WO2016158185A1 - Centrifugal pump device - Google Patents

Abstract

A drive unit (9) for driving an impeller (10) maintains the impeller (10) at a position in the center of the movement range of the impeller (10) within a pump chamber (7) in a direction along the rotation axis by means of an attractive force, which has been adjusted through vector control, which acts between a permanent magnet (17) and the drive unit (9). Consequently, even in the event of changes in flow velocity, rotation speed, pressure, or the like, the position of the impeller (10) in the axial direction can be adjusted while changes in rotation torque of a motor are suppressed. Consequently, a centrifugal pump device with improved impeller stability during floating rotation can be provided.

Description

Centrifugal pump device

The present invention relates to a centrifugal pump device, and more particularly to a centrifugal pump device provided with an impeller that sends a liquid by a centrifugal force during rotation.

Recently, a canned motor having a structure in which a motor drive chamber and a rotor chamber are separated by a partition wall is often used. Such a motor is used, for example, in a pump for transporting chemical liquid or pure water in a semiconductor production line used in an environment where dust is not desired, or a pump for transporting biological fluid.

JP 2010-261394 A (Patent Document 1) describes an axial gap type centrifugal pump characterized by non-contact floating of an impeller by a fluid dynamic pressure bearing and a canned motor structure. In an axial gap type centrifugal pump characterized by non-contact floating of the impeller by the hydrodynamic bearing, it is arranged on the opposite side across the impeller so as to cancel the axial suction force acting between the impeller and the motor. A ring-shaped permanent magnet or the like balances the attractive force in the axial direction.

However, the attractive force by these permanent magnets or the like is a component of negative rigidity (unstable element) that tends to approach the impeller more in one direction when the impeller approaches one direction.

Also, for example, when the flow rate is large, the impeller is eccentric in the radial direction due to a pressure difference in the circumferential direction due to the influence of the position of the fluid outlet. As a result, there is a difference between the amount of decrease in the attractive force on the ring magnet side and the amount of decrease in the attractive force on the motor side, so the balance of the axial direction attractive force changes and the impeller's floating position in the axial direction deviates from the center. There was a problem.

Furthermore, in a centrifugal pump, the fluid inlet is often installed on one side near the impeller center. In this case as well, there has been a problem that the floating position of the impeller deviates from the steady floating position due to the fluid force.

As described above, as a method for controlling the axial attractive force changed due to the eccentricity of the impeller, in Japanese Patent Application Laid-Open No. 2010-261394 (Patent Document 1), the motor-side attractive force is balanced with the attractive force change of the ring-shaped magnet portion. In addition, the motor current phase was adjusted. As a result, even if the impeller is eccentric in the radial direction due to disturbances or operating conditions, stable rotation can be maintained without changing the floating position of the impeller in the axial direction.

JP 2010-261394 A

As described above, in the centrifugal pump as described in Japanese Patent Application Laid-Open No. 2010-261394, even if the impeller is eccentric in the radial direction due to disturbance or operating conditions, the impeller is stable without changing the floating position in the axial direction. In order to maintain the rotation, measures are taken such as adjusting the motor current phase so that the motor-side attractive force balances the change in the attractive force of the ring-shaped permanent magnet portion.

However, changing the motor current phase may cause various problems. For example, if the current phase changes from the situation where the motor is operated at the maximum efficiency point, the motor efficiency may be reduced. Further, for example, when the current phase is changed from the situation where the operation is performed at the maximum torque point, the generated torque is reduced, and there is a possibility that the pump output is reduced or the motor is stepped out.

In pump applications where clean conditions are essential, it is necessary to reliably prevent the generation of contaminants and their contamination due to contact between the impeller and the inner wall of the pump chamber. On the other hand, it is desirable to avoid motor efficiency reduction and pump output reduction as much as possible.

The present invention has been made to solve the above-described problems, and an object thereof is to provide a centrifugal pump device capable of achieving both prevention of contamination and prevention of reduction in efficiency and performance. .

In summary, the present invention is a centrifugal pump device that includes a housing including first and second chambers partitioned by a partition wall, and a shaft that intersects the partition wall in the first chamber so as to be rotatable about a rotation axis. An impeller that sends liquid by centrifugal force during rotation, a drive unit that is provided in the second chamber and rotationally drives the impeller via the partition, and an impeller along the first wall surface facing the partition of the first chamber A first magnetic body disposed along the same circle, a second magnetic body embedded in the first wall surface and attracting the first magnetic body, and an impeller along the partition wall A plurality of third magnetic bodies provided on the other surface and arranged along the same circle. The drive unit includes a plurality of coils provided to face the plurality of third magnetic bodies and generate a rotating magnetic field. A first dynamic pressure groove is formed on one surface of the impeller or the first wall surface facing the one surface. A second dynamic pressure groove is formed on the other surface of the impeller or a partition wall facing the other surface. At least one of the first magnetic body and the second magnetic body is formed in an annular shape around the rotation center line of the impeller. The drive unit maintains the position of the impeller at the center of the movable range of the impeller in the direction along the rotation axis in the first chamber by the attractive force acting between the third magnetic body and the drive unit adjusted by vector control. To do.

Preferably, the drive unit increases the magnetic flux current in the vector control when the position of the impeller changes in the direction away from the partition from the center of the movable range.

Preferably, when the drive unit performs high-efficiency driving during steady rotation of the impeller, the first attractive force acting between the first magnetic body and the second magnetic body, and the third The second attractive force acting between the magnetic body and the drive unit is balanced when the impeller is positioned at the center of the movable range of the impeller in the first chamber.

Preferably, the centrifugal pump device further includes a phase estimator for estimating the rotation angle of the impeller for use in vector control.

Preferably, the centrifugal pump device preliminarily measures and stores the fluid force acting on the impeller using at least one or more of the rotation speed, flow rate, discharge pressure, and physical property value of the fluid as parameters. The apparatus further includes a storage unit and a drive control unit that controls the second suction force according to a value stored in the storage unit.

Preferably, the centrifugal pump device further includes a rotation detector for detecting the rotation angle of the impeller for use in vector control.

More preferably, the rotation detector is a magnetic sensor, and the flying position of the impeller is detected by the magnetic sensor.

Preferably, a third dynamic pressure groove is formed on the outer peripheral side surface of the impeller.
Preferably, a third dynamic pressure groove is formed on the wall surface of the first chamber facing the outer peripheral side surface of the impeller.

Preferably, the centrifugal pump device is used for circulating food.
Preferably, the centrifugal pump device is used for circulating pharmaceutical products.

In the conventional configuration, the attractive force on the drive motor side and the negative stiffness component in the axial direction of the attractive force on the ring magnet side to cancel it cause instability of the impeller behavior. By vector control of the motor current, only the field current component (Id component) is actively changed while maintaining the rotational torque of the motor, and the change in the flying position of the impeller is controlled.

In particular, according to the present invention, the electromagnetic force generated in the motor is separated (vector control) into a torque current component (Iq component) and a field current component (Id component), and the Id component is positively changed, so that the impeller It is possible to control the axial suction force acting on the impeller, and to improve the stability of the impeller during the floating rotation.

