JP6732329B2 - Fluid servo valve and fluid servo device - Google Patents

Description

The present invention relates to a fluid control device for controlling fluid pressure/flow rate, and more specifically, a fluid servo valve for controlling fluid pressure/flow rate by moving a flapper and a nozzle relatively by an axial drive mechanism. It is about. And a fluid servo device equipped with this fluid servo valve.

1. Trends in the world: Requests from the product side In various fields such as semiconductor manufacturing processes, liquid crystal manufacturing processes, and precision machining, the use of vibration control to block and suppress minute disturbance vibrations is expanding. The fine processing/inspection apparatus such as the scanning electron microscope and the semiconductor exposure apparatus (stepper) used in the above process is required to have a strict vibration permissible condition to guarantee the performance of the apparatus. An active type precision vibration isolation table has been used in which an actuator is used to support a device that is easily affected by disturbance vibration, and an actuator is controlled so as to reduce the vibration.

In recent years, in addition to demands for miniaturization and precision with the improvement of product performance, equipment for improving productivity has been increased in size and speed. Along with this, also in the vibration isolator that supports the apparatus, in addition to the improvement in the performance of removing the disturbance vibration, it is required that the apparatus itself be upsized, the control force be increased, and the multifunction be increased.

2. Conventional example of active vibration isolation table With the trend of increasing the size of the device supported by the vibration isolation table, the air spring type vibration isolation table, which takes advantage of the pneumatic actuator, is widely used for the micro vibration control of ultra-precision equipment. It is like this. FIG. 49 shows a model diagram of a conventional active vibration isolation table using a pneumatic actuator. This active vibration isolation table is a known one as described in Patent Document 1 and Patent Document 2. Plural sets of pneumatic actuators (582a, 582b) for supporting the surface plate 581 are arranged on the floor surface 580. A precision device (details not shown) is mounted on the surface plate 581. The pneumatic actuator is composed of an air chamber 583, which is internally filled with high-pressure air, for supporting a vertical load, and a piston 585 which is inserted above the air chamber via a diaphragm 584. The active vibration isolation table is composed of two control systems of feedback control (FB control) and feedforward control (FF control).

(1) Feedback Control 586, 587a, 587b are acceleration sensors and displacement sensors for respectively detecting vertical and horizontal accelerations of the surface plate 581 and relative displacement of the surface plate 581 with respect to the floor surface 580. Reference numeral 588 denotes an acceleration sensor that detects the acceleration of the floor surface 580 (basic vibration state). Output signals from each of these sensors are input to a feedback control controller G ₁ 589.

(2) Feedforward control 590 is a stage mounted on the surface plate 581. The behavior signal (chain line) of this stage is input to the controller G ₂ 591 for feedforward control. The input signal (two-dot chain line 592) of the driver that drives the stage is used as the behavior signal of the stage.
A servo valve 593 controlled by the outputs of the _two controllers G ₁ and G ₂ is connected to the air chamber 583 via a pipe 593. The internal pressure P _a of the air chamber 583 is controlled by adjusting the flow rate of the compressed air supplied to and discharged from the air chamber 583 by the servo valve 593 which is an electropneumatic converter. Here, the servo valve 593 externally supplies a gas having a supply pressure P _S , receives a control signal from the controllers G ₁ and G ₂ , adjusts to _a desired gas pressure P _a , and outputs the gas. It has a function of exhausting to the atmosphere P ₀ .

(3) Conventional Servo Valve Structure In an active vibration isolation table which is a pneumatic servo device using a pneumatic actuator, a nozzle flapper valve has been mainly used to control the pressure and flow rate of the actuator. This nozzle flapper valve is disclosed, for example, in Patent Document 3, and its operating principle is generally widely used. FIG. 50 is a structural diagram modeling the operating principle of a conventional servo valve. The structure of the servo valve can be roughly divided into an actuator section A-1 and a fluid control section B-1. In the actuator unit A-1, 551 is a magnet (permanent magnet), 552 is a coil, 553 is a body that houses this coil, 554 is a flapper, 555a and 555b are a pair of yokes mounted with their tips facing each other, and 556 is This is the tip of the flapper on the actuator side. Reference numeral 557 is a leaf spring which also serves as a seal member, and 558 is a supporting center portion of the leaf spring.

In the fluid control unit B-1, 560 is a forward nozzle, 561 is a backward nozzle, and 562 is a flapper tip on the fluid control unit side. 563 is a supply port, 564 is an exhaust port, 565 is a load port (control port), and 566 is a control chamber. The gas having the supply pressure P _S is supplied to the control chamber 566 via the forward nozzle 560. At the same time, the gas in the control chamber 566 is discharged to the atmosphere through the backward nozzle. The difference between the inflow amount from the forward direction nozzle and the outflow amount from the reverse direction nozzle determines the control pressure P _a in the control chamber 566 and the outflow amount from the load port 565. However, the structure of the servo valve actually used is a three-dimensional structure in which the magnetic circuit of the permanent magnet and the magnetic circuit of the electromagnet are arranged orthogonally. The basic structure of the pneumatic servo valve was derived by applying the hydraulic servo valve technology with a long history, and is used as a primary control valve (pilot valve) of an electrohydraulic control valve. Is.

Patent Document 4 discloses a gas pressure control valve which is an improved spool type valve used as a secondary control valve of the electrohydraulic control valve. The gas pressure control valve shown in FIG. 51 is roughly divided into a spool 501, a linear motor 502, a speed detection sensor 503, and a control unit 504. The spool 501 is supported by a sleeve 505 so as to be movable in the axial direction, and a left land portion 506, a first plunge 507, a second plunge 508, a central land portion 509, and a right land portion 510 are formed on the spool. There is. The left land portion 506 and the right land portion 510 are supported in a non-contact manner on the sleeve 505 by hydrostatic bearings. A supply port 511, a load port 512, and an exhaust port 513 are formed in the sleeve 505. Reference numeral 514 is a control pressure chamber. The linear motor 502 includes a coil 515, a magnet 516, and a moving body 517. Similarly, the speed detection sensor 503 includes a coil 518, a magnet 519, and a moving body 520. The spool 501 is axially moved and driven by a linear motor 502. 521 and 522 are balance springs that support the spool 501 at both ends. The pressure in the control pressure chamber 514 changes depending on the relative positional relationship between the load port 512 of the sleeve 505 and the central land portion 509. However, by forming the spool 501 and the sleeve 505 in a stepped structure, the drive current and output pressure of the linear motor can be reduced. Can be made proportional.

JP 2006-283966 JP 2007-155038 JP-A-11-294627 Japanese Patent No. 4636830

1. Correspondence to increase in size and speed of on-board equipment In recent years, due to increase in size and speed of the stage included in the equipment of the vibration isolation device, vibration is added to the vibration isolation device during acceleration/deceleration of the stage. The generated vibration was a problem. The solution to this problem is feedforward control (FF control) using the behavior signal of the stage. Feedback control (FB control) improves the time for free vibration to settle, but it cannot reduce the response at the moment of stage acceleration/deceleration. On the other hand, stage FF control has the effect of suppressing momentary disturbance input.

Further, in the case of a heavy stage, the stage tilts due to a change in the load on the vibration isolator accompanying the movement of the stage, and low frequency vibration occurs. When stage FF is applied, vibration can be greatly suppressed compared to the case of only FB control.

2. Conditions Required for Pneumatic Servo Valve Now, the conditions required for the pneumatic servo valve, which is an important basic element that constitutes the active vibration isolation table, are as follows.
(1) High-speed response (2) The primary resonance point of the pneumatic servo valve is sufficiently high and is several hundreds Hz or higher (3) Large exhaust flow rate/low air consumption flow rate

The reason for (1) above is as follows. For example, when the stage mounted on the anti-vibration table starts and stops, the driving reaction force due to the mass movement is input to the surface plate, which is the stage installation surface, as a linear disturbance. In this case, it is possible to reduce the surface plate vibration during acceleration/deceleration by performing stage feedforward control on the vibration isolation device using the stage acceleration signal. In order to quickly converge the platen vibration, a high responsiveness is required for the air valve that drives the pneumatic actuator.

The reason for (2) above is as follows. Even though the response of the pneumatic active vibration isolation system is in the order of several Hz to several tens of Hz, the reason why the servo valve needs a high resonance frequency of several hundred Hz is that it is unique to the pneumatic active vibration isolation system. It is based on.

In order to reduce the peak of the resonance point determined by the air spring of the actuator and the mounting mass, it is essential to apply the acceleration feedback control to the active vibration isolation table. However, when the acceleration feedback control is performed, the open loop characteristic of the active vibration isolation system is such that the gain increases and the phase lags in a wide frequency range. As a result, the gain margin and the phase margin are reduced at the resonance point of the pneumatic servo valve. If the resonance point of the servo valve is low, the active control system will not exhibit sufficient performance.

The above (3) is due to the fact that the feedforward control has come to play an important role in dealing with the increase in the size and speed of the mounted equipment. To that end, a pneumatic servo valve that can control a large flow rate at high speed has become essential. In the behavior signal (two-dot chain line 592) of the stage in the active vibration isolation table of FIG. 49, when the stage starts or stops, the reaction force due to the inertial force is applied to the stage installation surface. For example, if a stage of m=100Kg rises to a constant acceleration of a=10m/s ² in 25ms, a reaction force of F=ma=1000N will be generated. Therefore, in order to suppress this reaction force by feedforward control, a high-speed response/large exhaust flow rate servo valve that can generate a load of 1000 N in 25 ms is required.
3. Problems with conventional pneumatic servo valves

(I) Nozzle flapper valve Conventionally, many vibration isolation devices have been applied for the purpose of vibration isolation from the floor of the installation floor in order to improve the installation environment of precision equipment. In this case, a nozzle flapper valve with a small exhaust flow rate has been mainly used to control the pressure and flow rate of the actuator.

When the mounted equipment includes a source of a fluctuating load such as a stage, FF control can be performed by arranging a plurality of small exhaust flow rate nozzle flapper valves in parallel to increase the flow rate. However, as described later, the nozzle flapper valve consumes the largest amount of flow at the operating point (neutral point) due to its operating principle. As an example of configuring an active vibration isolation table, 4-point support active control is assumed. In this case, the pneumatic actuators are arranged at the four corners, and the units are installed in the horizontal X direction at two points and in the Y direction at two points diagonally. Each actuator also incorporates an actuator that supports the load in the Z direction. Therefore, a total of eight pneumatic actuators are arranged, requiring a total of eight pneumatic servovalves to control each actuator. Therefore, there is a problem that the air consumption flow rate in the steady state significantly increases in the entire active vibration isolation table.

(Ii) Spool valve As shown in the example of the spool type gas pressure control valve disclosed in Patent Document 4, a major problem in applying the spool type valve to the active vibration isolation table is a primary resonance frequency of the movable part. The point is that you can't raise it. In FIG. 51, the resonance frequency of the movable portion including the spool 501 is determined by the mass m of the movable portion and the spring constant K of the balance springs 521 and 522. Since the primary resonance frequency is proportional to (K/m) ^1/2 , the smaller the mass m and the larger the spring constant K, the higher the resonance frequency. However, there is a limit to the weight reduction of the spool shaft that forms the large flow rate opening. Further, in the case of the linear motor 502 utilizing the Lorentz force, the electromechanical conversion efficiency of the generated force with respect to the input current is small, and a large generated force cannot be obtained. Therefore, the stiffness K of the springs 522 and 522 must be reduced, and it is difficult to increase the resonance frequency of the movable portion.

