JPH09112481A - Discharge characteristic control type centrifugal pump - Google Patents

Abstract

PROBLEM TO BE SOLVED: To promote the widenization and energy-saving of a pressure control range, in this discharge characteristic control type centrifugal pump having an open type impeller. SOLUTION: An open type impeller 5 and a blade plate 6 being opposed to the former are made to be free of relatively approaching or separating in the axial direction, and an insertional groove, where each blade 5b is insertable in the axial direction, is formed in this blade plate 6, inserting the blade 5b into this insertional groove, and thereby the axial effective width d3 of the blade 5b is made adjustable by a relative axial movement between the blade 5b and blade plate 6. Therefore, pressure control can be done up to a low pressure area, and thus a control range is yet more widened and simultaneously the promotion of energy-saving is thus achievable.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a discharge characteristic control type centrifugal pump having, in a pump casing, an impeller having one axially open side and a blade plate axially opposed to the impeller.

[0002]

2. Description of the Related Art In this type of centrifugal pump having an open impeller and an impeller plate, a gap is conventionally formed between the impeller and the impeller plate, and one of the members is moved in the axial direction. By doing so, the size (width in the axial direction) of the gap is increased or decreased to control the discharge characteristics. As a specific example, Japanese Patent Application No. 4-254 capable of constant discharge pressure control.
No. 525 and the like.

FIG. 14 is a performance characteristic diagram of a conventional gap control type centrifugal pump. A curve T1 is a pump head curve when the gap is minimized, and is a maximum pump head curve. By gradually increasing the gap from this state, the amount of water that escapes to the gap in front of the impeller increases, and the pump head curve decreases, as in T2 and T3. As is clear from the figure, the manner of the decrease is such that the movement is substantially parallel to the lower left from T1, and thereby the closing lift (when the water volume becomes zero).
Has also been gradually decreasing.

[0004]

In the clearance control type centrifugal pump having the above-mentioned performance characteristics, when the discharge pressure constant control is performed, for example, the discharge pressure is detected and amplified or reduced by a pressure responsive valve. The gap is controlled by moving the blade plate or the impeller with thrust, and is controlled as shown by the line segment A-E in FIG.

However, in the case of performing the discharge amount constant control, as described above, when the gap is enlarged and the head curve is lowered, if the size of the gap becomes larger than a certain value, the gap is once added to the front gap. Since the leaked water recirculates into the impeller to generate a head, low pressure control below a certain limit (for example, below T3 in FIG. 14) becomes practically impossible. That is, T3 becomes the minimum pump performance. For example, when the discharge amount is controlled at the discharge amount Q3 as the discharge amount constant control, only the range from the maximum head HA to the minimum head HB can be controlled, and the control from the head HB to the head 0 cannot be performed. Become.

[0006] Further, as an irregular discharge pressure constant control, in applications where the suction pressure changes, the discharge pressure = total head +
If the change in the suction pressure may reach the width equivalent to the total head of the pump due to the suction pressure, the pressure control is not achieved unless the total head of the pump becomes zero theoretically.

For example, when the suction pressure is ± 0 m, the total pump head is 20 m, and the discharge pressure is 20 m (0 + 20),
If the suction pressure rises to 20 m, the pump discharge pressure rises to 20 + 20 = 40 m. In order to keep this at a constant discharge pressure of 20 m, the suction pressure is increased to 20 m as described above.
m, the pump head needs to be 0 m. In order to cope with such an application, it is necessary to change the pump performance up to a pressure of 0, but it is difficult to achieve it in the gap control of the performance as shown in FIG.

In the above-described system in which the pressure is controlled by the pressure responsive valve, the applicant of the present invention uses an electric pilot valve capable of controlling the set pressure in accordance with the discharge terminal pressure as the pressure responsive valve, thereby achieving a constant estimated terminal pressure control. Has developed and filed an application (Japanese Patent Application No. 7-23).
No. 1064). In such a device, the performance characteristics are set by the electric pilot valve so as to have a gradient as close as possible to the pipeline resistance curve R in FIG. 14, and along the straight line AB in the range of the water amount Q4 or more, labor saving is achieved. .