It is an electric circuit diagram of the motor drive system of the centrifugal pump device according to the present embodiment. It is a wave form diagram for demonstrating the output signal output to an inverter from a PWM driver. It is the control block diagram which showed the motor control part 324 and its periphery structure in a motor drive system. 4 is a control block diagram showing a motor control unit 324 and its peripheral configuration in a motor drive system when a sensorless angle detection process is used for motor drive, which is a modification of FIG. 3. FIG. It is a block diagram of the voltage equation of a motor. It is a vector diagram of motor control. It is a front view which shows the external appearance of the pump part 1 of the centrifugal pump apparatus of embodiment of this invention. It is a side view of the pump part 1 shown in FIG. FIG. 9 is a sectional view taken along line IX-IX in FIG. 8. FIG. 10 is a sectional view taken along line XX in FIG. 9. It is sectional drawing which shows the state which removed the impeller from FIG. FIG. 10 is a cross-sectional view showing a state where the impeller is removed from the cross-sectional view taken along the line XII-XII in FIG. 9. FIG. 10 is a sectional view taken along line XIII-XIII in FIG. 9. The magnitude of the resultant force of the attractive force F1 between the permanent magnets 15 and 16 and the attractive force F2 between the permanent magnet 17 and the magnetic body 18 is zero at a position P1 other than the central position of the movable range in the pump chamber 7 of the impeller 10. It is a figure which shows the force which acts on the impeller 10 when it adjusts so that it may become. The magnitude of the resultant force of the attractive force F1 between the permanent magnets 15 and 16 and the attractive force F2 between the permanent magnet 17 and the magnetic body 18 is zero at the center position P0 of the movable range in the pump chamber 7 of the impeller 10. It is a figure which shows the force which acts on the impeller 10 when adjusting to. It is a figure which shows the radial direction displacement of the impeller rotation speed and the rotation centerline of an impeller. It is the figure which showed arrangement | positioning of the magnetic sensor S for detecting a rotational speed in sectional drawing shown in FIG. It is a time chart which shows the output signal of the magnetic sensor shown in FIG. It is the figure which showed the example of a change of arrangement | positioning of a magnetic sensor. It is a figure for demonstrating the control which feeds back the output of a magnetic sensor to the electric current of a coil. It is a wave form diagram for demonstrating the process which estimates an angle from a magnetic sensor output. It is a schematic block diagram for demonstrating the process in the case of performing sensorless control by a motor drive part. It is the figure which showed the relationship between a flow volume and a floating position during rotation of an impeller. It is the figure which showed the relationship between discharge pressure and a floating position during rotation of an impeller. It is the figure which showed the relationship between rotation speed and a floating position. It is a figure which shows the relationship between the electric current Id and attractive force. It is a figure which shows the example of a change of the structure shown in FIG. FIG. 28 is a sectional view taken along line XXVIII-XXVIII in FIG. 27. It is a figure which shows the modification which further provided the radial dynamic pressure groove in the pump chamber peripheral surface. 4 is a view showing a modification in which a radial dynamic pressure groove is further provided on the outer peripheral surface of the impeller 10. FIG. It is the figure which showed the modification which eliminated the ring magnet from the example shown in FIG. It is a figure which shows the 1st example of the dynamic pressure groove formed in the outer peripheral surface of a shroud. It is a figure which shows the 2nd example of the dynamic pressure groove formed in the outer peripheral surface of a shroud. It is the figure which showed the 1st example of the specific structure of the radial dynamic pressure groove formed in the pump chamber internal peripheral surface. It is the figure which showed the 2nd example of the concrete structure of the radial dynamic pressure groove formed in the pump chamber internal peripheral surface. FIG. 3 is a diagram showing a detailed arrangement of magnets embedded in a shroud of the impeller 10. It is a 1st modification of arrangement | positioning of the permanent magnet shown in FIG. It is a 2nd modification of arrangement | positioning of the permanent magnet shown in FIG. FIG. 37 is a third modification of the arrangement of the permanent magnets shown in FIG. 36. FIG. FIG. 37 is a fourth modification of the arrangement of the permanent magnets shown in FIG. 36. FIG. FIG. 37 is a fifth modification of the arrangement of the permanent magnets shown in FIG. 36. FIG.

Hereinafter, embodiments of the present invention will be described in detail with reference to the drawings. In the drawings, the same or corresponding parts are denoted by the same reference numerals and description thereof will not be repeated.

[Outline of motor control circuit]
FIG. 1 is an electric circuit diagram of a motor drive system of a centrifugal pump device according to the present embodiment. Referring to FIG. 1, motor drive system 300 includes a battery unit 310, an inverter device 320, and a motor unit 331.

The motor unit 331 is a three-phase synchronous motor, such as an SPM type (surface magnet type) synchronous motor, and includes a rotor 332 including a permanent magnet and a stator including a three-phase stator coil. The rotation angle of the rotor 332 is detected by an angle detector 334.

The inverter device 320 includes a smoothing unit 328 that smoothes a DC voltage supplied from the battery unit 310, a three-phase inverter 330, and a calculation unit 322. The calculation unit 322 includes a motor control unit 324 and a PWM driver 326.

The three-phase inverter 330 is composed of six drive elements that are semiconductor switching elements, and outputs a drive current of each phase of the three phases (U, V, W phase) of the motor in a pulse waveform.

FIG. 2 is a waveform diagram for explaining an output signal output from the PWM driver to the inverter. For example, as shown in FIG. 2, the PWM driver 326 performs pulse width modulation on the input current command so as to obtain a current output for driving the three-phase inverter 330 in a sine wave, and gives on / off commands to the six drive elements. .

1, the PWM driver 326 and the motor control unit 324 constitute a calculation unit 322 that is a weak electric circuit portion of the inverter device 320. The arithmetic unit 322 includes a computer, a program executed on the computer, and an electronic circuit.

FIG. 3 is a control block diagram showing the motor control unit 324 and its peripheral configuration in the motor drive system. Referring to FIG. 3, motor control unit 324 includes current command calculation unit 340, torque current control unit 341, magnetic flux current control unit 343, αβ coordinate conversion unit 350, and two-phase / three-phase coordinate conversion unit 352. And a three-phase / two-phase coordinate conversion unit 354 and a rotation coordinate conversion unit 356 on the detection side.

The current command calculation unit 340 includes a torque current command unit 362 and a magnetic flux current setting unit 364. The torque current command unit 362 outputs a torque current command value Iqref in accordance with the torque command given from the host controller. The torque command is calculated from the external rotational speed or torque command. The magnetic flux current setting unit 364 outputs a command value Idref in which the magnetic flux current is determined. The magnetic flux current command value Idref is appropriately set according to the motor characteristics and the like, and is normally set to “0”.

The torque current is hereinafter referred to as “q-axis current”. The magnetic flux current is hereinafter referred to as “d-axis current”. The magnetic flux current is also called exciting current or field current. Regarding the voltage, the torque voltage is referred to as “q-axis voltage”, and the magnetic flux voltage is referred to as “d-axis voltage”. The q axis is an axis indicating a component in the motor rotation direction, and the d axis is an axis perpendicular to the q axis.

The torque current control unit 341 includes a subtraction unit 342 that subtracts the q-axis current detection value Iq from the q-axis current command value Iqref, and an arithmetic processing unit 346 that performs a predetermined arithmetic process on the output of the subtraction unit 342. The arithmetic processing unit 346 performs proportional integration processing in the example shown in FIG.

The torque current control unit 341 controls the q axis current detection value Iq to follow the q axis current command value Iqref given from the torque current command unit 362 of the current command calculation unit 340. The q-axis current detection value Iq is obtained from the detection values Iu and Iv of the current detectors 336 and 337 that detect the drive current of the motor via the three-phase / two-phase coordinate conversion unit 354 and the rotation coordinate conversion unit 356. Torque current control unit 341 outputs q-axis voltage command value Vq.

The magnetic flux current control unit 343 includes a subtraction unit 344 that subtracts the d-axis current detection value Id from the d-axis current command value Idref, and an arithmetic processing unit 348 that performs a predetermined arithmetic process on the output of the subtraction unit 344. . The arithmetic processing unit 348 performs proportional integration processing in the example shown in FIG.

The magnetic flux current control unit 343 performs control so that the d-axis current detection value Id follows the d-axis current command value Idref given from the magnetic flux current setting unit 364 of the current command calculation unit 340. The d-axis current detection value Id is obtained from the detection values Iu and Iv of the current detectors 336 and 337 that detect the drive current of the motor via the three-phase / two-phase coordinate conversion unit 354 and the rotation coordinate conversion unit 356. The magnetic flux current control unit 343 outputs the d-axis voltage command value Vd.

The three-phase / two-phase coordinate conversion unit 354 includes two or three currents flowing through the U phase, V phase, and W phase of the motor, for example, a U phase current Iu and a V phase current Iv. Are detected values Iα and Iβ of actual currents (actual current on the α axis and actual current on the β axis) of the stationary quadrature two-phase coordinate component.