As described above, in order to apply the pneumatic servo valve to the active vibration isolation table, it is necessary to satisfy the condition "the primary resonance point of the servo valve is sufficiently high and several hundred Hz or more". Therefore, in the spool type control valve, acceleration feedback control is performed using the signal of the speed detection sensor 503. That is, the low frequency transfer characteristic is compensated by utilizing the fact that the natural frequency of the system can be reduced from the control side by performing acceleration feedback. In summary, in addition to the linear motor 502 that drives the spool 501, the gas pressure control valve that incorporates the speed detection sensor 503 and the control unit 504 must have a complicated structure, a large size, and a high cost. There were challenges.

Specifically, the invention of claim 1 is a forward nozzle in which a flow path communicates with a fluid supply side,
A reverse nozzle having a flow path communicating with the fluid exhaust side, a flapper provided so as to face the forward end of the forward nozzle and the tip of the reverse nozzle, and a flapper support member for fixing a part of the flapper. A driving means that acts on the flapper so that the flapper moves between the forward nozzle and the backward nozzle; and a supply side formed between the forward nozzle and the facing surface side of the flapper. A flow rate control section or an exhaust side flow rate control section formed between the reverse direction nozzle and the facing surface side of the flapper, and the working fluid supplied from the fluid supply side passes through the forward direction nozzle. Then, it is configured to flow into a control chamber that is a space for accommodating the flapper, pass through the reverse direction nozzle from the control chamber, and flow out to the fluid exhaust side. Alternatively, the effective cross-sectional area of the exhaust side flow rate control unit is configured to change with a non-linear characteristic with respect to the displacement of the flapper, the non-linear characteristic is represented by a downward convex curve or polygonal line, When the effective area at the bending point is A ₀ and the maximum value of the effective area is A _max , 0<A ₀ /A _max <0.5.

Here, the effective cross-sectional area can be defined as, for example, a flow passage area that acts as flow passage resistance on the fluid passing through the supply side flow rate control unit or the exhaust side flow rate control unit. Further, the bending point is a point on the curve where the amount of change in inclination before and after that point changes more than at other points, or a point expressed as an intersection of envelopes. In addition, the bending point is an intersection of two straight lines having different inclinations in the polygonal line.

In the present invention as described above, the facing surface side of the nozzle and the flapper constitutes, for example, a flow rate control unit that structurally forms a narrow flow path, and (1) when one nozzle opening increases, the other nozzle The opening is sufficiently small. (2) Both nozzle openings do not increase at the same time, and the air consumption in a steady state can be suppressed at any position of the flapper. With the above (1) and (2), it is possible to realize a servo valve that can simultaneously realize a large exhaust flow rate control and a low air consumption flow rate.

Specifically, the invention of claim 2 is an annular flow passage forming structure in which the supply-side flow rate control unit and the exhaust-side flow rate control unit form a flow channel having a substantially annular cross section. The forming structure is configured by an inner peripheral surface of a cylindrical portion forming an outer boundary of the generally annular flow path, and an insert body inserted to be spaced apart from the inner peripheral surface in a radial direction.

That is, in the present invention, the flow rate passing through the annular flow path is small when the insert body is deeply entering the tubular portion, and flows in from the tubular portion when the insert body is separated from the tubular portion. The flow rate that flows out is used to make a sharp increase.

Specifically, in the invention of claim 3, a convex portion is formed at a substantially central portion on the supply side and the exhaust side of the flapper to form the insert body, and the inner surface of the forward direction nozzle and the reverse direction nozzle is provided with the convex portion. A tubular part is formed.

That is, in the present invention, the annular flow path forming structure is formed by combining the convex portion formed in the central portion of the flapper and the nozzle. The convex portion formed on the flapper may have a small outer diameter and height, and the resonance frequency of the flapper can be set sufficiently high.

Specifically, the invention of claim 4, the schematic outer diameter of the fluid supply side of the insert .PHI.d _in, as schematically outer diameter .PHI.d _ou fluid exhaust side, which was constructed such that .PHI.d _in> .PHI.d _ou is there.

That is, in the present invention, it has been found that a mechanical pressure feedback action can be obtained with a simple configuration _in which the outer diameters of the inserts provided on the front and back of the flapper are Φd _in >Φd _ou . By this mechanical pressure feedback action, it is possible to achieve a further lower air consumption flow rate in a steady state, and it is possible to greatly improve the linearity of the control pressure characteristic with respect to the current.

Specifically, in the invention of claim 5, P _max is the maximum value of the control chamber pressure, P ₀ is the atmospheric pressure, the control chamber pressure at the neutral point is P _B , the supply side effective sectional area is A ₀ , and the supply side effective disconnection is The maximum value of the area is defined as A _max , η=A ₀ /A _max , and the intermediate pressure ratio β=(P _B -P ₀ )/(P _max -P ₀ ), and the envelope curve of β curve for η The above A ₀ and A _max are set so that η>η _c , where η at the intersection point is η _c .

That is, in the present invention, (i) the outer diameters of the micro pistons formed on the front and back of the flapper are made asymmetric, and (ii) the intake/exhaust passage area is changed in two steps. The inventors have found that the synergistic effect of combining the above (i) and (ii) can further improve the linearity of the control pressure characteristic with respect to the current.

Specifically, in the invention of claim 6, in the annular flow path forming structure composed of the tubular portion and the insert, either the tubular portion or a component constituting the insert is the annular It is configured as a separate member with respect to the housing that houses the flow passage forming structure, and is movable in the radial direction for adjusting the axial center of the annular flow passage with the fastening members such as bolts removed. is there.

That is, in the present invention, it is difficult to secure the following problems based on the unique valve structure different from the conventional valve, (1) ensuring the coaxial accuracy of the two micro pistons. (2) The flapper is an elastic body that is easily deformed. The present invention proposes an assembly/adjustment method that solves the problems (1) and (2) above, and a valve structure that enables this construction method. The insert or the annular flow path forming structural component forming the tubular portion is movable in the radial direction with respect to the housing for adjusting the axial center. The above methods and structures can solve the above problems (1) and (2).

Specifically, in the invention of claim 7, the gap movable in the radial direction is 0.05 mm or more.

That is, in the present invention, the problems (1) and (2) can be cleared by setting the gap in which the annular flow path forming structural component is movable in the radial direction to be 0.05 mm or more.

Specifically, in the invention of claim 8, when the current applied to the drive means is zero, the annular flow path connecting from the fluid supply side to the fluid exhaust side is in close contact between the flapper and the supply side nozzle. It is configured to be shielded by.

That is, in the present invention, the fact that the valve of the present invention has two flow passages, that is, a "radial flow passage" formed by a convex portion and a nozzle and an "axial flow passage" formed by a flapper and a nozzle are used. doing. When the valve drive current I=0, the flow path between the fluid supply source and the pneumatic actuator can be completely cut off, which can be used as a fail-safe function during a power failure. Due to this fail-safe function, even when a heavy load is mounted on the pneumatic actuator, the heavy load does not easily drop.

Specifically, the invention of claim 9 is characterized in that an electromagnet provided so that an attractive force is generated with respect to the flapper, the flapper is deformed by the attractive force of the electromagnet, and the tip portion of the forward direction nozzle, And, or, the drive means is configured to change the distance between the tip end portion of the reverse direction nozzle and the flapper.

That is, in the present invention, by using the magnetic attraction type actuator as the means for driving the flapper, the mass of the movable portion can be reduced and the resonance frequency of the valve can be set high.

Specifically, the invention of claim 10 forms a closed loop magnetic circuit with the flapper, a magnetic pole provided opposite to the flapper for giving an attraction action to the flapper, and a magnetic material component including the magnetic pole and the flapper. Then, the magnetic characteristic of this closed loop magnetic circuit is a linear region in which the characteristic of the magnetic flux density with respect to the magnetizing force is approximately proportional,
The magnetic material has a magnetic saturation region in which the inclination angle of the magnetic flux density characteristic with respect to the magnetizing force changes smaller than the linear region, and when the current flowing to the electromagnet is increased in the displaceable range of the flapper, the magnetic material. The magnetic flux density of the magnetic flux flowing through the component is configured to enter the magnetic saturation region.

That is, in the present invention, the following magnetic attraction type actuator found in the proposal is applied. When a magnetic material suitable for the flapper and a thin disk are used for a conventional actuator having a limited stroke, the displacement characteristic of the flapper with respect to the electric current can be excellent in linearity. By utilizing this magnetic saturation phenomenon, the stroke can be greatly increased without losing the linearity of the flapper displacement characteristic with respect to the current. This actuator can be effectively applied to the servo valve of the present invention which requires a large stroke.

Specifically, in the invention of claim 11, the driving means is arranged with a gap maintained between a coil bobbin fixed to the flapper, an electromagnetic coil wound around the coil bobbin, and an outer peripheral portion of the electromagnetic coil. A permanent magnet, and is configured as a linear actuator using a Lorentz force by the coil bobbin, the electromagnetic coil, and the permanent magnet.

That is, in the present invention, the servo valve of the present invention uses a linear motor type actuator utilizing Lorentz force. When there is no great restriction on the resonance frequency of the movable part, the linear motor type has no restriction on the stroke, which is advantageous for further increasing the flow rate.

Specifically, the invention of claim 12 is arranged such that the first nozzle, the second nozzle of which the flow path communicates with the fluid supply side, and the first nozzle and the second nozzle are arranged between the first nozzle and the second nozzle. A flapper provided so as to face the tip of the second nozzle; driving means that acts on the flapper so that the flapper moves between the first nozzle and the second nozzle; A first flow rate control section formed between the first nozzle and the flapper, a second flow rate control section formed between the second nozzle and the flapper, and a first opening section communicating with the outside. When a second flow rate control flow path having a first flow rate control flow path and a second opening communicating with the outside and the flapper approach the first nozzle, the first flow rate control flow from the supply side through the first flow rate control unit. At the same time that the flow path is opened to the path, the flow path is opened from the second flow rate control path through the second flow rate control unit, and when the flapper approaches the second nozzle, the second flow rate control is performed from the supply side. The flow path is opened to the second flow rate control flow path through the section, and at the same time, the flow path is opened from the first flow rate control flow path via the first flow rate control section.

That is, also in the present invention, it is possible to control the flow rate of the fluid for various purposes while obtaining the same effect as that of the invention of claim 1.

Specifically, the invention of claim 13 relates to the fluid servo valve according to any one of claims 1 to 12, a gas spring for supporting an object to be isolated against a foundation, and a vibration state of the object to be isolated. An acceleration sensor for detecting, and by adjusting the fluid servo valve based on the information from the acceleration sensor, an active controller that gives a gas pressure to the gas spring to reduce the vibration of the vibration isolation target, The fluid servo valve is configured to supply gas from the supply side to the gas spring and exhaust the gas to the exhaust side, and the primary natural frequency of the flapper is configured to be 200 Hz or higher. It is a fluid servo device.

