the velocity vector in the radial direction is small because the flow rate is small, and if it flows to the diffuser outlet without the blade, the flow will be decelerated by the diffuser radius ratio by deceleration (dashed line in the figure). ).
In this figure, the flow is shown as the average velocity, so no backflow occurs. However, in the actual flow, the flow separates near the wall surface due to the boundary layer, and the backflow occurs.

When the flow lost from the impeller reaches the opening between the blades, the flow path is narrowed, so that the flow is accelerated according to the area ratio shown in FIG. 11 and the flow angle becomes larger than that at the inlet. The velocity of this part is expressed as a substantially vertical vector of the flow path shown by the solid line in the figure, and its magnitude is determined by the law of conservation of mass.

As is apparent from this figure, the velocity vector in the radial direction is several times larger than the inlet part due to the acceleration due to the decrease in the area of the flow path part, and the backflow is less likely to occur downstream.
As a result, the unstable flow in the diffuser due to the decrease in the flow rate can be eliminated.

Further, in the present invention, since the diffuser blade angle and the area can be changed at the same time, the backflow in the diffuser which is generated at a low flow rate can be more effectively suppressed, and the operation can be performed while avoiding surging. is there.

[0056]

As described above, according to the present invention,
An unstable phenomenon such as surging that occurs when the fluid machine is operated in a flow rate range equal to or lower than the design flow rate can be avoided, and the fluid machine can be operated in a wide flow rate range.

[Brief description of the drawings]

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

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

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 graph showing a relationship between a blockage coefficient and a flow rate.

FIG. 6 is a longitudinal sectional view showing a single-stage centrifugal compressor which is an example of a fluid machine with variable guide blades according to the present invention.

FIG. 7 shows a difference between a blade angle of a diffuser blade capable of avoiding surging and an angle obtained by equation (2) under the condition, in which the compressor shown in FIG. FIG.

FIG. 8 is a graph in which the diffuser blade angle of the formula (2) of the present invention is plotted against the flow rate.

FIG. 9 is a flowchart showing a processing procedure of the fluid machine with variable guide vanes of the present invention.

FIG. 10 is a diagram showing a relationship between a dimensionless flow rate, efficiency, and a head coefficient.

FIG. 11 shows a relationship between an area at a radius where the diffuser blade is attached and an area ratio between an opening between the diffuser blades and a suction flow ratio with respect to a suction flow rate at a design point of the fluid machine. .

FIG. 12 is a diagram showing a blade shape when the angle of a flat diffuser blade is 0 degrees, and a shape of an opening formed between the blades.

FIG. 13 is a diagram showing a blade shape when the angle of a flat 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 a flat 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 a flat 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 the angle of a flat diffuser blade is 60 degrees, and a shape of an opening formed between the blades.

FIG. 17 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. 18 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. 19 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. 20 is a diagram showing a blade shape when the angle of the blade-shaped diffuser blade is 60 degrees, and a shape of an opening formed between the blades.

FIG. 21 is a diagram showing the shape of a blade when the angle of an arc-shaped diffuser blade is 10 degrees, and the shape of an opening formed between the blades.

FIG. 22 is a diagram showing the shape of a blade when the angle of an arc-shaped diffuser blade is 20 degrees, and the shape of an opening formed between the blades.

FIG. 23 is a view 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. 24 is a diagram showing the shape of the blade when the angle of the arc-shaped diffuser blade is 60 degrees, and the shape of the opening formed between the blades.

FIG. 25 is a diagram showing a vector of an absolute velocity from a diffuser entrance (impeller exit) to a diffuser exit when the diffuser blade angle is at a certain angle, and velocity component vectors in a radial direction and a circumferential direction.

[Explanation of symbols]

 2 Impeller 3 Diffuser part 3a Diffuser blade 4 Gear 5 Actuator 6 Detecting device 7 Control device 8, 9 Detecting device 10 Inlet guide blade

──────────────────────────────────────────────────続 き Continued on the front page (58) Field surveyed (Int.Cl. ⁷ , DB name) F04D 27/02

Claims (2)

(57) [Claims] In a fluid machine provided with a diffuser blade, a detection device for detecting a suction flow rate of the fluid machine, and a suction flow rate detected by the detection device based on the following relationship between the suction flow rate and the diffuser blade angle: And a control device for controlling the diffuser blade angle from the above. [Number 1] _{α = arctan (Q / (K} 1 N-K 2 Q)) symbols in this case are as follows. α: diffuser blade angle Q: suction flow rate K ₁ = (πD ₂ ) ² σb ₂ B: constant N: rotation speed K ₂ = cotβ ₂ : constant σ: slip coefficient β ₂ : blade exit angle from the circumferential direction D ₂ : Impeller outer diameter b ₂ : Impeller exit width B: Blockage coefficient 2. A fluid machine equipped with a diffuser blade, wherein a detecting device for detecting a suction flow rate of the fluid machine, a detecting device for detecting a pressure ratio between an inlet and an outlet, and a detector for detecting a suction flow rate and a pressure ratio and a diffuser blade angle. A fluid machine with variable guide blades, comprising: a control device for controlling a diffuser blade angle from a suction flow rate and a pressure ratio detected by the detection device based on the following relationship. [Number 2] _{α = arctan ((1 / P} r) 1 / κ Q /
(K ₁ N− (1 / P _r ) ^{1 / κ} K ₂ Q) The symbols here are as follows. alpha: diffuser blades angle Q: intake flow rate ^{_{K 1 = (πD 2) 2}} σb 2 B: constant N: rotational speed K ₂ = cotβ _2: Constant P _r: compressor inlet outlet pressure ratio kappa: specific heat ratio sigma: Slip Coefficient β ₂ : blade exit angle from the circumferential direction D ₂ : impeller outer diameter b ₂ : impeller exit width B: blockage coefficient