Based on the motor rotor angle θmea detected by the angle detector 334, the rotational coordinate conversion unit 356 converts the detected values Iα and Iβ of the static quadrature two-phase coordinate component into the detected values Iq and Id on the q and d axes. Convert to

The αβ coordinate conversion unit 350 uses the q-axis voltage command value Vq and the d-axis voltage command value Vd based on the motor rotor angle θ detected by the motor angle detector, that is, the motor rotor phase, to command the actual voltage of the fixed two-phase coordinate component. Convert to values Vα and Vβ.

The two-phase / three-phase coordinate conversion unit 352 uses the actual voltage command values Vα and Vβ output from the αβ coordinate conversion unit 350 as the three-phase AC voltage command values Vu for controlling the U phase, V phase, and W phase of the motor. , Vv, Vw.

The PWM driver 326 and the three-phase inverter 330 convert the voltage command values Vu, Vv, and Vw output from the two-phase / three-phase coordinate conversion unit 352 as described above to convert the motor drive currents Iu, Iv, and Iw into power. Output.

FIG. 4 is a modified example of FIG. 3 and is a control block diagram showing a motor control unit 324 and its peripheral configuration in the motor drive system when sensorless angle detection processing is used for motor drive.

FIG. 4 shows a state in which control is performed using the motor rotor angle θest output from the phase estimator 358.

In the following, a formula for estimating the angle by the phase estimator 358 will be described.
The motor equivalent circuit equation shown in the following (Equation 1) is expressed by the equation shown in (Equation 2) in the dq coordinate system. When the expression shown in (Expression 2) is further converted into the α-β coordinate system, the expression shown in (Expression 3) is obtained.

Here, when the expression shown in (Expression 3) is expressed using Iq, the following expression (Expression 4) is obtained. Based on this equation, the phase estimator 358 estimates the angle (phase θ) of the motor rotor.

In the above equation, R is the resistance value of the armature winding, Ld is the d-axis inductance, Lq is the q-axis inductance, and KE is the induced voltage constant. R, Ld, Lq, and KE are known values, Iα and Iβ are detected values, Vα and Vβ are calculated values during vector control, and are known during position estimation. Therefore, the angle (phase) of the motor rotor can be estimated.

Here, a block diagram and a vector diagram are shown for the equation (Equation 1). FIG. 5 is a block diagram of the motor voltage equation. FIG. 6 is a vector diagram of motor control.

In FIG. 6, the d-axis and the q-axis indicate the magnetic pole directions of the permanent magnet. The electron interlinkage magnetic flux φa is on the d axis. Ldid and Lqiq are the interlinkage magnetic flux components of each axis, and the total interlinkage magnetic flux of the motor is φ0 from FIG. Here, the input current ia is a combined current of id and iq. Also, β represents the phase angle of the input current ia with respect to the q axis. β represents a current phase difference based on the induced voltage ωφa at no load.

The terminal voltage Va is a voltage obtained by adding a voltage drop Raia to the electric coil winding resistance Ra to V0. V0 is a voltage obtained by adding the electric reaction voltages ωLdid and ωLqiq of each axis to the induced voltage ωφa. The phase difference between the voltages V0 and φ0 is 90 degrees.

From the vector diagram of FIG. 6, it can be seen that the interlinkage magnetic flux φa is the field magnetic flux of the permanent magnet, and the magnetic flux of the armature reaction that affects the magnitude of the field magnetic flux is Ldid. Since the d-axis component affects the suction force, controlling id to zero suppresses a change in the suction force. When β is zero, the current phase of each phase coincides with the induced voltage ωφa at no load. In this case, the current vector ia moves on the q axis according to the load state.

It can be seen that the suction force can be changed by changing the d-axis component. Here, the d-axis component is dared to control the suction force in the axial direction.

Before explaining the details of the suction force control, the basic configuration of the pump unit will be described below.

[Description of basic configuration of pump unit]
The motor control of the centrifugal pump device according to the present embodiment has been described with reference to FIGS. Next, the basic configuration of the pump unit of the centrifugal pump device of the present embodiment will be described.

FIG. 7 is a front view showing an appearance of the pump unit 1 of the centrifugal pump device according to the embodiment of the present invention. FIG. 8 is a side view of the pump unit 1 shown in FIG. 9 is a cross-sectional view taken along line IX-IX in FIG. 10 is a cross-sectional view taken along line XX of FIG.

7 to 10, the pump unit 1 of the centrifugal pump device includes a housing 2 formed of a nonmagnetic material. The housing 2 includes a columnar main body 3, a cylindrical inflow port 4 erected at the center of one end surface of the main body 3, and a cylindrical outflow port 5 provided on the outer peripheral surface of the main body 3. Including. The outflow port 5 extends in the tangential direction of the outer peripheral surface of the main body 3.

In the housing 2, as shown in FIG. 9, a pump chamber 7 and a motor chamber 8 partitioned by a partition wall 6 are provided. As shown in FIGS. 9 and 10, a disc-like impeller 10 having a through hole 10 a at the center is rotatably provided in the pump chamber 7. The impeller 10 includes two shrouds 11 and 12 each having a donut plate shape and a plurality of (for example, six) vanes 13 formed between the two shrouds 11 and 12. The shroud 11 is disposed on the inflow port 4 side, and the shroud 12 is disposed on the partition wall 6 side. The shrouds 11 and 12 and the vane 13 are made of a nonmagnetic material.

Between the two shrouds 11 and 12, a plurality of (in this case, six) liquid passages 14 partitioned by a plurality of vanes 13 are formed. As shown in FIG. 10, the liquid passage 14 communicates with the central through hole 10 a of the impeller 10, and starts from the through hole 10 a of the impeller 10 and extends so that the width gradually increases to the outer peripheral edge. In other words, the vane 13 is formed between two adjacent liquid passages 14. In the present embodiment, the plurality of vanes 13 are provided at equiangular intervals and formed in the same shape. Accordingly, the plurality of liquid passages 14 are provided at equiangular intervals and are formed in the same shape.

When the impeller 10 is driven to rotate, the liquid flowing in from the inflow port 4 is sent from the through hole 10a to the outer periphery of the impeller 10 through the liquid passage 14 by centrifugal force and flows out from the outflow port 5.

<Description of dynamic pressure groove>
11 is a cross-sectional view showing a state where the impeller is removed from FIG. 12 is a cross-sectional view showing a state where the impeller is removed from the cross-sectional view taken along the line XII-XII in FIG.

As shown in FIGS. 11 and 12, a plurality of dynamic pressure grooves 21 are formed on the surface of the partition wall 6 facing the shroud 12 of the impeller 10, and a plurality of dynamic pressures are formed on the inner wall of the pump chamber 7 facing the shroud 11. A groove 22 is formed. When the rotational speed of the impeller 10 exceeds a predetermined rotational speed, a dynamic pressure bearing effect is generated between each of the dynamic pressure grooves 21 and 22 and the impeller 10. Thereby, a drag force is generated from each of the dynamic pressure grooves 21 and 22 against the impeller 10, and the impeller 10 rotates in a non-contact state in the pump chamber 7. That is, the position of the impeller 10 in the axial direction is supported by the dynamic pressure groove 21 and the dynamic pressure groove 22.

More specifically, the plurality of dynamic pressure grooves 21 are formed in a size corresponding to the shroud 12 of the impeller 10, as shown in FIG. Each dynamic pressure groove 21 has one end on the periphery (circumference) of a circular portion slightly spaced from the center of the partition wall 6 and has a width up to the vicinity of the outer edge of the partition wall 6 in a spiral shape (in other words, curved). It extends to gradually spread. The plurality of dynamic pressure grooves 21 have substantially the same shape and are arranged at substantially the same interval. The dynamic pressure groove 21 is a recess, and the depth of the dynamic pressure groove 21 is preferably about 0.005 to 0.4 mm. The number of the dynamic pressure grooves 21 is preferably about 6 to 36.