However, even in the estimated terminal pressure control using the electric pilot valve, the control can be performed only by the pump performance T3 when the water amount is equal to or less than Q3, and the control in the low pressure range cannot be performed.

As another conventional example, there is a pump having a structure in which discharge pressure is controlled by controlling the rotation speed of a pump driving motor using an inverter or the like. In addition to this, the efficiency is reduced due to an increase in heat loss in the motor accompanying a decrease in the rotation speed.

[0011]

SUMMARY OF THE INVENTION It is an object of the present invention to provide a discharge pressure characteristic control type centrifugal pump capable of controlling a discharge pressure over a wide range from a high pressure range to a low pressure range of a head, particularly a range close to zero in a low pressure range. I have.

[0012]

In order to solve the above problems, in the invention according to claim 1, one of an impeller whose one side in the axial direction is open and a vane plate which is arranged to face the open side of the impeller. , One of which is fixed in the axial direction and the other of which is movable in the axial direction, the blade plate is formed with an insertion groove into which each blade can be inserted in the axial direction, and each blade is inserted into the insertion groove, and the blade and the blade plate The effective width of the blade in the axial direction can be adjusted by moving the blade in the axial direction relative to.

According to the second aspect of the present invention, in order to further improve the control accuracy, in the discharge characteristic control type centrifugal pump according to the first aspect, the insertion groove is formed so as to have a bottom in the axial direction and is formed in the radial direction. The inner and outer ends are closed, and the groove depth in the axial direction is equal to or greater than the entire width of the blade in the axial direction.

[0014]

BEST MODE FOR CARRYING OUT THE INVENTION

(First Embodiment) FIG. 1 shows a single-stage centrifugal pump to which the present invention is applied. A main shaft 1 is rotatably supported by a casing side wall member 3 via a bearing seal 2 and a sleeve. The rear end portion in the axial direction protrudes into the spiral casing 4, and the casing 1 of the main shaft 1
An open impeller 5 and a blade plate 6 opposed thereto are mounted. The front end of the main shaft 1 is rotatably supported by a bearing case 10 and the like, and is interlockingly connected to a prime mover such as an electric motor (not shown).

The casing 4 is placed on a base or the like, and the casing side wall member 3 is connected to the bearing case 1 via a bracket 7.
Tied to 0.

A suction pipe 14 projecting downward in a curved shape is connected to the axial rear end of the casing 4, and an upward projecting discharge port 13 is connected to the radial end of the spiral casing 4. The discharge port 13 is connected to a discharge water supply conduit 18 via a discharge short pipe or the like, and a control pressure outlet 19 is provided on the side surface.

The impeller 5 has a side plate 5a and the side plate 5a.
And a plurality of centrifugal blades 5b that are integrally formed on the rear surface and are open at the rear surface.
While being fixed to the spindle 1 in the rotation direction by a key,
It is also fixed in the axial direction via a shaft sleeve 8 by a tightening nut 9 screwed to the rear end of the main shaft. A liner ring portion 5c is integrally formed on the front surface of the impeller 5, and is fitted to the liner ring 12 of the casing side wall portion 3.

The blade plate 6 is formed in a thick disk shape having a constant axial thickness, and a boss 21 is integrally formed at a radially inner end portion with a plurality of radial ribs 20 interposed therebetween. The boss 21 is fitted to the outer periphery of the shaft sleeve 8 so as to be movable in the axial direction, and the nut 9 restricts the maximum rearward movement position. A disk 11 having a hole in the center is fixed to the rear surface of the blade plate 6, and a cylindrical guide 22 projecting rearward is integrally formed on the rear surface of the disk 11. By fitting the cylindrical guide 22 to the liner ring 23 of the casing 4 movably in the axial direction,
At the rear of the disk 11, a control pressure chamber 25 for moving the blade plate 6 is formed.

FIG. 2 is a rear view of the impeller 5, in which, for example, six curved blades 5b are arranged.
As shown in FIG. 3, the blade plate 6 is formed with insertion grooves 26 having the same arrangement as the blades 5b. Each insertion groove 2
6 has an inner surface shape substantially corresponding to the cross-sectional shape of the blade 5b, but is formed to be slightly wider than the thickness t1 of the blade 5b, and has a slight margin c1, at the inner and outer ends in the radial direction.
It is formed in a length having c2. In addition, insertion groove 2
Both the inner and outer ends of 6 in the radial direction are closed.