That is, in the present invention, it is possible to realize a fluid servo device having excellent response characteristics and vibration isolation characteristics.

That is, in the present invention, two control flow paths, a first flow control flow path and a second flow control flow path, are provided, and when one control flow path communicates with the supply side, the other control flow path is exhausted. It is configured to contact the side. With this valve configuration, the servo valve of the present invention can also be applied as a four-way valve which has many applications in the hydraulic and pneumatic fields.

Now, when the features of the fluid servo valve according to the present invention are listed,
(1) Low air consumption flow rate (2) Large exhaust flow rate (3) High resonance frequency (4) High-speed response (5) The valve of the present invention that greatly eliminates the drawbacks of the conventional valve that can greatly simplify the structure The widespread use of pneumatic servo systems is expected to greatly accelerate. The effect is remarkable.

1 is a front sectional view of a fluid servo valve according to a first embodiment of the present invention. FIG. 2 shows the positional relationship between each nozzle and the flapper when the valve input current is changed in the first embodiment. FIG. 2a shows a state where the valve input current I=0, FIG. 2b shows a state where 0<I<I _max /2, and FIG. Shows a state of I≈I _max /2 (neutral point), FIG. 2d shows a state of I _max /2<I _max , and FIG. 2e shows a state of I=I _max (maximum value). The AA arrow view of FIG. 2a. 3 is a graph showing the analysis result of the effective area with respect to the current in the first embodiment. The graph which shows the analysis result of the effective area with respect to the electric current of the conventional nozzle flapper valve. The graph which shows the analysis result of control pressure with respect to drive current. The graph which shows the analysis result of the air consumption flow volume with respect to a drive current. The graph which shows the transient response analysis result of the flow volume with respect to time. FIG. 9a is a graph showing the transient response analysis result of the generated load with respect to time, and FIG. 9b is a partially enlarged view of the graph of FIG. 9a. The graph which shows the analysis result of control pressure with respect to drive current. The graph which shows the analysis result of the air consumption flow volume with respect to a drive current. FIG. 6 is a front sectional view of a fluid servo valve according to a second embodiment of the present invention. The AA section enlarged view of FIG. FIG. 14 shows an analytical model diagram of the asymmetrical micro piston and the nozzle part according to the second embodiment, FIG. 14a shows a state where P _S =P ₀ in which a high pressure source is not connected to the valve supply side, and FIG. 14b shows high pressure on the valve supply side. The figure which shows the state which the source was connected. The block diagram when the drive current I is input and the control pressure P _a is output. The graph which shows the effective cross-sectional area with respect to a drive current. The graph which shows the analysis result of control pressure with respect to drive current. The graph which shows the analysis result of control pressure with respect to drive current. 19A and 19B show analysis results of piston displacement/flow rate/pressure characteristics with respect to time, FIG. 19a is a transient response characteristic of piston displacement, FIG. 19b is a transient response characteristic of flow rate, and FIG. 19c is a graph showing transient response characteristic of control pressure. 20A and 20B show the positional relationship between each nozzle and the flapper when the valve input current is changed in the second embodiment. FIG. 20A shows a state where the valve input current I=0, and FIG. 20B shows a state where I≈I _max /2 (neutral point). FIG. 20c is a model diagram showing a state of I=I _max (maximum value). 21a is a graph showing a transient response analysis result of the generated load with respect to time, and FIG. 21b is a partially enlarged view of the graph of FIG. 21a. The figure which shows the concept of the principle of the fluid servo valve which concerns on embodiment of this invention. The model figure which shows the partial enlarged view of a micro piston part. FIG. 24A shows the positional relationship between each nozzle and the flapper when the valve input current is changed in the third embodiment. FIG. 24A shows a state where the valve input current I=0, and FIG. 24B shows a state where I≈I _max /2 (neutral point). FIG. 24c is a model diagram showing a state of I=I _max (maximum value). The graph which shows the effective cross-sectional area with respect to a drive current. The graph which shows the analysis result of control pressure with respect to drive current. 27A and 27B show analysis results of piston displacement/flow rate/pressure characteristics with respect to time, FIG. 27a is a transient response characteristic of piston displacement, FIG. 27b is a transient response characteristic of flow rate, and FIG. 27c is a graph showing transient response characteristic of control pressure. The graph which shows the analysis result of control pressure with respect to drive current. The graph which shows the analysis result of the intermediate pressure ratio with respect to (eta). The graph which shows the analysis result of the air consumption flow volume with respect to a drive current. The graph which shows the analysis result of control pressure with respect to drive current. 32a is a graph showing the transient response analysis result of the generated load with respect to time, and FIG. 32b is a partially enlarged view of the graph of FIG. 32a. FIG. 6 is a front sectional view of a four-way valve fluid servo valve according to a fourth embodiment of the present invention. The figure which shows the state which connected the valve of this embodiment to the pneumatic actuator. FIG. 34A is a model diagram showing the positional relationship between the first flow path part and the second flow path part. FIG. 34A shows the flapper in close contact with the second flow path part, and FIG. 34B shows the flapper with the first flow path part and the second flow path part. FIG. 34c is a view showing the state in the middle of the parts, in which the flapper is in close contact with the first flow path part. The front sectional view of the fluid servo valve concerning Embodiment 5 of the present invention. The front sectional view of the fluid servo valve concerning Embodiment 5 of the present invention. Front sectional drawing of a flapper. Front sectional drawing of a master disk flapper. Front sectional drawing of the servo valve in process Step1. Front sectional drawing of the servo valve in process Step2. Front sectional drawing of the servo valve in process Step3. The graph which shows the analysis result of the flapper displacement with respect to a drive current. FIG. 3 is a model diagram when a cavity is formed in a convex portion of the annular flow path forming structure. FIG. 6 is a model diagram when the convex portion of the annular flow path forming structure is formed on the nozzle side and the cylindrical portion is formed on the flapper side. The case where the convex portion of the annular flow path forming structure is tapered is shown in FIG. 46a where the valve input current I=0, FIG. 46b where I≈I _max /2, and FIG. 46c where I=I _max . FIG. FIG. 5 is a model diagram in which a surface diaphragm is used instead of the annular flow path forming structure on the exhaust side. The model figure which used the surface diaphragm on the supply side and the exhaust side. The model figure of the active anti-vibration table using a pneumatic servo valve. The figure which modeled the conventional nozzle flapper type pneumatic servo valve. The front sectional view of the conventional spool type pneumatic servo valve.

First Embodiment In this embodiment, in addition to the following problems that cannot be achieved by the conventional valve, high-speed response, high resonance frequency, large exhaust flow rate, low air consumption flow rate, a great simplification of the structure is achieved at the same time. It proposes a new valve structure to be realized.

1 is a front sectional view of a pneumatic servo valve according to a first embodiment of the present invention, in which 50 is a central axis of a cylindrical portion, 51 is a bottom portion of the central axis, and 52 is a concentric circle with the axis of the central axis. The outer frame portion of the central shaft formed by (3), 53 is a coil bobbin mounted on the central shaft, and 54 is a coil wound on the coil bobbin. The central shaft 50, the central shaft bottom portion 51, the central shaft outer frame portion 52, the coil bobbin 53, and the coil 54 constitute a magnetic attraction electromagnetic actuator that attracts a flapper (described later) and controls its displacement. In the present embodiment, the present inventor uses a magnetic attraction type actuator proposed in Japanese Patent Application No. 2015-024794.

Reference numeral 55 denotes a cylindrical exhaust side housing that houses the center shaft bottom portion 51 and outer frame portion 52, 56 denotes this exhaust side housing bottom portion, 57 denotes a bolt for fastening the exhaust side housing bottom portion 56 and the center shaft bottom portion 51, and 58 denotes exhaust gas. The exhaust side flow passage formed in the bottom portion 56 of the side housing, and 59 are discharge ports formed in the central shaft 50. Reference numeral 60 is a supply side housing, 61 is a supply side flow path formed in the center of the supply side housing, and 62 is a control side flow path connected to a pneumatic actuator (not shown). Reference numeral 63 denotes a convex disk-shaped flapper, which is composed of a convex portion (magnetic path portion) 64a having a thick plate thickness and an outer peripheral portion (elastic deformation portion) 64b having a thin plate thickness. Incidentally, the flapper is generally called a flapper as shown in FIG. 48, which is a model of a conventional valve. In the present invention including this embodiment, a member that is arranged on the facing surface of the nozzle and adjusts the flow passage area of the fluid with the nozzle is called a flapper regardless of the shape of the member. Reference numeral 65 is an intake port, 66 is a supply side gap formed between the supply side housing 60 and the flapper 63, and 67 is an exhaust side gap formed between the flapper 63 and the exhaust side housing side. 68a, 68b, 68c and 68d are flow holes (68b and 68d are not shown) formed in the flapper 63, 69 is a supply side nozzle (forward nozzle), and 70 is an exhaust side nozzle (reverse direction nozzle). Reference numeral 71 denotes the flapper valve side end surface of the central shaft 50 (central shaft end surface), which is the first magnetic pole of the electromagnet. Reference numeral 72 denotes a second magnetic pole formed on the flapper-side end surface of the outer frame portion 52, and 73, the flapper is sandwiched between the supply-side housing and the exhaust-side housing to fasten the supply-side housing and the exhaust-side housing. It is a welded part. Reference numeral 74 is a supply-side convex portion formed at the supply-side central portion of the flapper, and 75 is an exhaust-side convex portion formed at the exhaust-side central portion of the flapper. Further, the supply side space and the exhaust side space form a control chamber 76 of the present valve.

By the way, by the magnetic flux generated by energizing the coil 54, a closed loop of “first magnetic pole 71→gap portion→flapper 63→gap portion→second magnetic pole 72→outer frame portion 52→center shaft bottom portion 51→center shaft 50” A magnetic circuit is formed. However, when the direction of the current flowing through the coil is opposite, the direction of the magnetic flux is opposite. In the present embodiment, a disk with a thin plate is used for the flapper, which is an element forming the closed loop magnetic circuit, and the magnetic saturation phenomenon of the magnetic flux flowing through the disk is used to obtain sufficient displacement characteristics with respect to the current value. Got linearity and big stroke. The magnetic attraction type actuator utilizing the magnetic saturation phenomenon is proposed by the present inventor in Japanese Patent Application No. 2015-024794.

As described above, the valve of this embodiment includes the forward nozzle 69 connected to the fluid supply side, the reverse nozzle 70 provided on the fluid discharge side, and the flapper 63 provided between the two nozzles. It is a bidirectional nozzle flapper valve constructed. The forward nozzle and the flapper adjust the nozzle opening (effective cross-sectional area) of the supply side flow passage, and the reverse nozzle and the flapper adjust the nozzle opening (effective cross-sectional area) of the exhaust side flow passage. There is. A portion of the supply side flow passage for adjusting the nozzle opening is referred to as a supply side flow rate control unit, and a portion of the exhaust side flow passage for adjusting the nozzle opening degree is referred to as an exhaust side flow rate control unit.

By the way, the valve of this embodiment can sufficiently reduce the air consumption flow rate in the steady state, although the large exhaust flow rate can be controlled. This is for the following reason.
(1) When one nozzle opening increases, the other nozzle opening becomes sufficiently small.
(2) Both nozzle openings do not increase at the same time, and the steady-state air consumption is suppressed at any position of the flapper.