In FIG. 11, ten dynamic pressure grooves 21 are arranged at an equal angle with respect to the central axis of the impeller 10. Since the dynamic pressure groove 21 has a so-called inward spiral groove shape, when the impeller 10 rotates in the clockwise direction, the liquid pressure increases from the outer diameter portion to the inner diameter portion of the dynamic pressure groove 21. For this reason, a repulsive force is generated between the impeller 10 and the partition wall 6, and this becomes a dynamic pressure.

Thus, due to the hydrodynamic bearing effect formed between the impeller 10 and the plurality of hydrodynamic grooves 21, the impeller 10 is separated from the partition wall 6 and rotates in a non-contact state. For this reason, a liquid flow path is ensured between the impeller 10 and the partition wall 6. Further, in a normal state, a portion between the impeller 10 and the partition wall 6 is stirred by the dynamic pressure groove 21 and a liquid flow (leakage flow rate) due to a pressure difference between the inner and outer diameter portions of the impeller generated by the pump operation. Generation of typical liquid retention can be prevented.

Also, the corner portion of the dynamic pressure groove 21 is preferably rounded so as to have an R of at least 0.05 mm.

Further, as shown in FIG. 12, the plurality of dynamic pressure grooves 22 are formed in a size corresponding to the shroud 11 of the impeller 10 as with the plurality of dynamic pressure grooves 21. Each dynamic pressure groove 22 has one end on the periphery (circumference) of a circular portion slightly spaced from the center of the inner wall of the pump chamber 7, and spirally (in other words, curved) on the inner wall of the pump chamber 7. It extends so that the width gradually increases to the vicinity of the outer edge. The plurality of dynamic pressure grooves 22 have substantially the same shape and are arranged at substantially the same interval. The dynamic pressure groove 22 is a recess, and the depth of the dynamic pressure groove 22 is preferably about 0.005 to 0.4 mm. The number of the dynamic pressure grooves 22 is preferably about 6 to 36. In FIG. 12, ten dynamic pressure grooves 22 are arranged at an equal angle with respect to the central axis of the impeller 10.

Note that the corners of the dynamic pressure grooves 22 are preferably rounded so as to have an R of at least 0.05 mm.

Thus, due to the hydrodynamic bearing effect formed between the impeller 10 and the plurality of hydrodynamic grooves 22, the impeller 10 is separated from the inner wall of the pump chamber 7 and rotates in a non-contact state. Moreover, when the pump part 1 receives an external impact or when the dynamic pressure by the dynamic pressure groove 21 becomes excessive, it is possible to prevent the impeller 10 from closely contacting the inner wall of the pump chamber 7. The dynamic pressure generated by the dynamic pressure groove 21 and the dynamic pressure generated by the dynamic pressure groove 22 may be different.

However, it is preferable that the impeller 10 rotates with the gap between the shroud 12 of the impeller 10 and the partition wall 6 and the gap between the shroud 11 of the impeller 10 and the inner wall of the pump chamber 7 being substantially the same. When disturbance such as fluid force acting on the impeller 10 is large and one gap is narrowed, the dynamic pressure by the dynamic pressure groove on the narrowing side is made larger than the dynamic pressure by the other dynamic pressure groove, To make the dynamic pressure grooves 21 and 22 different in shape.

11 and 12, each of the dynamic pressure grooves 21 and 22 has an inward spiral groove shape, but the dynamic pressure grooves 21 and 22 having other shapes may be used.

<Description of arrangement of permanent magnet and coil>
Referring to FIGS. 9 and 10 again, a plurality of (for example, eight) permanent magnets 17 are embedded in the shroud 12. The plurality of permanent magnets 17 are arranged with gaps along the same circle at equal angular intervals so that adjacent magnetic poles are different from each other. In other words, the permanent magnet 17 with the N pole facing the motor chamber 8 side and the permanent magnet 17 with the S pole facing the motor chamber 8 side are alternately provided along the same circle with gaps provided at equal angular intervals. Has been placed.

FIG. 13 is a cross-sectional view taken along line XIII-XIII in FIG. With reference to FIGS. 9 and 13, a plurality of (for example, nine) coils 20 are provided in the motor chamber 8. The plurality of coils 20 are arranged along the same circle at equal angular intervals so as to face the plurality of permanent magnets 17 of the impeller 10 with a partition wall interposed therebetween. In the coil 20, coil wiring is wound around a core portion (magnetic body 18) where a magnetic body or the like is disposed.

However, the magnetic body in the coil of FIG. 13 may be a laminated steel plate, a dust core or other magnetic body.

Although not shown, the core portion may be an air core.
A magnetic body 19 serving as a back yoke is disposed on the opposite side of the partition walls 6 of the plurality of coils 20 to strengthen the magnetic flux of the coils 20. The back yoke may not be provided.

The voltage is applied to the nine coils 20 by, for example, a 120-degree energization method. That is, nine coils 20 are grouped by three. U-phase, V-phase, and W-phase three-phase voltages VU, VV, and VW are applied to the first to third coils 20 of each group. A positive voltage is applied to the first coil 20 during a period of 0 to 120 degrees, 0 V is applied during a period of 120 to 180 degrees, a negative voltage is applied during a period of 180 to 300 degrees, and 300 to 360 degrees. 0V is applied during this period. Therefore, the end face of the first coil 20 (the end face on the impeller 10 side) becomes the N pole during the period of 0 to 120 degrees and becomes the S pole during the period of 180 to 300 degrees. The phase of the voltage VV is 120 degrees behind the voltage VU, and the phase of the voltage VW is 120 degrees behind the voltage VV. Therefore, by applying the voltages VU, VV, and VW to the first to third coils 20 respectively, a rotating magnetic field can be formed, and the attractive force between the plurality of coils 20 and the plurality of permanent magnets 17 of the impeller 10. The impeller 10 can be rotated by the repulsive force.

In FIG. 14, the magnitude of the resultant force of the attractive force F <b> 1 between the permanent magnets 15 and 16 and the attractive force F <b> 2 between the permanent magnet 17 and the magnetic body 18 is other than the central position of the movable range in the pump chamber 7 of the impeller 10. It is a figure which shows the force which acts on the impeller 10 when it adjusts so that it may become zero in the position P1. However, the rotation speed of the impeller 10 is kept at the rated value.

14 represents the position of the impeller 10 (the left side in the figure is the partition wall 6 side), and the vertical axis represents the acting force on the impeller 10. When the acting force on the impeller 10 acts on the partition wall 6 side, the acting force is negative. The acting force on the impeller 10 includes an attractive force F1 between the permanent magnets 15 and 16, an attractive force F2 between the permanent magnet 17 and the magnetic body 18, a dynamic pressure F3 in the dynamic pressure groove 21, and a dynamic force in the dynamic pressure groove 22. The pressure F4 and the resultant force “net force F5 acting on the impeller” are shown.

The attraction force F1 between the permanent magnets 15 and 16 is set smaller than the attraction force F2 between the permanent magnet 17 and the magnetic body 18, and the floating position of the impeller 10 at which the resultant force becomes zero is higher than the middle of the impeller movable range. It is on the 6th side. The shapes of the dynamic pressure grooves 21 and 22 are the same.

As can be seen from FIG. 14, the floating position of the impeller 10 is largely deviated from the center position of the movable range of the impeller 10 at the position where the net force F5 acting on the impeller 10 becomes zero. As a result, the distance between the rotating impeller 10 and the partition wall 6 is narrowed, and the impeller 10 contacts the partition wall 6 even if a small disturbance force acts on the impeller 10.

On the other hand, in FIG. 15, the magnitude of the resultant force of the attractive force F1 between the permanent magnets 15 and 16 and the attractive force F2 between the permanent magnet 17 and the magnetic body 18 is the movable range in the pump chamber 7 of the impeller 10. It is a figure which shows the force which acts on the impeller 10 when it adjusts so that it may become zero in the center position P0. Also in this case, the rotational speed of the impeller 10 is kept at the rated value.

That is, the attractive force F1 between the permanent magnets 15 and 16 and the attractive force F2 between the permanent magnet 17 and the magnetic body 18 are set to be substantially the same. Further, the shapes of the dynamic pressure grooves 21 and 22 are the same. In the case shown in FIG. 15, the support rigidity with respect to the floating position of the impeller 10 is higher than in the case of FIG. 14. Since the net force F5 acting on the impeller 10 is zero at the center of the movable range, the impeller 10 floats at the center position when no disturbance force acts on the impeller 10.