FIG. 4 is an enlarged longitudinal sectional view of the impeller 5 and the blade plate 6. The insertion groove 26 penetrates the blade plate 6 in the axial direction, but its rear surface is closed by the disk 11. . As shown in the figure, the relative axial positional relationship between the impeller 5 and the blade plate 6 is at least the rear end of the blade 5b and the insertion groove 26 even when the blade plate 6 is lowered to the maximum rearward movement position. The engagement state of is maintained. The relative relationship between the axial full width d1 of the blade 6 and the axial depth d2 of the insertion groove 26 is such that the axial effective width d3 of the blade 5b can be adjusted to be as small as 0 at least. The d2 is the wing 5
The size is set to be equal to or greater than the total axial width d1 of b.

In the control pressure chamber 25 at the rear of the blade plate,
A secondary pressure outlet 28a of a pilot valve (pressure-responsive valve) 28 shown in FIG. 1 is connected, and a primary pressure inlet 28b of the pilot valve 28 is connected to a control pressure outlet 19 of the discharge port 13. .

The pilot valve 28 is a valve that can take out a secondary pressure that increases in proportion to an increase in the primary pressure.
When the discharge pressure in the discharge port 13 increases from the set pressure, the pressure in the control pressure chamber 25 increases in proportion thereto,
The forward thrust on the blade plate 6 is increased.

FIG. 5 shows an embodiment of the pilot valve 28, which has a primary pressure chamber 31 and a secondary pressure chamber 32, a coil spring 33 for setting pressure, and a diaphragm 34.
And a valve body 35 that adjusts the opening amount of the valve port 37 by connecting the primary pressure to the primary side set pressure set by the coil spring 33. The diaphragm 34 is pushed up to widen the space between the valve port 37 and the valve body 35, thereby increasing the secondary pressure.

The adjustment of the set load of the coil spring 33 is as follows.
By rotating the adjusting bolt 36 and adjusting the height of the vertically slidable spring receiving nut 45 screwed to the adjusting bolt 36, the adjustment can be performed. That is, when the adjustment bolt 36 in FIG. 5 is rotated in the reverse rotation direction (counterclockwise),
As the spring load increases, the valve body 35 descends and the valve opening 37
To reduce the secondary pressure. Conversely, when the motor rotates in the forward rotation direction (clockwise), the spring load decreases, and the valve element 35 returns upward to increase the secondary pressure.

In order to change the pressure set by the spring 33 so as to follow a predetermined pressure control characteristic by changing the current flow rate and the current pressure, as shown in FIG.
8, an electric pilot valve unit 55 including an operation part thereof, a pressure control circuit device 56, and a flow meter 54.
It has.

The electric pilot valve unit 55 has a small control motor 59 having a small-diameter drive gear 58 that meshes with the large-diameter gear 50 of the electric pilot valve 28, and a pressure control circuit controller that detects the discharge pressure of the discharge port 13. And a pressure transmitter 60 that transmits the pressure to the electronic unit 63 that functions as a pressure sensor. The pressure control circuit device 56 includes a constant voltage unit 61, a relay circuit 62, the electronic unit 63, and a preamplifier 64 functioning as an arithmetic circuit.

The constant voltage unit 61 supplies the pressure transmitter 60 with a constant voltage for operating it. The flowmeter 54 is arranged in the middle of the discharge water supply pipe line 18.

Various pressure control characteristics (gradients) are stored in the preamplifier 64 and can be set arbitrarily. When a load signal (flow signal) of the amount of water used is input from the flow meter 54, the preamplifier 64 is set. A target pressure change command corresponding to the pressure control characteristic is input to the electronic unit 63.

The electronic unit 63 receives the target pressure change command from the preamplifier 64 and the current pressure signal from the pressure transmitter 60, detects a deviation between the current pressure and the target pressure,
The control pressure of the electric pilot valve 28 is adjusted by rotating the control motor 59 so as to reach the target pressure.