An example of a structure that realizes the above (1) and (2) is an annular flow path that forms a flow path between the supply side nozzle 69 and the flapper 63, and between the exhaust side nozzle 70 and a substantially circular cross section. It is a formation structure. More specifically, the annular flow path forming structure has a cylindrical inner peripheral surface at the tip of each of the nozzles 69 and 70, and an insert body (74) which is inserted so as to be radially separated from the inner peripheral surface. And 75). That is, by using a convex portion that protrudes perpendicularly to the face plate portion of the flapper 63 as an insert body (micro piston), the length of the annular flow path in the axial direction can be adjusted by adjusting the insertion of the insert body into the nozzles 69 and 70. The flow rate characteristic can be changed by changing the flow rate characteristic. The valve structure described in this embodiment will be referred to as a "micro piston system".

2a to 2e are model diagrams showing the positional relationship between the nozzles 69 and 70 and the flapper 63 when the valve input current is changed. FIG. 4 shows the effective cross-sectional areas of the intake side and the exhaust side with respect to the drive current corresponding to the above model diagram. 2a shows a state where the valve input current I=0 (initial value), FIG. 2b shows a state where 0<I<I _max /2, FIG. 2c shows a state where I≈I _max /2 (neutral point), and FIG. 2d shows I. The state of _max /2<I _max , and FIG. 2e shows the state of I=I _max (maximum current value). In the figure, 77 is a supply-side nozzle orifice (supply-side annular flow path) formed by a narrow annular gap between the supply-side nozzle 69 and the supply-side convex portion, and 78 is between the exhaust-side nozzle 70 and the exhaust-side convex portion 75. Is an exhaust side nozzle orifice (exhaust side annular flow path) formed with a narrow annular gap. Reference numerals 79a, 79b and 79c are flow grooves formed on the flapper side end surface of the first magnetic pole 71, 80 is a supply side convex end surface, 81 is an exhaust side convex end surface, 82 is a supply side nozzle end surface, and 83 is an exhaust side nozzle end surface. is there.

When the valve drive current I=0 in FIG. 2A, the supply side convex portion 74 of the flapper deeply penetrates into the supply side nozzle orifice 77, and the supply side nozzle end surface 82 is in close contact with the flapper 63. At this time, the supply-side nozzle orifice (supply-side annular flow path) 77 is in a blocked state. Therefore, at this stage, there is no inflow of air from the supply-side flow passage 61 into the control chamber 76. Although details will be described later in [Supplement], when the valve driving current I=0 or I=I _max , the effective cross-sectional area A=0 (flow rate Q=0) of the present invention can be achieved. This is a characteristic peculiar to the micro piston type servo valve in the embodiment.

FIG. 2b shows the intermediate stage (0<I<I _max /2) of FIGS. 2a and 2c when the current is applied (I>0), and the flapper 63 moves away from the supply-side nozzle end surface 82. Attention is now paid to the dashed circle AA in FIG. 2b. At this stage, the flow rate is the "radial flow passage (gap δ _r )" (supply side annular flow passage 77) formed by the outer peripheral portion of the convex portion and the inner surface of the nozzle, and the "axial flow" formed by the flapper and the end surface of the nozzle. Path (gap δ _x )”. That is, the micro-piston valve also has the nozzle flow rate characteristic of the conventional nozzle flapper valve (Patent Document 3). Immediately after the flapper 63 separates from the supply-side nozzle end surface 82 (FIG. 2b), the fluid resistance of the axial flow path (gap δ _x ) is sufficiently large. Therefore, the flow rate (effective cross-sectional area) does not increase sharply but gradually increases.

At the drive current I≈I _max /2 (neutral point) in FIG. 2C, the supply-side convex end surface 80 is in a state of being close to the supply-side nozzle end surface 82. The exhaust side convex portion 75 is also in the same state as the exhaust side nozzle end surface 83. At this stage, the flow rate (effective cross-sectional area) flowing from the supply side nozzle into the control chamber and the flow rate (effective cross-sectional area) flowing from the control chamber into the exhaust side nozzle are also suppressed by the narrow gap δ _r .

FIG. 2d shows the intermediate stage (I _max /2<I<I _max ) between FIGS. 2c and 2e. As shown in the figure, the distance between the supply-side convex end surface 80 and the supply-side nozzle end surface 82 is X _d , and the supply-side nozzle inner diameter is d _n . The size of S ₁ and S ₂ above is defined as the flow passage area S ₁ =(d _n /2)π determined by the nozzle inner diameter d _n and the flow passage area S ₁ =2πd _n X _d determined by the virtual cylindrical surface at the nozzle opening end. To compare. In the state where the supply-side convex portion end surface 80 is sufficiently close to the supply-side nozzle end surface 82, S ₂ <<S ₁ , so that the effective cross-sectional area A of the supply-side flow path is A=S ₂ .

At the drive current I=I _max in FIG. 2e, the nozzle-side end surface 80 of the supply-side convex portion 74 is sufficiently separated from the supply-side nozzle end surface. In this case, the flow rate flowing from the supply side nozzle into the control chamber becomes maximum. Further, the exhaust side nozzle end surface 83 is in close contact with the flapper 63, and the exhaust side annular flow path 78 is in a blocked state. Therefore, in this state (FIG. 2e), there is no outflow of air from the control chamber 76 to the exhaust side flow passage 58.

By the way, the effective cross-sectional area of the flow path (nozzle part) in the pneumatic circuit can be measured by measuring the flow rate by the method defined in JIS B 8390-2000. For example, the supply side effective area shown in the graph of FIG. 4 is obtained by the following method. The discharge port 59 of the exhaust side flow passage 58 is shielded, a constant pressure source is connected to the supply side flow passage 61, and the control chamber 76 is opened to the atmosphere. In this state, the valve drive current I is changed and the air flow rate Q is measured. The function f (P _H , P _L , T) is calculated with the pressure of the pressure source as the upstream pressure _H , the downstream pressure P _L as the atmospheric pressure, and the temperature T [see equation (1)]. The relation between A and I may be obtained by setting the effective area A=Q/f (P _H , P _L , T) at the current I. Equation (1) is intended for compressible fluid with a relatively large flow passage area, but if the flow passage gap is on the order of several microns, it becomes a viscous fluid, so the equation (known) derived from the viscous fluid theory is applied. Then, the effective area A may be obtained.

As described above, the other embodiments are similar, but the effective sectional area A=0 (flow rate Q=0) when the valve drive current I=0 is unique to the micro piston type servo valve of the present invention. It is a feature. The reason for this is that the micro-piston type of the present embodiment has an annular "axial flow path" formed by the convex portion and the nozzle portion and a "radial flow path" formed by the flapper and the nozzle end surface for adjusting the flow rate. This is because two channels can be used. In the case of the spool type gas pressure control valve (FIG. 49) disclosed in Patent Document 4, the first plunge 507 and the second plunge 508 are slidably accommodated inside the cylindrical sleeve 505. Therefore, the obstruction of the flow path (reduction of the leak flow rate) must depend only on the narrow gap (axial flow path) between the outer peripheral portions of the first plunge 507 and the second plunge 508 and the sleeve 505. In order to keep the leak flow rate small, it is necessary that the narrow gap be in the viscous flow region (10 μm or less) of the fluid. In order to maintain this narrow gap, a non-contact floating is performed by the spool 501 with a static pressure bearing. There is. In Embodiments 1 and 3 of the present invention, the leak of the annular flow path is used as an effect of linearization, but the clearance of the annular flow path (the clearance between the micro piston and the nozzle) may be on the order of 0.1 mm. Compared to the spool type, the precision of parts can be greatly eased. The leak in the annular flow path in the above embodiment may be set in consideration of the balance with the air consumption flow rate that is allowed in a steady state.

In the present embodiment, when the valve drive current I=0, the flow path between the fluid supply source and the pneumatic actuator can be completely cut off, which can be utilized as a fail-safe function during a power failure. Due to this fail-safe function, even when a heavy load is mounted on the pneumatic actuator, the heavy load does not easily drop. This function is difficult in the conventional spool type in which the valve is constituted only by the radial passage (gap δ _r ).

FIG. 4 shows an example of the effective cross-sectional areas of the intake side and the exhaust side with respect to the drive current of the “micro-piston type valve” in the present embodiment. In the embodiment, A ₀ /A _max =0.15 is set, where A _{max is} the maximum value of the effective cross-sectional area on the supply side and A ₀ is the effective cross-sectional area of both at the neutral point. With the drive current I ₀ =0.025A at the neutral point, the intake side opening area gradually increases in the range of 0<I<I ₀ . In the range of I ₀ <I<I _max , the opening area on the supply side sharply increases.

FIG. 5 shows the characteristics of the conventional nozzle flapper valve (the opening areas on the intake side and the exhaust side with respect to the drive current) in which the effective cross-sectional areas of the intake side and the exhaust side at the operating point are set to be the same as those of the micro-piston method. Is. Unlike the above-mentioned micro-piston method, the effective cross-sectional areas of the intake side and the exhaust side with respect to the drive current have a linear proportional relationship in the entire current region (0<I<I _max ). The nozzle opening areas at the neutral points (I=I ₀ ) of both bars are set to the same value (A=0.25 m ² ).

Hereinafter, the pressure/flow rate characteristics of the servo valve of this embodiment when the effective cross-sectional areas of the intake side and the exhaust side with respect to the drive current are given will be determined in comparison with the conventional nozzle flapper valve. For the mass flow rate of the gas passing through the nozzle of the servo valve, the nozzle equations (1) and (2) in the isentropic flow of the compressive fluid are used. The effective cross-sectional area between the nozzle flappers is the annular flow path area formed between the nozzle tip and the flapper, where the nozzle inner diameter is d and the supply-side effective cross-sectional area a _in =dπx and the exhaust-side effective cross-section a _out =dπ. (δ _a -x). The mass flow rate G _in of the gas flowing into the air chamber from the supply source side is shown below. Here, P _s is the supply source pressure, P _a is the control chamber pressure of the servo valve, ρ _s is the supply source gas density, and κ is the specific heat ratio.

However, when P _a /P _s <{2/(κ+1)} ^2/(κ-1)

The mass flow rate G _out of the gas flowing out from the control chamber to the atmosphere side is expressed by the following equations (1) and (2): P _s →P _a , P _a →P ₀ , ρ _s →ρ _a , a _out =dπ( δ _a -x). V _c is the volume of the control chamber 33, and R is the gas constant. From the mass flow rates G _in and G _out , the pressure P _a of the control chamber 33 is obtained by the following equation.

FIG. 6 shows the analysis result of the control pressure in the steady state with respect to the current value in the pneumatic valve according to the first embodiment, in comparison with the conventional nozzle flapper valve. The control pressure is the pressure P _a of the control chamber 33. The analysis conditions are: supply pressure P _S =0.6MPa(abs), atmospheric pressure P ₀ =0.1MPa(abs), pneumatic actuator outer diameter D=Φ96mm, pneumatic actuator gap δ ₀ =3mm, pneumatic actuator The volume V _c =2.17×10 ^-5 m ³ .