Thus, the floating position of the impeller 10 is generated in the dynamic pressure grooves 21 and 22 when the impeller 10 rotates, and the attractive force F1 between the permanent magnets 15 and 16, the attractive force F2 between the permanent magnet 17 and the magnetic body 18, and the impeller 10. It is determined by the balance with dynamic pressures F3 and F4. By making F1 and F2 substantially the same and making the shape of the dynamic pressure grooves 21 and 22 the same, the impeller 10 can be floated at the substantially central portion of the pump chamber 7 when the impeller 10 rotates. As shown in FIGS. 9 and 10, since the impeller 10 has a shape in which blades are formed between two disks, the two surfaces facing the inner wall of the housing 2 can have the same shape and the same size. Therefore, it is possible to provide the dynamic pressure grooves 21 and 22 having substantially the same dynamic pressure performance on both sides of the impeller 10.

In this case, since the impeller 10 floats at the center position of the pump chamber 7, the impeller 10 is held at a position farthest from the inner wall of the housing 2. As a result, even if a disturbance force is applied to the impeller 10 when the impeller 10 is lifted and the floating position of the impeller 10 is changed, the possibility that the impeller 10 and the inner wall of the housing 2 are in contact with each other is reduced.

FIG. 16 is a diagram showing the impeller rotational speed and the radial displacement of the impeller rotation center line. More specifically, FIG. 16 shows a centrifugal pump device in which the center line (hereinafter referred to as center line L2) of the permanent magnet 16 is aligned with the center line (hereinafter referred to as center line L1) of the side wall of the pump chamber 7. The relationship between the discharge flow rate of the pump device and the moving direction and moving amount of the rotation center line (hereinafter referred to as rotation center line L3) of the impeller 10 is shown. In FIG. 16, the center line L1 of the side wall of the pump chamber 7 and the center line L2 of the permanent magnets 16a and 16b are arranged at the origin, and the direction of the upstream end of the opening 7a shown in FIG. . It can be seen from FIG. 16 that the rotation center line L3 of the impeller 10 moves so as to be sucked to the upstream end of the opening 7a as the discharge flow rate increases.

Therefore, in the present embodiment, when the impeller 10 is rotated at the rated rotational speed, the inner wall of the pump chamber 7 is arranged such that the rotation center line L3 of the impeller 10 and the center line L1 of the side wall of the pump chamber 7 coincide. The center line L2 of the permanent magnets 16a and 16b is disposed on the opposite side of the opening 7a when viewed from the center line L1, and the impeller 10 is attracted (in other words, energized) to the opposite side of the opening 7a. The distance R between the center line L1 of the side wall of the pump chamber 7 and the center line L2 of the permanent magnet 16a is set according to the operating conditions. That is, the displacement amount may be read from FIG. 16 based on the discharge flow rate when the impeller 10 is rotated at the rated rotation speed, and the interval R between the center lines L1 and L2 may be set as the displacement amount. The distance between the center lines L1 and L2 (the amount of eccentricity) varies depending on the size of the pump and the like, but is preferably 0.1 to 1.0 mm. Thereby, when the impeller 10 is rotated at the rated rotation speed, the rotation center line L3 of the impeller 10 and the center line L1 of the side wall of the pump chamber 7 coincide.

[Control of pump unit impeller position]
Next, the application of vector control to adjust the axial position of the pump impeller having the above-described structure and having a dynamic pressure groove formed in the inner wall of the pump chamber will be described.

FIG. 17 is a view showing the arrangement of the magnetic sensor S for detecting the rotational speed in the cross-sectional view shown in FIG. FIG. 18 is a time chart showing output signals of the magnetic sensor shown in FIG.

Referring to FIG. 17, three magnetic sensors S are provided between three adjacent four magnetic bodies 18 out of nine magnetic bodies 18. The three magnetic sensors S are arranged to face the passage paths of the plurality of permanent magnets 17 of the impeller 10. When the impeller 10 rotates and the S poles and N poles of the plurality of permanent magnets 17 alternately pass near the magnetic sensor S, the level of the output signal of the magnetic sensor S changes in a sine wave shape as shown in FIG. To do. Therefore, the positional relationship between the plurality of permanent magnets 17 and the plurality of magnetic bodies 18 can be detected by detecting the time change of the output signal of the magnetic sensor S, and the timing of flowing current through the plurality of coils 20; The rotation speed of the impeller 10 can be obtained.

Further, when the gap between the impeller 10 and the partition wall 6 is wide, the magnetic field in the vicinity of the magnetic sensor S becomes weak and the amplitude A1 of the output signal of the magnetic sensor S becomes small. When the gap between the impeller 10 and the partition wall 6 is narrow, the magnetic field in the vicinity of the magnetic sensor S becomes strong and the amplitude A2 of the output signal of the magnetic sensor S increases.

Therefore, by detecting the amplitude of the output signal of the magnetic sensor S, the position of the impeller 10 within the movable range of the impeller 10 in the axial direction can be detected.

FIG. 19 is a diagram showing a modification example of the arrangement of the magnetic sensors. In this modified example, nine coils 20 are divided into three groups of three, and three magnetic sensors S are respectively disposed between three of the three groups. Therefore, since the mechanical angle between the three magnetic sensors S is 120 degrees, the floating posture of the rotating impeller 10 can be easily calculated. The timing of flowing current through the nine coils 20 is calculated based on the output signal of any one of the three magnetic sensors S.

19 By arranging the magnetic sensor as shown in FIG. 19, it is possible to separate the rotation (tilt) and translation (parallel movement) of the impeller and detect the flying position more accurately.

FIG. 20 is a diagram for explaining control for feeding back the output of the magnetic sensor to the coil current. For ease of understanding, FIG. 3 is simplified and shown in FIG. 20 along with a sectional view of the pump. The controller 42 includes a rotation speed calculator 50 and a position determiner 51. The rotational speed calculator 50 obtains the rotational speed of the impeller 10 based on the output signals of the three magnetic sensors S, and outputs a signal φR indicating the rotational speed. The position determiner 51 is based on the signal φP indicating the position of the impeller 10 generated by the position calculator 49 and the signal φR indicating the rotation speed of the impeller 10 generated by the rotation speed calculator 50. Is within the normal range, and a signal φD indicating the determination result is output. The reason why the rotational speed of the impeller 10 is referred to at the time of determination is that the hydrodynamic bearing effect of the dynamic pressure grooves 21 and 22 changes depending on the rotational speed of the impeller 10, the position of the impeller 10 changes, and the impeller 10 is generated depending on the rotational speed This is because the fluid force to be changed changes. In addition, when the rotation speed is fixed, the rotation speed calculator 50 may be removed.

Further, when determining whether or not the position of the impeller 10 is within the normal range, the viscosity information of the liquid may be referred to instead of the rotational speed of the impeller 10 or in addition to the rotational speed of the impeller 10. This is because the dynamic pressure bearing effect of the dynamic pressure grooves 21 and 22 changes depending on the viscosity of the liquid, and the position of the impeller 10 changes.

Further, in this centrifugal pump device, when the impeller 10 is not rotating, the dynamic pressure bearing effect of the dynamic pressure grooves 21 and 22 does not occur. Therefore, the attractive force F1 between the permanent magnets 15 and 16 and the permanent magnet 17 and The impeller 10 and the inner wall of the housing 2 are in contact with each other by the attractive force F2 between the magnetic bodies 18. Therefore, the impeller 10 does not rotate at the normal axial position at the start of rotation and at the time of low speed rotation. Therefore, when the signal φR indicating the rotation speed is not used for position determination, the output signal φD of the position determination device 51 is forcibly set to a normal position for a certain time from the start of rotation until the rated rotation speed is reached. You may make it the signal which shows that there exists.