An example of various pressure control characteristics that can be set by the preamplifier 64 is shown in FIG. 7. In FIG. 7, a characteristic line AB having a gradient closest to the pipeline resistance curve R of the apparatus, and a gradient having a gentler gradient than that. Characteristic line AC or straight line A-
D can be freely set to an arbitrary gradient, and the discharge pressure can be set to be constant regardless of the flow rate as shown by the characteristic line AE by not using the flow rate signal.

, Tn in FIG. 7 indicate pump head curves corresponding to changes in the effective blade width d3. Curve T1
Is the maximum pump lift, which is a change when the blade effective width d3 is the largest (for example, at 10 mm). The curves T2, T3, T4, as the blade effective width d3 is gradually reduced.
It changes like T5. Effective blade width d3 is approximately several mm
Until (curve T5) is reached, the deadline head HS = 1
00 is almost unchanged, and the maximum discharge amounts QM1, QM2, etc. are gradually decreased, and the downward slope of the slope becomes steeper. Then, when the blade effective width d3 further decreases, the pressure internal loss relatively increases, so that the shutoff head gradually decreases as in the head curve below the curve T6, and finally the blade effective width d3 becomes smaller. 0, lift 0
And converges to the origin B where the amount of water is zero. That is, control can be performed up to a pressure range of zero head.

Next, the control operation will be described. However, it is assumed that the suction pressure is 0 and does not fluctuate, and the discharge pressure and the head are equal. In the case where the set pressure control characteristic by the preamplifier 64 in FIG. 6 is selected so as to have the gradient of the characteristic line AB in FIG. 7, if there is no change in the flow rate during operation with the water amount Q1 and the head (discharge pressure) H1, Operation in this state is maintained.

When the amount of water changes to Q2, the load signal of the amount of water Q2 is input from the flowmeter 54 of FIG. 6 to the preamplifier 64, and in the preamplifier 64, the determined set pressure control characteristic line AB (see FIG. 7). ), A target pressure H2 corresponding to the water amount Q2 is determined, and a pressure change command is input to the electronic unit 63.

The electronic unit 63 compares the current pressure input from the pressure transmitter 60 with the target pressure H2,
The pressure deviation is detected, and a corresponding pressure correction command is input to the relay circuit 62.

The control motor 59 is driven via the relay circuit 62 to rotate the electric pilot valve 28 until the discharge pressure reaches the target pressure, thereby increasing the pressure in the control pressure chamber 25 and reducing the effective width d3. I do. As a result, the pump performance becomes a curve T2, the pressure is set to H2, and the effective blade width control of the open impeller 5 ends.

Then, in the operating state with the water amount Q2 and the pressure H2, a change in the water amount is detected and controlled as described above.

Accordingly, the pressure control characteristic line AB of FIG.
Is set between the points A and B, that is, the discharge amount Q
From 1 to around O, the pump performance is controlled steplessly, and so-called constant estimated end pressure control that approximates the pipeline resistance curve R can be performed. Thereby, compared with the performance curve AE at the time of the constant pressure control, the difference between the two indicated by the shaded portions is the energy saving.

In the example shown in FIG. 15, A-
A relay point M is provided in the middle of the resistance curve R in place of the above-mentioned control characteristic that connects lines B with a straight line, and a line segment AM and a line segment MB are provided.
This is an example in which pressure control is set in two stages. According to the pressure control setting, the control is performed as close as possible to the resistance curve R, so that the energy saving degree is higher than the pressure control setting of AB in FIG.

In order to perform such two-stage control, as shown in FIG. 16, a control device is provided between the preamplifier 64 and the electronic unit 63 as shown in FIG.
An additional path via the ratio setting unit 64A is used. That is, in the case between the characteristic lines AM of FIG.
In the case between the characteristic lines MB in FIG. 15, control is performed from the preamplifier 64 of FIG. 6 to the electronic unit 63 via the ratio setting unit 64 </ b> A in the case of the characteristic line MB. Control by route. It should be noted that three or more levels of control are possible by increasing the ratio setting device.