FIG. 7 shows an analysis result of the air consumption flow rate in a steady state with respect to a current value in the pneumatic valve according to the first embodiment, in comparison with a conventional nozzle flapper valve. As described above, the nozzle opening area at the neutral point (I=I ₀ ) of both valves is set to the same value (A=0.25 m ² ), and the air consumption flow rate of both valves is almost the same level. ..

FIG. 8 shows the transient response characteristic of the valve flow rate with respect to time when the drive current of the example valve is suddenly changed. The purpose is to perform the feedforward control for suppressing the reaction force due to the stage movement, as described above. In the same figure, the transient response characteristics when the drive current is suddenly changed in the conventional flapper valve are shown in comparison. As shown in FIG. 6, the current value at the operating point of the embodiment valve is I=0.022A, and the conventional slip flapper valve is I=0.017A. At this time, the control pressures are both Pa=2.5×10 ⁵ Pa (0.25 MPa).

The example valve shows the case where I=0.022→0.040A, and the conventional slip flapper valve shows the case where I=0.017→0.040A. The valve of the embodiment can obtain a flow rate up to a maximum flow rate of Q=90 NL/min, whereas the conventional slipper flapper valve can only obtain a flow rate of up to a maximum flow rate of Q=33 NL/min. By the way, the air consumption flow rate in the steady state before the current is suddenly changed is in the range of Q=14 to 18 NL/min.

FIG. 9 shows the transient response characteristics of the generated load with respect to time when the drive current of the embodiment valve is suddenly changed, in comparison with the conventional slip flapper valve. FIG. 9b is a partially enlarged view of FIG. 9a. In FIG. 9, the time required for the generated load to reach 0.632 times the final reached value is defined as the time constant T. The time constant T _{1 of the} example valve (micro piston system) is 15.5 ms, and the time constant T ₂ of the conventional slip flapper valve is T ₂ =60 ms. Therefore, the valve of this embodiment can obtain a predetermined generated load in about 1/4 of the time as compared with the conventional slip flapper valve.

As described above with reference to FIGS. 1 and 2, the valve of the present embodiment has a schematic cross section between the supply side nozzle 69 and the flapper 63 and between the exhaust side nozzle 70 and the flapper 63. Each is provided with an annular flow path forming structure that forms an annular flow path. That is, in the annular flow path forming structure, the cylindrical inner peripheral surface of the tip end of each of the nozzles 69 and 70 and the insert bodies (74 and 75) inserted with a gap in the radial direction from the inner peripheral surface. ) And. The present invention has found that by appropriately setting the size of the above-mentioned gap, there is a controllable linear region in the control pressure characteristic with respect to the drive current while keeping the air consumption flow rate in the steady state at a low level. Is.

FIG. 10 shows the control pressure characteristics with respect to the drive current, with the ratio of the effective area A _{0 at} the neutral point and the maximum effective area A _max , that is, A ₀ /A _max as a parameter. Incidentally, the effective cross-sectional area A _{0 at} the neutral point is smaller as the radial gap of the annular flow path is smaller, and the maximum effective cross-sectional area A _max is larger as the nozzle inner diameter is larger. In FIG. 11, the air consumption flow rate with respect to the drive current is obtained under the same conditions as in FIG.

In FIG. 10, when A ₀ /A _max =0, as the current is increased, the control pressure Pa=0 in the range of 0<I<0.025A. At I=0.025A, the control pressure sharply increases to Pa=6.0×10 ⁵ Pa (constant value). Therefore, in this case, there is no linear region in which the control pressure is proportional to the drive current. However, from the graph of FIG. 11, the air consumption flow rate Q=0 in the entire current region.

As the value of A ₀ /A _max increases, the width of the linear region in which the control pressure is proportional to the drive current increases. At the same time, the air consumption flow rate also increases from the graph of FIG. When A ₀ /A _max =0.5, the characteristics are equivalent to those of the conventional nozzle flapper valve (Fig. 4). Therefore, by setting the range of 0<A ₀ /A _max <0.5, the control pressure has a linear region proportional to the drive current, and the air consumption flow rate can be made smaller than that of the conventional nozzle flapper valve. As shown in the equations (1) and (2) of the mass flow rate passing through the nozzle, the mass flow rate is proportional to the effective sectional area of the nozzle part. Therefore, if the leak flow rate of the annular flow path at I=I _max /2 is Q ₀ , and the maximum exhaust flow rate determined by A _max is Q _max , even if 0<Q ₀ /Q _max <0.5 is set. Good.

Now, in addition to the feature that the low air consumption flow rate and the large exhaust flow rate can be realized at the same time, the valve of this embodiment has the following effects.
(1) Resonance frequency can be set high (2) Valve can be driven with low power (3) High-speed response can be obtained

The reason for (1) above is as follows. In this research, we paid attention to the fact that when the member corresponding to the flapper of the valve is made into a thin disk shape, the effective mass m of the movable part that becomes the inertial load is only the elastically deformed part near the nozzle tip. Since the primary resonance frequency is proportional to (K/m) ^1/2 , the smaller the mass m and the larger the spring constant K, the higher the resonance frequency. In the conventional nozzle flapper valve (FIG. 50), a flapper 554, which is a rigid body having a mass of m, is supported by a spring 557, whereas in the servo valve of the present invention, the flapper itself is an elastic body (spring). Therefore, compared with the conventional nozzle flapper valve, the valve of this embodiment can set the resonance frequency higher. Further, in the spool valve (FIG. 51) disclosed in Patent Document 4, as described above, the resonance frequency of the movable portion including the spool 501 is equal to the mass m of the movable portion and the spring constants of the balancing springs 521 and 522. Determined by K. However, there is a limit to the weight reduction of the spool shaft that forms the large flow rate opening.

Now, in the case of an active vibration isolation system in which acceleration feedback is applied to the passive vibration isolation system, the peak of the resonance point can be reduced without deterioration of the vibration isolation performance in the high frequency range. Therefore, application of acceleration feedback control is essential for the active vibration isolation table (FIG. 49). The details are described in the previously proposed Japanese Patent Application No. 2015-024794, but when acceleration feedback control is applied, the open-loop characteristics of the active vibration isolation system show that the open-loop gain increases over a wide frequency range and the phase It becomes a delayed characteristic. Further, at the resonance point of the pneumatic servo valve incorporated in the closed loop control system, the open loop gain has a resonance peak and the phase is delayed by 180 degrees or more. As a result, there is a problem that the control system cannot have a sufficient margin for stability unless the resonance frequency of the pneumatic servo valve is set sufficiently high, for example, 200 Hz or higher. The valve of the present invention can easily solve this problem, which was difficult with the conventional spool type.

The reason for (2) above is as follows. The reason why the servo valve of the present invention can be driven with a small electric power (small electric current) is that the driving source uses Maxwell's stress acting on the conductor surface. Usually, a voice coil motor (linear motor) is used as an actuator for linearly moving a small displacement of the order of 0.1 mm to several mm. However, since the voice coil motor uses Lorentz force, a large thrust constant (electromechanical conversion efficiency) cannot be obtained. In this embodiment, paying attention to the magnetic saturation phenomenon of the magnetic flux flowing through the thin disk, it is possible to use the Maxwell stress having a thrust constant much higher than the Lorentz force. For the above reason, the capacity of the power source for driving the servo valve of this embodiment is sufficiently small, and a small current is sufficient.

The reason for (3) above is inevitably derived from the features of the valve of this embodiment of (1) and (2) above. That is, since the inertial load m and the spring load K are small and the electromechanical conversion efficiency is high, the number of turns of the coil is small and the inductance in the electric circuit is small. Therefore, the flapper displacement (flow rate) transfer characteristic with respect to the input current can have sufficiently high responsiveness.

Second Embodiment This embodiment proposes a valve structure in which a further lower air consumption flow rate can be achieved in a steady state and the linearity of the control pressure characteristic with respect to the current is improved as compared with the above-mentioned embodiments. That is, it is focused on that a mechanical feedback action can be obtained by asymmetrically forming the outer diameter of the insert (micro piston) formed on the front and back of the flapper.

(1) Configuration of Asymmetric Micro Piston Method FIG. 12 is a front sectional view of a pneumatic servo valve according to the second embodiment of the present invention. FIG. 13 shows a partially enlarged view of the micro piston part. Reference numeral 150 is a cylindrical central axis, 151 is a bottom portion of the central axis, 152 is an outer frame portion of the central axis formed concentrically with the central axis of the central axis, and 153 is a coil bobbin attached to the central axis. Reference numeral 154 is a coil wound around the coil bobbin. The central shaft 150, the central shaft bottom portion 151, the central shaft outer frame portion 152, the coil bobbin 153, and the coil 154 constitute an electromagnetic actuator that sucks a flapper (described later) and controls its displacement.

155 is a cylindrical exhaust side housing that houses the bottom 151 of the central shaft and the outer frame 152, 156 is this exhaust side housing bottom, 157 is a bolt that fastens the housing bottom 156 and the central shaft bottom 151, and 158 is the exhaust side housing. The exhaust side flow passage 159 formed in the bottom portion 156 is a discharge port formed in the central shaft 150. Reference numeral 160 is a supply-side housing, 161 is a supply-side flow path formed in the center of the supply-side housing, and 162 is a control-side flow path connected to a pneumatic actuator (not shown). A convex disk-shaped flapper 163 is composed of a convex portion (magnetic path portion) 164a having a thick plate thickness and an outer peripheral portion (elastic deformation portion) 164b having a thin plate thickness. Reference numeral 165 is a suction port, 166 is a supply side gap formed between the supply side housing 160 and the flapper 163, and 167 is an exhaust side gap formed between the flapper 163 and the exhaust side housing. 168a, 168b, 168c, 168d are flow holes (168b, 168d not shown) formed in the flapper 163, 169 is a supply side nozzle (forward nozzle), and 170 is an exhaust side nozzle (reverse direction nozzle). Reference numeral 171 denotes a flapper valve side end surface of the central shaft 150 (central shaft end surface), which is a first magnetic pole of the electromagnet. Reference numeral 172 is a second magnetic pole formed on the flapper side end surface of the outer frame portion 152, and 173 is a welding portion for fastening the supply side housing 160 and the exhaust side housing 155. Reference numeral 174 is a supply side convex portion (supply side micro piston) formed in the flap side supply side central portion, and 175 is an exhaust side convex portion (exhaust side micro piston) formed in the flap side exhaust side central portion. is there. As shown in FIG. 13, a Φd _in> Φd _out the outer diameter .PHI.d _in, the outer diameter of the exhaust-side protrusion 175 as .PHI.d _out of the supply-side protrusion 174. Further, the supply side void portion and the exhaust side void portion form a control chamber 176 of the present valve.

(2) Theoretical analysis of asymmetric micro-piston type valve The equilibrium condition of the force applied to the left and right disk surfaces is obtained. FIG. 14 shows an analytical model diagram of the asymmetric micro piston and the nozzle portion. FIG. 14a shows (i) a state in which a high-voltage source is not connected to the valve supply side, P _S =P ₀ (atmospheric pressure), and the drive current of the electromagnet is I=0, (ii) a high-voltage source on the valve supply side ( For example, P _S =0.6 MPa) is connected, the drive current is I=0, and any one of the above (i) and (ii) is shown. In any of the above cases (i) and (ii), the pressures in the control chamber 176 and the exhaust side flow path 158 are all atmospheric pressure P ₀ . FIG. 14b shows a case where the drive current is I>0 under the condition (ii) in which the high voltage source is connected to the valve supply side.