FIG. 21 is a waveform diagram for explaining the process of estimating the angle from the magnetic sensor output. Specifically, the rotation angle estimator 372 shown in FIG. 3 has a multiplication processing unit, uses at least one output signal of the magnetic sensor, and uses, for example, the zero cross position of the magnetic sensor output signal as a reference. The rotation synchronization pulse is generated and multiplied to generate a multiplied pulse to estimate the angle of the motor rotor.

FIG. 22 is a schematic block diagram for explaining processing when sensorless control is performed by the motor drive unit. For ease of understanding, FIG. 4 is simplified and shown in FIG. 22 along with a sectional view of the pump. Referring to FIG. 22, at least one piece of information of rotation speed N, flow rate J, discharge pressure P, and fluid information (physical property value) Y is input to impeller attitude determination unit 70 of controller 42A. The storage unit 71 stores impeller behavior measured in advance under each condition. Based on the information stored in the storage unit 71, the Id component of the motor drive current is controlled so that the impeller is levitated and rotated to a substantially central position.

When the pump operation is started, fluid force is applied to the impeller. In the configurations shown in FIGS. 7 to 9, the asymmetry of the shape due to the inlet (inlet) being one place on one side with respect to the axial direction, the impeller moves toward the inlet as the flow rate and the number of rotations increase.

As shown in FIG. 10, the impeller 10 moves to the outlet (opening 7a) side due to the pressure balance around the impeller due to the asymmetry of the shape having one outlet (opening 7a) in the circumferential direction. At this time, when the suction force in the axial direction varies depending on the eccentricity of the impeller 10, the impeller is displaced in the axial direction. In the configuration shown in FIG. 9, since one side is a motor and one side is a ring-shaped permanent magnet, adjustment is made so that the attractive force is balanced at zero eccentricity, but the attractive force balance may be lost during eccentricity.

FIG. 23 is a diagram showing the relationship between the flow rate and the flying position during the rotation of the impeller. If such a relationship is examined in advance, the deviation d1 of the flying position can be obtained if the flow rate J1 is known. FIG. 24 is a diagram illustrating the relationship between the discharge pressure and the flying position during the rotation of the impeller. If such a relationship is examined in advance, the deviation d2 of the flying position can be obtained if the discharge pressure P2 is known. From experience, the influence on the flow rate is larger than the influence on the discharge pressure. FIG. 25 is a diagram showing the relationship between the rotational speed and the flying position. When the rotation is stopped, no dynamic pressure is generated, and the impeller is in contact with one of the partition walls. However, when the impeller starts rotating, the impeller floats to the center position due to the balance between the upper and lower dynamic pressure and the magnetic attractive force. Also in this case, the impeller is displaced in the axial direction when the flow rate or the discharge pressure increases due to the increase in the rotational speed. If such a relationship is examined in advance, the deviation d3 of the flying position can be obtained if the rotational speed N3 is known.

FIG. 26 is a diagram showing the relationship between the current Id and the attractive force. In the pump device of the present embodiment, when the Id component is zero at low flow rate, the upper and lower force balance is achieved and the impeller floats to the center position. Even in the state where the Id component is zero, the motor drive unit generates a magnetic attractive force. When the impeller flying position shifts with increasing flow rate, the suction force of the motor driving unit can be changed by increasing or decreasing the Id component. The Id component may be increased or decreased in accordance with the relationship of FIG. 26 so that the amount of deviation of the impeller flying position is estimated from the relationship of FIGS. 23 to 25 and the suction force is changed to correspond to this.

[Various modifications]
FIG. 27 is a diagram illustrating a modification of the configuration illustrated in FIG. FIG. 28 is a sectional view taken along line XXVIII-XXVIII in FIG. In this modification, a tooth 18 </ b> A is disposed at the tip of the core 18. FIG. 27 shows a configuration in which a magnetic body having a larger area is arranged on the side of the stator of FIG. 9 that faces the rotor. As a result, a large area of the rotor facing the permanent magnet can be ensured, so that motor output and motor efficiency can be improved.

Fig. 29 and Fig. 30 show further modifications (addition of radial dynamic pressure grooves). FIG. 29 is a view showing a modification in which a radial dynamic pressure groove is further provided on the peripheral surface of the pump chamber. Thereby, the disturbance resistance can be further ensured, and the impeller can be stably rotated. FIG. 30 is a view showing a modification in which a radial dynamic pressure groove is further provided on the outer peripheral surface of the impeller 10. As shown in FIG. 30, the formation of the dynamic pressure grooves 61 and 62 on the outer peripheral surface of the impeller 10 is more workable than the process of forming the dynamic pressure grooves 161 and 162 on the inner peripheral surface of the pump chamber 7 as shown in FIG. Good. The example shown in FIG. 30 can ensure more disturbance resistance as in the example shown in FIG. 29, and the impeller can be stably rotated.

Fig. 31 shows a further modification (the ring magnet is omitted). FIG. 31 is a view showing a modification in which the ring magnet is eliminated from the example shown in FIG. When a sufficient radial supporting force can be obtained by the radial dynamic pressure, the number of parts can be reduced by eliminating the ring magnet, and the size, weight, and cost can be reduced. In this case, the performance of the dynamic pressure groove provided on the inner wall of the pump chamber is designed to be different between the drive unit side and the other, so that the magnetic attraction force by the drive unit and the floating force by the dynamic pressure unit are approximately the impeller movable range. Set to balance in the center.

<Example of configuration of radial dynamic pressure groove provided on outer surface of shroud of impeller>
FIG. 32 is a view showing a first example of a dynamic pressure groove formed on the outer peripheral surface of the shroud. FIG. 33 is a diagram illustrating a second example of the dynamic pressure grooves formed on the outer peripheral surface of the shroud.

32, the dynamic pressure grooves 61 and 62 are formed on the outer peripheral surfaces of the shrouds 11 and 12, respectively. The tips of the dynamic pressure grooves 61 and 62 are directed in the direction opposite to the rotation direction of the impeller 10. When the impeller 10 rotates in the direction of the arrow, the liquid pressure increases toward the tip portions of the dynamic pressure grooves 61 and 62. For this reason, a repulsive force is generated between the impeller 10 and the inner peripheral surface of the pump chamber 7, and this becomes a dynamic pressure.

33, the dynamic pressure grooves 64 and 65 are formed not on the inner peripheral surface side of the pump chamber 7 but on the outer peripheral surfaces of the shrouds 11 and 12, respectively. The depth of each of the dynamic pressure grooves 64 and 65 is gradually shallower in the direction opposite to the rotation direction of the impeller 10. Also in this modified example, when the impeller 10 rotates in the direction of the arrow, the pressure of the liquid increases toward the tips of the dynamic pressure grooves 64 and 65. For this reason, a repulsive force is generated between the impeller 10 and the inner peripheral surface of the pump chamber 7, and this becomes a dynamic pressure.

<Configuration example of radial dynamic pressure groove provided on pump chamber inner circumferential surface>
FIG. 34 is a view showing a first example of a specific configuration of a radial dynamic pressure groove formed on the peripheral surface of the pump chamber. In FIG. 34, V-shaped dynamic pressure grooves 161 are formed at a predetermined pitch in the rotation direction of the impeller 10 in a region facing the outer peripheral surface of the shroud 11 in the inner peripheral surface of the pump chamber 7. The tip (acute angle portion) of the V-shaped dynamic pressure groove 161 is directed in the rotational direction of the impeller 10. Similarly, V-shaped dynamic pressure grooves 162 are formed at a predetermined pitch in the rotation direction of the impeller 10 in a region facing the outer peripheral surface of the shroud 12 in the inner peripheral surface of the pump chamber 7. The tip (acute angle portion) of the V-shaped dynamic pressure groove 162 is directed in the rotation direction of the impeller 10. A groove 63 having a predetermined depth is formed in a ring shape in a region of the inner peripheral surface of the pump chamber 7 facing the gap between the shrouds 11 and 12. When the impeller 10 rotates in the direction of the arrow, the liquid pressure increases toward the tips of the dynamic pressure grooves 161 and 162. For this reason, a repulsive force is generated between the impeller 10 and the inner peripheral surface of the pump chamber 7, and this becomes a dynamic pressure.