Next, the case where the discharge pressure constant control is performed will be described. The pressure control characteristic set by the preamplifier 64 is selected so as to have a gradient (horizontal) of the characteristic line AE in FIG. In this case, it is not necessary to detect the amount of water. Initially, the operation is performed with the water amount Q1 and the head H1 (point A), the water amount decreases to Q2, and the head becomes H
2 (point b1), the discharge pressure of the blade plate 6 from the initial equilibrium stationary state to H2 increases through the pilot valve 28 to the pressure in the control pressure chamber 25 from the initial state. The plate 6 is pushed forward, and the blade effective width d3 decreases. As a result, the pump head decreases from the curve T1 to T2, and the operation state (point b2) with the head H1 and the water amount Q2 is established.

Further, a head curve T3, a water amount Q3, a head H1
, The amount of water increases to Q2, and the head becomes H
3 (point b4), since the secondary pressure of the pilot valve 28 is reduced, the pressure in the control pressure chamber 25 is reduced, the balance is broken, and the blade plate 6 moves rearward, and the blade effective width d3 Expands. As a result, the head becomes a curve T2, and the operation state (point b2) of the discharge pressure H1 and the water amount Q2 is achieved. In this way, the discharge pressure constant control is performed.

(Second Embodiment) FIG. 9 shows an example in which the present invention is applied to a multistage pump in which a pair of open type impellers 5 are fixed on the same main shaft 1, and is a basic example. The configuration is shown in Figure 1.
The same components are denoted by the same reference numerals. Each blade plate 6 is fitted and supported on the main shaft 1 so as to be movable in the axial direction, and has a control pressure chamber 25 at a rear portion thereof.
1, 28-2. The first pilot valve 28-1 on the rear side has a primary pressure inlet 28b connected to the intermediate casing 7.
2 and the secondary pressure outlet 28a is in communication with the first control pressure chamber 25 on the rear side. The front-side second pilot valve 28-2 has a primary pressure inlet 28b connected to the control pressure outlet 19 of the discharge port portion 13, and a secondary pressure outlet 28a communicating with the front-side second control pressure chamber 25. ing.

Since the main shaft 1 is formed to be long to support the plurality of impellers 5, the rear end of the main shaft 1 is supported by a casing member via an end metal 70. In this embodiment, there are two stages of the motor pump type, but it is also possible to use a motor direct connection type via a joint mechanism.

(Third Embodiment) FIG. 10 shows an example in which the present invention is applied to a multistage centrifugal pump having a plurality of impellers similar to that of FIG. , The rear end of the main shaft 1 is directly connected to the motor. Other structures are the same as those in FIG. 9, and the same components are denoted by the same reference numerals. The control pressure outlet 19 is connected to the discharge short pipe 15 connected to the upper surface of the discharge port 13.
Is formed.

(Fourth Embodiment) FIG. 11 shows a single-stage pump to which the present invention is applied. An impeller 5 is fixed to the main shaft 1, and the axially movable blade plate 6 is connected to the pump. Instead of using a control pressure chamber in such a casing, it is connected to a piston 81 fitted to an external cylinder 82. The blade plate 6 has a spindle 8 separate from the main shaft 1.
0, and the spindle 80 is connected to a rod 81a of the piston 81 via a joint mechanism 86 so as to be integrally movable in the axial direction.

The operation of the electric pilot valve 128 used in this embodiment is opposite to that of the pilot valve 28 shown in FIG. 5, and when the primary pressure increases with respect to the set pressure, the secondary pressure decreases and the When the secondary pressure decreases, the secondary pressure increases.

FIG. 12 shows a specific example of the electric pilot valve 128, in which the valve element 35 is disposed below the valve port 37 and the pressure set by the coil spring 33 increases.
When the pressure in the primary pressure chamber 31 increases, the valve body 35 rises against the spring, and the valve port 37 narrows, so that the pressure in the secondary pressure chamber 32 decreases.