13 and 14b, the pressure P _a of the control chamber 176, the area of the supply side micropiston S ₁ , the area of the exhaust side micropiston S ₂ , the micropiston displacement x, the electromagnet drive current I, the thrust constant K _i, the spring constant of the disc K _d, the pressure feedback constant is K _P. The equilibrium condition of the force determined by the load applied to the end face of each piston, the attraction force of the electromagnet, and the reaction force of the disc spring is

If I=0, x=0 and P _a =P ₀

Here, if the bias load F ₀ given to the disk at the time of initial setting and the bias displacement amount x ₀ are set,

Substituting equation (7) into equation (4)

K _P is the area difference between the two micro pistons

By solving the above equations (1) to (3) and equation (8) simultaneously, the pressure/flow rate characteristics when using the asymmetric micro-piston type valve can be obtained. For example, when P _S =0.6Mpa (abs), P ₀ =0.1Mpa (abs), Φd _in =4.5mm, Φd _out =1.8mm, and K _d =19200N/m, the micropiston is configured as follows: Bias load F ₀ =3.14N and bias displacement x ₀ =0.164mm.

FIG. 15 is a block diagram when the drive current I of the electromagnet is input and the control pressure P _a is output. However, in the figure, the transfer function G(s) of the control pressure P _a with respect to the disc displacement x is obtained by solving the nonlinear equations (1) to (3) by numerical analysis, and P _a and x The relationship is non-linear.

FIG. 16 shows that, under the condition that the exhaust side convex portion diameter (exhaust side micropiston diameter) Φd _out =1.8 mm is constant, the supply side convex portion diameter (intake side micropiston diameter) Φd _in is used as a parameter for the drive current supply side. The effective area is shown. However, the graph assumes a no-load state in which there is no pressure difference between the supply side and the exhaust side, and assumes the case where the flapper 163 moves from the supply side to the exhaust side in proportion to the drive current value. Since it is assumed that the gap between the supply-side nozzle 169 and the supply-side convex portion 174 is small, the supply-side effective area A _in ≈0 in the range of current value 0<I<0.025A. When I>0.025A, the supply side effective area A _in increases linearly _in proportion to the current value. In the figure, the exhaust side effective sectional area (chain line) with respect to the drive current is shown for reference.

In this embodiment, the gap between the supply side nozzle 169 and the supply side convex portion 174 and the gap of the annular flow path formed between the exhaust side nozzle 170 and the exhaust side convex portion 175 are set to several μm to 10 μm. .. When the gap of the annular flow path is in this order, the flow of fluid (air) becomes a viscous flow region. On the other hand, the flow of the nozzle having an opening diameter of the order of 0.1 mm is a region of compressible fluid. Comparing the case where the same pressure difference is applied to the input side and the outlet side of the flow path, the flow rate in the viscous flow region is greatly reduced as compared with the compressive fluid region because of its high fluid resistance.

FIG. 17 shows the control pressure characteristics with respect to the drive current, based on the supply pressure P _S =0.6 MPa(abs), the atmospheric pressure P ₀ =0.1 MPa(abs), and the analysis conditions in Table 1. The outer diameters Φd _in and Φd _out of the micropiston are shown in FIG. When the current value is in the range of 0<I<0.025A, the control pressure Pa=0.1 MPa (atmospheric pressure) because the flow passage between the supply-side flow passage 161 and the control chamber 176 is blocked.

When Φd _in =1.8 mm, that is, when the intake-side micropiston diameter is equal to the exhaust-side micropiston diameter and Φd _in =Φd _in , the control pressure changes sharply around I=0.025A. That is, the atmospheric pressure P _a =P _{0 is} maintained in the range of I<0.025A, and the supply pressure P _a =P _S is maintained in the range of I>0.025A. Therefore, under this condition, proportional control of pressure by this valve is difficult. When Φd _in =2.7mm, the effect of mechanical/pressure feedback works in the range of 0.025<I<0.03A, and the control pressure is proportional to the current value. Furthermore, the linear region in which the control pressure is proportional to the current value increases as the intake-side micropiston diameter Φd _in increases.

FIG. 18 shows static characteristics of the control pressure with respect to the current value when Φd _in =4.5 mm and Φd _out =1.8 mm. This valve is an ON/OFF valve that has only two cases, an open state and a shutoff state, as shown in the graph. Nevertheless, the reason why the static characteristic of the control pressure with respect to the current value can have a proportional relationship will be described below.

19a to 19c show transient response characteristics of micropiston displacement, suction flow rate, and control pressure. That is, the comparison is made when the current value is changed to two target values of I=0→0.03A and I=0→0.04A at time t=0.4s. In the transient response characteristic of the micropiston displacement in FIG. 19a, both the current values I=0→0.03A and I=0→0.04A converge to the micropiston displacement x=0.35 mm. The reason is that, regardless of the final reached value of the control pressure, as will be described later with reference to FIG.

In the transient response characteristic of the intake flow rate in FIG. 19b, it can be seen that the state of the air consumption flow rate Q=0 is maintained at the stage of the current value I=0 and the steady state after the current reaches the target value. Further, in the transient response characteristic of the control pressure in FIG. 19c, at time t> 0.4 s, the pressure P _a reaches the static characteristic pressure (see FIG. 17) determined by the current value.

20a to 20c show the behavior of the asymmetric micropiston shown in FIG. 19a as a model. FIG. 20a shows the state before the current is applied, and the supply-side flow channel 161 and the control chamber 176 are in the cutoff state. FIG. 20b shows a state in which the space between the supply-side nozzle 169 and the supply-side convex portion 174 is open and air is filled in the control chamber 176 which is a closed space. At this time, the feedback effect is proportional to the area difference between the left and right micro piston _{[F B = (S1-S2} ) Pa ··· Equation (4) the second term on the right side of] is applied. This force F _B acts so as to move the supply side convex portion 174 to the supply side again. FIG. 20c shows a case where both the supply side nozzle 169 and the supply side convex portion 174 and the exhaust side nozzle 175 and the exhaust convex portion 175 are in a blocking state. At this time, the control chamber 176 becomes a closed space and converges to the control pressure (see FIGS. 18 and 19c) determined by the drive current.

Now, the case of applying the valve of the embodiment to the feedforward control of the active vibration isolation table will be considered below.
FIG. 21 shows the transient response characteristics of the generated load with respect to time when the drive current is suddenly changed. FIG. 21b is a partially enlarged view of FIG. 21a. In FIG. 21b, when the current I=0→0.04A is changed, the time constant at which the generated load reaches 0.632 times the final reached value (F=0→2700N) is T=13.6. is ms. Even when compared with the time constant T ₂ =60 ms of the conventional slip flapper valve described above, a sufficiently high response is obtained (see FIG. 9b). Therefore, it can be seen that sufficient performance required for feedforward control can be obtained even with the asymmetrical micro-piston system that can reduce the air consumption flow rate in the steady state to zero.

Third Embodiment This embodiment is based on the following two embodiments, that is, as shown in FIG.
(I) The outer diameter of the micro piston formed on the front and back of the flapper is made asymmetrical. (ii) The intake/exhaust flow passage area is changed in two steps. Due to the synergistic effect of the combination of (i) and (ii) above, the control pressure characteristic with respect to the current is controlled. We have found that the linearity of can be further improved.

FIG. 23 is a partially enlarged view of the micro-piston portion. 261 is a supply-side flow passage formed in the center of the supply-side housing, 262 is a supply-side void portion, and 263 is an exhaust-side void portion. Reference numeral 264 is a supply side nozzle (forward nozzle) and 265 is an exhaust side nozzle (reverse direction nozzle). 266 is a flapper, 267 is a central axis, and 268 is a flapper side end surface of the central axis, which is the first magnetic pole of the electromagnet. Reference numeral 269 is a supply side convex portion (intake side micro-piston) formed in the supply side central portion of the flapper, and 270 is an exhaust side convex portion (exhaust side micro piston) formed in the flap side exhaust side central portion. is there. Assuming that the outer diameter of the supply side convex portion 269 is Φd _in and the outer diameter of the exhaust side convex portion 270 is Φd _out , Φd _in >Φd _out as in the above-described embodiment. Reference numeral 271 is a control chamber formed by the supply side void portion and the exhaust side void portion. Focusing on the chain line circle AA, 272 is a supply side annular flow path formed by a gap between the supply side nozzle 264 and the supply side convex portion 269, and 273 is formed by an annular gap between the exhaust side nozzle 265 and the exhaust side convex portion 279. The exhaust side annular flow path (see FIG. 24c). Reference numeral 274 is a supply-side convex end surface, 275 is an exhaust-side convex end surface, 276 is a supply-side nozzle end surface, and 277 is an exhaust-side nozzle end surface.

The positional relationship between the flapper 266 and the two nozzles 264 and 265 in FIG. 23 shows the case where the drive current is in the range of 0<I<I _max /2. When the drive current is in the above range, the air supplied from the supply-side flow passage 261 leaks to the control chamber 271 via the supply-side annular flow passage 272.

24a to 24c show combinations of nozzle flapper combinations when the valve input current is changed. At the valve input current I=0 (initial value) in FIG. 24a, the supply side convex portion 269 of the flapper penetrates deeply into the supply side nozzle 264, and the supply side nozzle end surface 276 is in close contact with the flapper 266 and the supply side annular flow. Path 272 is blocked. Therefore, in this state (FIG. 24a), there is no inflow of air from the supply-side flow passage 261 into the control chamber 271.

When the current is applied (I>0), the flapper 266 separates from the supply side nozzle end surface 276 in the intermediate stage between FIGS. 24a and 24b (state of FIG. 23). However, the annular flow path formed by the supply side convex portion 269 and the supply side nozzle 264 has a narrow gap of the order of 0.1 mm in the embodiment, and flows from the supply side flow path 261 into the control chamber of the present valve. The flow rate of air is small. The exhaust side convex portion 270 is largely separated from the exhaust side nozzle 265, and the exhaust side is in an open state.

At the input current I≈I _max /2 (neutral point) in FIG. 24c, the supply-side convex end surface 274 is close to the supply-side nozzle end surface 276. The exhaust side convex portion 275 is also in the same state as the exhaust side nozzle end surface 277. The flow rate flowing from the supply side nozzle into the control chamber and the flow rate flowing from the control chamber into the exhaust side nozzle are also suppressed by the narrow gap.

At the input current I=I _max in FIG. 24c, the supply-side convex end surface 274 is sufficiently separated from the supply-side nozzle end surface 276. In this case, the flow rate flowing into the control chamber from the supply-side flow passage 261 is sufficiently large. Further, the exhaust side convex portion 270 deeply penetrates into the exhaust side nozzle orifice 265, the exhaust side nozzle end surface 277 is in close contact with the flapper 266, and the exhaust side nozzle orifice (exhaust side annular flow path) 273 is in a blocking state. .. Therefore, in this state (FIG. 24c), there is no outflow of air from the control chamber 271 to the exhaust side flow passage 278.