FIG. 35 is a view showing a second example of the specific configuration of the radial dynamic pressure groove formed on the peripheral surface of the pump chamber. In FIG. 35, in this modification, the dynamic pressure grooves 161 and 162 are replaced with dynamic pressure grooves 164 and 165, respectively. Each of the dynamic pressure grooves 164 and 165 is formed in a belt shape and extends in the rotation direction of the impeller 10. The depths of the dynamic pressure grooves 164 and 165 gradually become shallower in the rotation direction of the impeller 10. Also in this modification, when the impeller 10 rotates in the direction of the arrow, the pressure of the liquid increases toward the tips of the dynamic pressure grooves 164 and 165. For this reason, a repulsive force is generated between the impeller 10 and the inner peripheral surface of the pump chamber 7, and this becomes a dynamic pressure.

Finally, this embodiment will be summarized with reference to the drawings again. Referring mainly to FIG. 9, the centrifugal pump device according to the present embodiment intersects the partition wall 6 in the pump chamber 7 and the housing 2 including the pump chamber 7 and the motor chamber 8 partitioned by the partition wall 6. An impeller 10 that is rotatably provided with a shaft as a rotation shaft and that sends liquid by centrifugal force during rotation, a drive unit 9 that is provided in the motor chamber 8 and that rotates the impeller 10 via the partition wall 6, and a pump chamber 7 is provided on one surface of the impeller 10 along the first wall surface facing the partition wall 6, and the permanent magnet 15 disposed along the same circle, and the permanent magnet embedded in the first wall surface and attracting the permanent magnet 15. 16 and a plurality of permanent magnets 17 provided on the other surface of the impeller 10 along the partition wall 6 and arranged along the same circle. The drive unit 9 is provided to face the plurality of permanent magnets 17 and includes a plurality of coils 20 for generating a rotating magnetic field. A dynamic pressure groove 22 is formed on one surface of the impeller 10 or a first wall surface facing the one surface. A dynamic pressure groove 21 is formed on the other surface of the impeller 10 or the partition wall 6 facing the other surface. At least one of the permanent magnet 15 and the permanent magnet 16 is formed in an annular shape around the rotation center line of the impeller 10.

The drive unit 9 adjusts the position of the impeller 10 in the direction along the rotation axis in the pump chamber 7 by the attractive force (F2 in FIG. 15) acting between the permanent magnet 17 and the drive unit 9 adjusted by vector control. Maintain in the middle of 10 movable ranges. Thereby, even when the flow velocity, the number of revolutions, the pressure, and the like change, the position of the impeller 10 in the axial direction can be adjusted while suppressing the change in the rotational torque of the motor.

Preferably, as described with reference to FIGS. 23 to 26, when the position of the impeller 10 changes in the direction away from the partition wall 6 from the center of the movable range, the drive unit 9 increases the magnetic flux current Id in the vector control. . As a result, the suction force of the drive unit 9 increases, so that the position of the impeller 10 in the axial direction returns to the center.

Preferably, as shown in FIG. 15, when the drive unit 9 is driving with high efficiency during the steady rotation of the impeller 10, the first acting between the permanent magnet 15 and the permanent magnet 16. The attraction force F1 and the second attraction force F2 acting between the permanent magnet 17 and the drive unit 9 are when the impeller 10 is positioned at the center (point P0) of the movable range of the impeller 10 in the pump chamber 7. To balance.

Preferably, as shown in FIG. 4, the centrifugal pump device further includes a phase estimator 358 for estimating the rotation angle of the impeller 10 for use in vector control.

Preferably, as shown in FIG. 22, the centrifugal pump device uses the fluid force acting on the impeller 10 as the fluid information Y such as the rotational speed N, the flow rate J, the discharge pressure P, and the physical property value of the fluid. A storage unit 71 that stores a result measured in advance using at least one of them as a parameter, and a drive control unit that controls the second suction force according to a value stored in the storage unit 71 (Id component command generation) Part 74).

Preferably, the centrifugal pump device further includes a rotation detector for detecting the rotation angle of the impeller 10 for use in vector control. As shown in FIGS. 17 to 19, more preferably, the rotation detector is a magnetic sensor S, and the flying position of the impeller 10 is detected by the magnetic sensor.

Preferably, as shown in FIGS. 30 to 33, dynamic pressure grooves 61 and 62 or dynamic pressure grooves 64 and 65 are formed on the outer side surface of the impeller 10.

Preferably, as shown in FIGS. 29, 34, and 35, the dynamic pressure grooves 161, 162 or the dynamic pressure grooves 164 are provided on the wall surface of the pump chamber 7 that faces the side surface on the outer peripheral side of the impeller 10. , 165 are formed.

<Configuration example of permanent magnet arrangement>
FIG. 36 is a diagram illustrating a first example of a specific configuration of the permanent magnet arrangement. Referring to FIG. 36, a plurality (for example, eight) of permanent magnets 17 are embedded in shroud 12. The plurality of permanent magnets 17 are arranged with gaps along the same circle at equal angular intervals so that adjacent magnetic poles are different from each other. In other words, the permanent magnet 17 with the N pole facing the motor chamber 8 side and the permanent magnet 17 with the S pole facing the motor chamber 8 side are alternately provided along the same circle with gaps provided at equal angular intervals. Has been placed.

37 (a) and 37 (b), the impeller 10 is provided with a plurality of permanent magnets 17 and a plurality of permanent magnets 67. The number of permanent magnets 67 is the same as the number of permanent magnets 17. The permanent magnet 67 is magnetized in the circumferential direction (the rotation direction of the impeller 10). The plurality of permanent magnets 17 and the plurality of permanent magnets 67 are alternately arranged one by one at equal angular intervals along the same circle in a Halbach array structure. In other words, the permanent magnet 17 with the N pole facing the partition wall 6 side and the permanent magnet 17 with the S pole facing the partition wall 6 side are alternately arranged along the same circle with gaps provided at equal angular intervals. Yes. The N pole of each permanent magnet 67 is arranged toward the permanent magnet 17 with the N pole facing the partition 6 side, and the S pole of each permanent magnet 67 is arranged toward the permanent magnet 17 with the S pole facing the partition 6 side. Is done. The shapes of the plurality of permanent magnets 17 are the same, and the shapes of the plurality of permanent magnets 67 are the same. The shape of the permanent magnet 17 and the shape of the permanent magnet 67 may be the same or different. In this modified example, the attractive force between the permanent magnet 17 and the coil 20 can be suppressed, and the magnetic flux that causes the torque can be strengthened, so that the permanent magnet can be most miniaturized. That is, the impeller 10 can be most lightweight and energy efficiency can be increased even when the motor gap is wide.

In another modification shown in FIG. 38, the rotor (the shroud 12 of the impeller 10) includes a permanent magnet 17A magnetized in the rotation axis direction, a permanent magnet 67A magnetized in the circumferential direction, and a magnetic body 70A. It is out. The permanent magnet 17A is arranged so that the magnetic poles of adjacent magnets have different orientations, and the permanent magnet 67A is arranged so that the same magnetic pole as the permanent magnet 17A approaches the end face of the permanent magnet 17A on the partition wall 6 side.

The number of permanent magnets 17A and the number of permanent magnets 67A is the same. The magnetizing direction length of the permanent magnet 67A is shorter than the width of the permanent magnet 17A, and if the center of the magnetizing direction length of the permanent magnet 67A is made coincident with the boundary between adjacent magnets of the permanent magnet 17A, a gap is formed in the circumferential direction. The magnetic body 70A is disposed in the gap. In this case, the magnetic flux is focused on the magnetic body 70A, and a stronger field magnetic flux can be obtained and the torque can be increased compared to the case where there is no magnetic body or the configuration of the normal Halbach arrangement (FIG. 37). Furthermore, in the arrangement of FIG. 38, it is possible to suppress a decrease in the permeance coefficient of the permanent magnets 17A and 67A.

In FIG. 39, in the configuration of FIG. 38, a magnetic body 72 is disposed on the end surface of the permanent magnet 17A opposite to the partition wall 6. Magnetic flux can be further strengthened by the effect of the magnetic body 72.