In FIG. 11, the inside of the cylinder 82 is partitioned by a piston 81 into a balance chamber (variable pressure chamber) 84 on the rear side and a pressure chamber 85 on the front side. The inlet of the pressure chamber 85 is connected to the secondary pressure outlet 128 a, and the outlets of both chambers 84 and 85 are connected to, for example, the suction pipe 14. The control pressure outlet 19 connected to the primary pressure inlet 128b of the electric pilot valve 128 is formed in the discharge short pipe 15 connected to the discharge port 13. The same components as those in FIG. 1 are denoted by the same reference numerals.

In this embodiment, when the pressure of the discharge port 13 becomes equal to or higher than the pressure set by the spring 33, the pressure of the primary pressure inlet 128b increases, so that the pressure of the balance chamber 84 increases. Since the pressure in the pressure chamber 85 decreases due to the decrease in the pressure at the next pressure outlet 128a, the piston 81 moves forward, narrows the effective blade width d3, and reduces the discharge pressure to the set pressure.

(Fifth Embodiment) FIG. 13 shows a multistage pump having a plurality of impellers 5 fixed on a main shaft 1.
The impeller 5 is axially movable together with the main shaft 1, the vane plate 6 is fixed, and the piston 101 is connected to the front end of the main shaft 1 via a bearing joint 100.

As shown in FIG. 11, when the primary pressure rises, the secondary pressure falls,
A configuration in which the secondary pressure increases when the secondary pressure decreases is employed.

The interior of the cylinder 103 is divided into a rear balance chamber (variable pressure chamber) 111 and a front low pressure chamber 112, and the inlet of the balance chamber 111 is connected to the secondary pressure outlet 128 a of the pilot valve 128. The outlet of 112 is connected to the suction pipe 14.

A control pressure chamber 114 behind each blade plate 6 is connected to a low pressure portion such as the suction pipe 14.

In the second to fifth embodiments, their control operations are shown in FIG. 6 by using pilot valves 28 and 128 as electric pilot valves as in the first embodiment. By connecting to such a control device, similar control such as discharge pressure constant control, estimated terminal discharge pressure constant control, or discharge amount constant control can be performed.

[0055]

As described above, according to the present invention, (1) each blade 5b of the open impeller 5 is inserted into each insertion groove 26 of the blade plate 6 opposed thereto, and the blade 5b
The blade effective width d3 is adjusted by the relative axial movement of the blade and the blade plate 6, so that the discharge pressure or the head is changed and adjusted. 14, the both ends in the axial direction of the blade 5b are closed, and the efficiency of the pump is not reduced by escaping into the gap as compared with the gap control type centrifugal pump having the performance of FIG. However, high pump efficiency can be maintained.

(2) Since the structure is such that the effective width d3 is adjusted by movably inserting the blade 5b into the insertion groove 26, the blade 5b is recirculated from the clearance into the blade in the low pressure region as in the conventional clearance control type. The pressure drop does not remain at a certain limit, and the pressure can be controlled over a wide range up to a low pressure range, which was impossible with a gap control type pump, and energy saving can be achieved.

(3) As in the second aspect of the invention, the effective blade width d3 is changed to 0 (discharge pressure 0) by setting the depth d2 of the insertion groove 26 to be not less than the total width d1 of the blade 5b. Therefore, for example, in an application in which the suction pressure changes, the discharge pressure can be kept constant even if the suction pressure change width may reach the entire pump head, and the application range of the pump is expanded.

(4) Pressure control is performed by changing the blade effective width d3, and there is no need to change the speed of a pump drive mechanism such as a motor. There is no reduction in efficiency due to an increase in heat loss in the motor accompanying the decrease.

(5) Since both the pump and the motor have little increase in loss as described above, the pump efficiency is kept high and the energy saving effect is further enhanced.

(6) Compared with the case of electrically controlling the rotation speed of the motor using an inverter or the like, since the hydraulic control is used, the mechanism is simple and the cost is low.

[Brief description of the drawings]

FIG. 1 is a longitudinal sectional view of a discharge characteristic control type centrifugal pump to which the present invention is applied.

FIG. 2 is a rear view of the impeller alone.

FIG. 3 is a sectional view taken along line III-III in FIG. 1 of the vane plate alone.

FIG. 4 is an enlarged view of an impeller and a blade plate portion of FIG. 1;

FIG. 5 is a longitudinal sectional view of the electric pilot valve.