FIG. 25 shows the intake side effective sectional area with respect to the drive current in the present embodiment, in comparison with the second embodiment described above. In the present embodiment, A ₀ /A _max =0.15 is set as the maximum value of the effective cross-sectional area on the supply side and A _max , and the effective cross-sectional area A ₀ of both of them at the neutral point. However, in the present embodiment and the second embodiment, the exhaust side convex portion diameter (exhaust side micropiston diameter) Φd _out =1.8 mm, the intake side convex portion diameter (intake side micropiston diameter) Φd _in =4.5 mm Is set to. Further, the graph assumes a no-load state in which a pressure difference does not occur between the supply side and the exhaust side, and assumes the case where the flapper moves from the supply side to the exhaust side in proportion to the drive current value.

FIG. 26 shows the analysis result of the control pressure in the steady state with respect to the current value in the pneumatic servo valve according to the third embodiment, in comparison with the second embodiment. The analysis conditions such as the supply pressure, the outer diameter of the pneumatic actuator, the gap, and the volume are the same as those in the second embodiment. In the second embodiment described above, the control pressure and the characteristic with respect to the current have a linear region only when the drive current is in the range of 0.025<I<0.045A. In the third embodiment, the control pressure characteristic with respect to the current can have a linear region in a wide range of the entire current region 0<I<0.05A.

By the way, it is preferable that the servo valve characteristic in the active vibration isolation table (see FIG. 49) has a region in which the control pressure with respect to the current value maintains linearity in a wide range around the operating point. The reason is as follows. Since the servo valve is one element constituting the control system of the fluid servo device (active vibration isolation table), the ratio of the flow rate change to the current change is incorporated in the open loop gain as a flow rate gain. When the flow characteristic of the servo valve is non-linear, the open loop gain for predicting the stability margin of the entire active vibration isolation table must be determined by the maximum value of the flow gain. However, the operating point of the servo valve is usually used near the intermediate position of the drive current range (I≈Imax/2). Therefore, the more nonlinear the flow rate characteristic with respect to the current, the more excessive gain margin is set at the operating point. In this case, the active anti-vibration table cannot exhibit the sufficient performance that it originally has.
Further, the feedforward control (hereinafter referred to as FF control) is established only when the disturbance is known. A known stage behavior signal is used to perform the stage FF control. In order to effectively cancel the linear disturbance using the stage FF control, it is necessary to create a waveform of the generated force that faithfully reproduces the acceleration signal of the stage in antiphase. To do so, make the waveform of the valve drive current and the waveform of the generated pressure similar to each other, that is, make the region where the control pressure characteristic with respect to the current value is linear with the operating point of the valve drive current as wide as possible. It is preferable to have it in the range.

27a to 27c show transient response characteristics of the micro-piston displacement, the suction flow rate, and the control pressure in the pneumatic valve according to the third embodiment. That is, the comparison is made when the current value is changed to two target values of I=0→0.03A and I=0→0.04A at time t=0.4s. Hereinafter, description will be made in comparison with the case where there is a leak in the supply/exhaust flow path although there is a pressure feedback function (FIG. 19 of Embodiment 2).

When the current value is changed in two ways, I=0→0.03A and I=0→0.04A, the transient response characteristic of the piston displacement is that in the case of the second embodiment (FIG. 19a), the piston displacement x becomes a constant value. Converge. However, in the case of the third embodiment (Fig. 27a), the convergent value of the piston displacement x in each case is different.

In the transient response characteristic of the flow rate shown in FIG. 27b, it does not converge to zero in the case of the third embodiment. However, compared with the instantaneous maximum flow rate Q=250L/N (I=0→0.04A), the steady-state consumption flow rate can maintain a sufficiently small value Q=30-40L/N.

In the transient response characteristic of the generated pressure shown in FIG. 27c, compared with the second embodiment (FIG. 19c), the reached value of the pressure is higher when the current value I=0→0.03A. The reason for this is that the third embodiment has a wide range in which the “control pressure characteristic with respect to the current value” maintains linearity (see FIG. 26).

FIG. 28 shows the control pressure characteristics with respect to the drive current, with the ratio of the effective area A _{0 at} the neutral point and the maximum effective area A _max , that is, A ₀ /A _max as a parameter. Incidentally, the effective cross-sectional area A _{0 at} the neutral point is smaller as the radial gap of the annular flow path is smaller, and the maximum effective cross-sectional area A _max is larger as the nozzle inner diameter is larger. In FIG. 28, as A ₀ /A _max is decreased, the linear region where the control pressure is proportional to the drive current decreases. The graph of A ₀ /A _max =0 corresponds to the case where the radial gap of the annular flow path is small and the leak amount is small enough to be ignored, that is, the second embodiment. Here, the drive current I/I _max = 0.5 (I = 0.025 A) is set so that the neutral point is reached, and the control pressure at this neutral point is P _B , and P _max is the maximum control pressure (source Pressure), P ₀ is atmospheric pressure, and the intermediate pressure ratio β is

FIG. 29 shows the intermediate pressure ratio β with respect to η, where η=A ₀ /A _max . When η=0.2, β=0.5. At this time, the control pressure at the drive current I/I _max =0.5 is 1/2 of the maximum pressure, which shows an ideal linear characteristic capable of pressure control in the entire current range. When η=0, β=0. At this time, pressure control is possible only in the 1/2 current range.

In the graph of FIG. 29, the intersection point of the two envelopes of β with respect to η is designated as point C. η=η _c
The boundary of β changes greatly with respect to η. That is, the linearization effect brought about by the leakage of the annular flow path can be effectively utilized by setting η>η _c . In the case of this embodiment, η _c =0.03. As shown in the equations (1) and (2) of the mass flow rate passing through the nozzle, the mass flow rate is proportional to the opening area of the nozzle. Therefore, η=A ₀ /A _max =Q ₀ /Q _max . The Q _0, the Q _max Q ₀ / Q _max> may be set such that the eta _c.

In FIG. 30, the air consumption flow rate with respect to the drive current is obtained under the same conditions as in FIG. 28. In the figure, as A ₀ /A _max is reduced, the air consumption flow rate is reduced. For example, when A ₀ /A _max =0.15→0.03, the air consumption flow rate is reduced to Q=35→9NL/min.

Since the mass flow rate is proportional to the effective cross-sectional area of the nozzle portion, when A ₀ /A _max =0.03, the air consumption flow rate determined by the effective cross-sectional area A ₀ is only 3% of the maximum exhaust amount. Therefore, in the pneumatic servo valve according to the present embodiment, the linear region range of the control pressure characteristic can be significantly improved in a state where the air consumption flow rate in the steady state is maintained at a level that practically causes no hindrance.

FIG. 31 shows the control pressure characteristics with respect to the drive current, with the supply-side convex portion diameter Φd _in as a parameter under the conditions of the exhaust-side convex portion diameter Φd _out =1.8 mm and A ₀ /A _max =0.15. .. It can be seen that the larger the intake side convex portion diameter Φd _in , that is, the larger the piston area difference of the asymmetric micropiston, the wider the region where the “control pressure characteristic with respect to the current value” maintains linearity.

FIG. 32 shows the transient response characteristic of the generated load with respect to time when the drive current is suddenly changed. 32b is a partially enlarged view of FIG. 32a.

Fourth Embodiment All of the above-described embodiments of the servo valve of the present invention have a three-way valve structure. The basic structure of the servo valve of the present invention can also be applied as a four-way valve which has many applications in the hydraulic and pneumatic fields.

33 is a front sectional view of a pneumatic servo valve having the function of a four-way valve according to Embodiment 4 of the present invention, FIG. 34 is a view showing a state in which the valve of the present embodiment is connected to a pneumatic actuator, and FIGS. It is a model diagram which shows the positional relationship of the flapper and the 1st, 2nd flow path parts mentioned later. 750 is a first (right) central axis, 751 is a bottom portion of the first central axis, 752 is a first central axis outer frame portion provided concentrically with the axial center of the first central axis, 753 is a coil bobbin, 754 Is a coil wound around the coil bobbin. 755 is a cylindrical first housing for accommodating the first central shaft bottom portion and the outer frame portion 752, 756 is this first housing bottom portion, 757 is a bolt for fastening the first housing bottom portion and the first central shaft bottom portion, 758 is a first supply flow passage, and 759 is a first supply port. Reference numeral 760 is a second (left) housing, and 761 is a second central shaft mounted on the second housing. Reference numeral 762 denotes a central part of the second central axis, which is a second flow path part mounted on the flapper side, and 763 is a central part of the first central axis, which is a first flow path part mounted on the flapper side. Reference numeral 764 denotes the coil bobbin, the first central axis outer frame, a cylindrical portion made of a non-magnetic material mounted between the first central axes, and 765 a second supply port.

In FIG. 34, 766 is the valve of the present embodiment, 767 is the pneumatic actuator, 768 is the second air chamber of the pneumatic actuator, 769 is the first air chamber of the pneumatic actuator, 770 is the output shaft, and 771 is the valve of the present embodiment 766. And 772 is a second flow path connecting the second air chamber and 772 is a first flow path connecting the valve 766 of the present embodiment and the first air chamber.

In FIG. 33, 773 is a second control-side flow path connected to the second flow path 771, 774 is a first control-side flow path connected to the first flow path 772, 775 is a flapper, 776 is a second void portion, and 777 is a 1 void portion. Reference numeral 778 denotes the flapper side end surface of the first central shaft 750, which is the first magnetic pole of the electromagnet. Reference numeral 779 is a second magnetic pole formed on the flapper side end surface of the outer frame portion 752, and 780 is a welded portion that fastens the first housing 755 and the second housing 760 at the outer peripheral portion. In FIGS. 35a to 35c, 781 is a second convex portion formed on the second central axis side of the flapper, and 782 is a first convex portion formed on the first central axis side of the flapper. Further, an exhaust chamber 783 of the present valve is formed by the supply side void portion and the exhaust side void portion. Reference numeral 784 is an exhaust passage communicating with atmospheric pressure, 785 is a first supply nozzle that slidably fits the first convex portion, and 786 is a second supply nozzle that slidably fits the second convex portion. .. 787 is a flow passage formed in the outer peripheral portion of the first convex portion, 788 and 789 are radial passages of the first flow path part, and flow passages formed in the outer peripheral portion, 790 and 791 are the flow passages. The flow passage is formed in the radial direction of the first central axis and the outer peripheral portion. 792 is a flow passage formed in the outer peripheral portion of the second convex portion, 793 and 794 are radial passages of the second flow path part, and flow passages formed in the outer peripheral portion, 795 and 796 are the aforesaid It is a flow passage formed in the radial direction of the second central axis and in the outer peripheral portion. Reference numerals 797 and 798 are stoppers formed also on the flapper side of the first flow path part and the second flow path part and also serving as flow passages.

35a to 35c are model diagrams showing the positional relationship between the flapper, the first flow path part, and the second flow path part.