FIG. 40 shows another magnet arrangement. In the canned motor having the partition wall 6 between the stator and the rotor, the rotor is composed of a permanent magnet 17B magnetized in the rotation axis direction, a permanent magnet 67B magnetized in the circumferential direction, and a magnetic body 70B. The permanent magnet 17B is arranged with a gap in the direction of the magnetic poles of adjacent magnets, and the permanent magnet 67B is arranged in the gap on the partition 6 side. In the case where magnets are arranged in a limited space, this configuration allows the permeance coefficient to be larger than that in FIG. 38 because the flatness of the permanent magnet 17B is reduced. The number of permanent magnets 17B and the number of permanent magnets 67B is the same. The permanent magnet 67B is magnetized in the circumferential direction (rotation direction of the rotor). The plurality of permanent magnets 17B and the plurality of permanent magnets 67B are alternately arranged in a Halbach array structure along the same circle at equal angular intervals one by one. In other words, the permanent magnet 17B with the N pole facing the partition wall 6 and the permanent magnet 17B with the S pole facing the partition wall 6 are alternately arranged along the same circle with gaps provided at equal angular intervals. Yes. The N pole of each permanent magnet 67B is disposed toward the permanent magnet 17B with the N pole directed toward the partition wall 6, and the S pole of each permanent magnet 67B is disposed toward the permanent magnet 17B with the S pole directed toward the partition wall 6. Is done. The shapes of the plurality of permanent magnets 17B are the same, and the shapes of the plurality of permanent magnets 67B are the same. The axial length of the permanent magnet 17B is shorter than the width of the permanent magnet 67B. When the permanent magnet 17B is disposed, a step is formed on the partition wall 6 side, and the magnetic body 70B is disposed at the step portion. Also in this case, the magnetic flux is focused on the magnetic body 70B, and a stronger field magnetic flux can be obtained and the torque can be increased as compared with the case where there is no magnetic body or a normal Halbach arrangement (FIG. 36). Further, it is possible to suppress a decrease in permeance coefficient of the permanent magnets 17B and 67B.

41, the magnetic body 72 is disposed on the end surface of the permanent magnet 17B opposite to the partition wall 6 in the configuration of FIG. Magnetic flux can be further strengthened by the effect of the magnetic body 72.

Further, the plurality of permanent magnets 17 embedded in the shroud 12 may be arranged without providing gaps along the same circle at equal angular intervals so that adjacent magnetic poles are different from each other.

As described above, in this embodiment, in the axial gap type centrifugal pump characterized by the non-contact levitation of the impeller by the fluid dynamic pressure bearing, the torque current component ( Iq component) and field current component (Id component).

And, only the field current component (Id component) is actively changed while maintaining the rotational torque of the motor, and the axial attractive force acting on the impeller is controlled. Thereby, the floating position of the impeller can be controlled, and the stable floating can always be performed near the center of the movable range.

The configuration of the pump unit of the present embodiment is a configuration in which a ring-shaped magnet (second magnetic body) is embedded in a partition opposite to the drive unit 9, and a permanent magnet (first magnetic body) embedded in an impeller. ) Is substantially balanced with the attractive force between the drive unit 9 and the permanent magnet (third magnetic body) in the impeller, and further the restoring force in the radial direction by the first and second magnetic bodies. The impeller can be levitated with a flat structure and a compact configuration.

Note that the centrifugal pump device of the above embodiment may be used for circulating food. Moreover, the centrifugal pump apparatus of the said embodiment may be used in order to circulate a pharmaceutical.

The embodiment disclosed this time should be considered as illustrative in all points and not restrictive. The scope of the present invention is defined by the terms of the claims, rather than the description above, and is intended to include any modifications within the scope and meaning equivalent to the terms of the claims.

1 pump part, 2 housing, 3 body part, 4 inflow port, 5 outflow port, 6 partition, 7 pump room, 7a opening, 8 motor room, 10 impeller, 10a through-hole, 11,12 shroud, 13 vane, 14 Liquid passage, 15, 16, 16a, 16b, 17 permanent magnet, 18, 19 magnetic body, 18A teeth, 20 coils, 21, 22, 61, 62, 64, 65, 161, 162, 164, 165, dynamic pressure grooves, 42, 42A controller, 49 position calculator, 50 rotation number calculator, 51 position determiner, 71 storage unit, 63 groove, 70 impeller attitude determination unit, 74 Id component command generation unit, 300 motor drive system, 310 battery unit, 320 inverter device, 322 arithmetic unit, 324 motor control unit, 3 6 driver, 328 smoothing unit, 330 inverter, 331 motor unit, 332 rotor, 334 angle detector, 336, 337 current detector, 340 current command calculation unit, 341 torque current control unit, 342, 344 subtraction unit, 343 magnetic flux current Control unit, 346, 348 arithmetic processing unit, 350, 352, 354, 356 coordinate conversion unit, 358 phase estimator, 362 torque current command unit, 364 magnetic flux current setting unit, 372 rotation angle estimator, S magnetic sensor.

Claims (11)

1. A housing including first and second chambers partitioned by a partition;
   An impeller that is rotatably provided with an axis that intersects the partition in the first chamber as a rotation axis, and that sends liquid by centrifugal force during rotation;
   A drive unit provided in the second chamber and configured to rotationally drive the impeller via the partition;
   A first magnetic body provided on one surface of the impeller along a first wall surface facing the partition wall of the first chamber and disposed along the same circle;
   A second magnetic body embedded in the first wall surface and attracting the first magnetic body;
   A plurality of third magnetic bodies provided on the other surface of the impeller along the partition wall and disposed along the same circle;
   The drive unit includes a plurality of coils provided to face the plurality of third magnetic bodies and generate a rotating magnetic field,
   A first dynamic pressure groove is formed on the one surface of the impeller or the first wall surface facing the one surface, and a second dynamic pressure groove is formed on the other surface of the impeller or the partition wall facing the first surface;
   At least one of the first magnetic body and the second magnetic body is formed in an annular shape around a rotation center line of the impeller,
   The drive unit adjusts the position of the impeller in the direction along the rotation axis in the first chamber by an attractive force acting between the third magnetic body and the drive unit adjusted by vector control. A centrifugal pump device that maintains the center of the movable range.
2. The centrifugal pump device according to claim 1, wherein the drive unit increases the magnetic flux current in the vector control when the position of the impeller changes in a direction away from the partition from the center of the movable range.
3. During the steady rotation of the impeller, when the drive unit is driving with high efficiency, the first attractive force acting between the first magnetic body and the second magnetic body, The second attraction force acting between the third magnetic body and the drive unit is balanced when the impeller is positioned at the center of the movable range of the impeller in the first chamber. The centrifugal pump device described.
4. The centrifugal pump device according to any one of claims 1 to 3, further comprising a phase estimator for estimating a rotation angle of the impeller for use in the vector control.
5. A storage unit that measures and stores the fluid force acting on the impeller in advance using at least one of the rotational speed, flow rate, discharge pressure, and physical property value of the fluid as parameters;
   The centrifugal pump device according to claim 3, further comprising a drive control unit that controls the second suction force according to a value stored in the storage unit.
6. The centrifugal pump device according to any one of claims 1 to 3, further comprising a rotation detector for detecting a rotation angle of the impeller for use in the vector control.
7. The centrifugal pump device according to claim 6, wherein the rotation detector is a magnetic sensor, and the flying position of the impeller is detected by the magnetic sensor.
8. The centrifugal pump device according to any one of claims 1 to 7, wherein a third dynamic pressure groove is formed on an outer peripheral side surface of the impeller.
9. The centrifugal pump according to any one of claims 1 to 7, wherein a third dynamic pressure groove is formed on a wall surface of the first chamber facing a side surface on the outer peripheral side of the impeller. apparatus.
10. The centrifugal pump device according to any one of claims 1 to 9, wherein the centrifugal pump device is used for circulating food.
11. The centrifugal pump device according to any one of claims 1 to 9, wherein the centrifugal pump device is used for circulating a pharmaceutical product.

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