FIG. 6 is a schematic view of a control mechanism of the electric pilot valve.

FIG. 7 is a graph showing control characteristics by the device of the present invention.

FIG. 8 is a graph showing control characteristics of the device of the present invention.

FIG. 9 is a schematic vertical cross-sectional view showing a second embodiment of the present invention.

FIG. 10 is a schematic longitudinal sectional view showing a third embodiment of the present invention.

FIG. 11 is a schematic longitudinal sectional view showing a fourth embodiment of the present invention.

FIG. 12 is a longitudinal sectional view showing a modification of the electric pilot valve.

FIG. 13 is a schematic longitudinal sectional view showing a fifth embodiment of the present invention.

FIG. 14 is a graph showing control characteristics in a conventional device.

FIG. 15 is a graph showing control characteristics of the device of the present invention.

FIG. 16 is a schematic diagram of a control mechanism for controlling the characteristics shown in FIG.

[Explanation of symbols]

 DESCRIPTION OF SYMBOLS 1 Main shaft 4 Pump casing 5 Open impeller 5a Blade 6 Blade plate 25 Control pressure chamber 26 Insertion groove 28,128 Pilot valve

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

[Procedure amendment]

[Submission date] November 7, 1995

[Procedure amendment 1]

[Document name to be amended] Statement

[Correction target item name] 0007

[Correction method] Change

[Correction contents]

For example, when the suction pressure is ± 0 m, the total pump head is 20 m, and the discharge pressure is 20 m (0 + 20),
If the suction pressure rises to 20 m, the pump discharge pressure rises to 20 + 20 = 40 m. In order to keep this at a constant discharge pressure of 20 m, the suction pressure is increased to 20 m as described above.
m, the pump head needs to be 0 m. In order to cope with such an application, it is necessary that the pump performance can be changed to a pressure of 0, but it is difficult to achieve this in gap control of performance as shown in FIG.

[Procedure amendment 2]

[Document name to be amended] Statement

[Correction target item name] 0031

[Correction method] Change

[Correction contents]

.., T9 in FIG. 7 indicate pump lift curves corresponding to changes in the blade effective width d3. Curve T1
Is the maximum pump lift, which is a change when the blade effective width d3 is the largest (for example, at 10 mm). The curves T2, T3, T4, as the blade effective width d3 is gradually reduced.
It changes like T5. Effective blade width d3 is approximately several mm
Until (curve T5) is reached, the deadline head HS = 1
00 is almost unchanged, and the maximum discharge amounts QM1, QM2, etc. are gradually decreased, and the downward slope of the slope becomes steeper. Then, when the blade effective width d3 further decreases, the pressure internal loss relatively increases, so that the shutoff head gradually decreases as in the head curve below the curve T6, and finally the blade effective width d3 becomes smaller. 0, lift 0
And converges to the origin B where the amount of water is zero. That is, control can be performed up to a pressure range of zero head.

[Procedure 3]

[Document name to be amended] Drawing

[Correction target item name] Fig. 1

[Correction method] Change

[Correction contents]

FIG.

 ─────────────────────────────────────────────────── ─── Continuation of the front page (72) Inventor Kazushige Kobayashi 1-18-19 Komatsu, Higashiyodogawa-ku, Osaka City, Osaka Prefecture Mutual Pump Manufacturing Co., Ltd.

Claims (2)

[Claims] 1. An impeller having an axially open one side, and a blade plate opposing the open side of the impeller.
One is fixed in the axial direction, the other can be moved in the axial direction,
The blade plate is formed with an insertion groove into which each blade can be inserted in the axial direction, and each blade is inserted into the insertion groove, and the effective axial width of the blade is adjusted by the relative axial movement of the blade and the blade plate. Discharge characteristics control type centrifugal pump characterized by being possible. 2. The discharge characteristic control type centrifugal pump according to claim 1, wherein the insertion groove is formed to have a bottom in the axial direction, and the inner and outer ends in the radial direction are closed, and the groove depth in the axial direction. The discharge characteristic controlled centrifugal pump is characterized in that the effective width can be made substantially zero by setting the blade width to be equal to or larger than the entire width in the axial direction of the blade.