FIG. 34a shows a state where the flapper is in close contact with the second flow path part (driving current I=0), and the fluid flowing from the supply source side through the first supply flow passage 758→the first supply nozzle 785. Through the flow path 796→flow path 789→flow path 790→flow path 791→first control side flow path 774→first flow path 772 into the first air chamber 769 of the pneumatic actuator. On the other hand, the fluid in the second air chamber 768 of the pneumatic actuator passes through the second flow passage 771→the second control side flow passage 773, and then the flow passage 783→the flow passage 795→the flow passage 794→the flow passage 793→the flow passage 792. → It flows into the exhaust chamber 783 and is discharged to the atmosphere through the exhaust flow path 784. As a result, the pressure in the first air chamber 769 of the pneumatic actuator increases and the pressure in the second air chamber 768 decreases, so that the output shaft 770 of the pneumatic actuator moves to the left.

FIG. 35b shows a state where the flapper is in the middle of the first flow path part and the second flow path part (driving current I=I _max /2), and the first air chamber 769 of the pneumatic actuator, The two air chambers 768, the fluid supply source, and the flow path connecting to the atmospheric pressure are in the shielded state. As a result, the first air chamber 769 and the second air chamber 768 are closed.
FIG. 35c shows a state where the flapper is in close contact with the first flow path part (driving current I=I _max ). Contrary to the case of FIG. 35a, the fluid flows from the supply source side to the second side of the pneumatic actuator. The fluid in the first air chamber 769 flows into the air chamber 768 and is discharged to the atmosphere.

Contrary to the present embodiment, the supply source is connected to the space corresponding to the exhaust chamber 783, and the flow paths corresponding to the first supply flow passage 758 and the second supply flow passage 786 are connected to the exhaust side (atmospheric pressure). Even with such a configuration, the function as a four-way valve can be satisfied.

In the present embodiment, the magnetic attraction type is used as the means for driving the flapper, but a linear motor (voice coil motor) using Lorentz force may be used as described later. In this case, there is no restriction on the stroke of the flapper, and it is easy to configure a four-way valve with a larger displacement.

Fifth Embodiment In the above-described embodiment, the flapper is moved by the Maxwell attraction stress generated between the electromagnet and the flapper. In this embodiment, a linear motor (voice coil motor) is used as the flapper moving means.

FIG. 36 is a front sectional view of a pneumatic servo valve according to a fifth embodiment of the present invention, in which 350 is a central axis of a tubular portion, and 351 is an outer frame portion formed concentrically with the axis of the central axis. 352, a permanent magnet attached to the outer frame portion, 353, a flapper support member, 354, a cylindrical exhaust side housing for housing the bottom portion of the central shaft and the outer frame portion 351, 355, an exhaust side housing bottom portion, 356 Is a discharge port formed in the bottom of the exhaust side housing. 357 is a flapper, 358 is a bolt for fastening the flapper and the flapper support member 353, 359 is a coil bobbin attached to the flapper, and 360 is a coil wound around the coil bobbin. 361 is a bolt that fastens the flapper 357 and the coil bobbin 359, 362 is a flow hole formed in the flapper, 363 is an exhaust side nozzle fixed to the center shaft 350, and 364 is provided in the center part of the coil bobbin 359 on the flapper side. It is a convex portion on the exhaust side. Reference numerals 365 and 366 denote exhaust side flow passages formed in the exhaust side nozzle 363 and the central shaft 350.

The coil bobbin 359, the coil 360, the permanent magnet 352, the outer frame portion 351, the flapper 357, and the flapper support member 353 constitute a linear motor (voice coil motor) that controls the displacement of the flapper 357 by the current applied to the coil 360. There is. The principle of this linear motor utilizes the Lorentz force acting on the energizing coil 360 placed in a magnetic field.

367 is a supply side housing, 368 is a supply side flow path formed in the center of the supply side housing, and 369 is a control side flow path connected to a pneumatic actuator (not shown). Reference numeral 370 is an intake port, 371 is a supply side gap formed between the supply side housing 367 and the flapper 357, and 372 is an exhaust side gap formed between the flapper 357 and the exhaust side housing side wall surface. .. 373 is a supply side nozzle (forward nozzle), 374 is a supply side convex portion, 375 is a supply side flow passage formed in the supply side nozzle 373, and 376 is a bolt for fastening the supply side housing 367 and the exhaust side housing 354. .. A control chamber 377 of the present valve is formed by the supply side void portion and the exhaust side void portion.

In the present embodiment, the moving coil type in which the permanent magnet is arranged on the fixed side and the coil is arranged on the movable side is used, but a moving magnet type in which the permanent magnet is arranged on the movable side and the coil is arranged on the fixed side may be used.

In the above-described embodiment of the present invention, the case where the magnetic attraction type actuator is used as the driving unit that acts on the flapper has been described, but the linear motor of the present embodiment may be used. In the case of the linear motor type, since the mass of the movable portion is larger than that of the magnetic attraction type, the resonance frequency of the valve unit must be increased. However, the advantages of the linear motor type can be utilized in applications where a larger stroke is required by the flapper.

Sixth Embodiment The present embodiment relates to a pneumatic servo valve assembling method and a valve structure that enables this assembling method. Various features of the servo valve of the present invention are based on a unique valve structure different from the conventional spool valve (Patent Document 4). The practical problems shared by the valves of the above-described embodiments of the present invention are summarized as follows. Hereinafter, description will be made with reference to FIG. 1 (Embodiment 1) as an example.

(1) It is difficult to secure the coaxial accuracy of the two micropiston portions. Protrusions (74, 75) are formed on the front and back of the flapper 63, and these protrusions maintain a narrow gap, and each nozzle (69, 70). ). That is, the supply-side convex portion 74 is housed in the supply-side nozzle 69 to form a supply-side micro-piston portion, and the exhaust-side convex portion 75 is housed in the exhaust-side nozzle 70 to form an exhaust-side micro-piston portion. There is. The two micro-piston portions are composed of a combination of a plurality of members. For example, the exhaust side housing 55 that fixedly supports the flapper and the central shaft 50 that houses the exhaust side convex portion 75 are different members. Therefore, it is difficult to secure the coaxiality of the two micro piston parts only by the accuracy of each component alone.

(2) The flapper is an elastic body that is easily deformed. The flapper 63 is a disk-shaped elastic body that is easily deformed, and the protruding amount of each convex portion (74, 75) is small, for example, 0.5 to 1.0 mm, It is difficult to adjust the coaxiality during assembly.

First, a valve structure that enables the assembly method proposed in the present embodiment will be described below. FIG. 37 is a front sectional view of a pneumatic servo valve according to the fifth embodiment of the present invention, in which 650 is a central axis, 651 is a bottom portion of the central axis, and 652 is a concentric circle with the central axis of the central axis. A central axis outer frame portion, 653 are coil bobbins, and 654 is a coil wound around the coil bobbin. 655 is a cylindrical exhaust side housing that houses the bottom 651 of the central shaft and the outer frame 652, 656 is this exhaust side housing bottom, 657 is a bolt that fastens the exhaust side housing bottom 656 and the central shaft bottom 651, and 658 is the exhaust. The side flow passage 659 is a discharge port. Reference numeral 660 is a supply-side housing, and reference numeral 661 is a supply flow path component mounted on the supply-side housing.

Reference numeral 662 is a supply-side flow passage formed in the center of the supply flow passage component, and 663 is a radial gap (δ ₂ ) provided between the supply flow passage component and the supply-side housing 660. Reference numeral 664 is a bolt for fastening the supply flow path component to the supply side housing 660. Reference numeral 665 is a control side flow path connected to a pneumatic actuator (not shown), 666 is a flapper, 667 is a supply side void portion, and 668 is an exhaust side void portion. Reference numeral 669 is a supply side nozzle (forward nozzle), and 670 is an exhaust side nozzle (reverse direction nozzle). Reference numeral 671 denotes a flapper side end surface of the central shaft 650 (central shaft end surface), which is a first magnetic pole of the electromagnet. Reference numeral 672 is a second magnetic pole formed on the flapper side end surface of the outer frame portion 652, and 673 is a welded portion that fastens the supply side housing 660 and the exhaust side housing 655 at the outer peripheral portion.

Reference numeral 674 is a supply-side convex portion formed at the supply-side central portion of the flapper, and 675 is an exhaust-side convex portion formed at the exhaust-side central portion of the flapper. Further, the supply side void portion and the exhaust side void portion form a control chamber 676 of the present valve.
678 is a fitting portion of the supply side housing 660 and the exhaust side housing 655, 679 is a positioning pin for fixing the flapper 666 to the exhaust side housing 655, and 680 is a head accommodating portion of the positioning pin formed in the supply side housing 660. Is. Reference numeral 681 denotes a radial gap (δ ₁ ) provided between the central shaft outer frame portion 652 and the exhaust side housing 655. The central shaft bottom portion 651 and the central shaft outer frame portion 652 are fastened to each other at a joint portion 682 by adhesion, welding, or the like.

FIG. 38 shows a convex disk-shaped flapper 666 used in the final product, which is composed of a convex portion (magnetic path portion) 683a having a thick plate thickness and an outer peripheral portion (elastic deformation portion) 683b having a thin plate thickness. .. 684a, 684b, 684c and 684d are flow holes formed in the flapper 666 (684b and 684d are not shown), and 685 are hole portions into which positioning pins 679 are inserted, and a plurality of holes are formed on the circumference.

FIG. 39 shows a master disk 686 used as a substitute for the flapper 666 when assembling this servo valve. Compared with the flapper 666 which is easily elastically deformed, it is configured as a high-rigidity disk due to its sufficiently thick plate thickness.

Reference numeral 687 denotes a supply-side master protrusion, which has a slightly larger outer diameter and a longer length than the supply-side protrusion 674 of the flapper 666. Reference numeral 688 denotes an exhaust-side master convex portion, which similarly has a slightly larger outer diameter and a longer length than the exhaust-side convex portion 675 of the flapper 666. Reference numeral 689 denotes a master hole portion into which the positioning pin 679 is inserted, and a plurality of master hole portions are formed on the circumference at the same position as the hole portion 685 formed in the flapper 666. The assembly process of this embodiment will be described below.

Step 1 Adjustment of the central axis of the central shaft In FIG. 40, first, the central shaft 650 and the outer frame portion 652 integrated with the central shaft are housed in the exhaust side housing 655. Next, the master disk 686 is attached to the exhaust side housing 655. At this time,
(I) Insert the positioning pin 679 into the master hole 689 and the housing hole previously formed in the exhaust side housing 655.
(ii) Insert the exhaust side master protrusion 688 into the exhaust side nozzle 670.
(Iii) Due to the above (i) and (ii), the axial center of the central shaft 650 is within the range of the radial clearance (δ ₁ ) 681 with respect to the axial center C of the convex portions 687 and 688 as shown by the arrow AA. Move with to perform automatic alignment. In this state, the exhaust side housing bottom portion 656 and the central shaft bottom portion 651 are fastened with bolts 657.

Step 2 Shaft Core Adjustment of Supply Channel Component (i) In FIG. 41, the supply side housing 660 is inserted into the fitting portion 678 of the exhaust side housing 655 while maintaining the state of Step 1 above. Next, the supply flow path component 661 is attached to the supply side housing 660.

(Ii) The supply side nozzle 669 is inserted into the supply side master convex portion 687.
(Iii) Due to the above (i) and (ii), the axial center of the supply flow path component 661 has a radial gap (δ ₂ ) 663 with respect to the axial center C of the convex portions 687 and 688 as shown by an arrow BB. It moves within the range and is automatically aligned. In this state, the supply flow path component 661 and the supply side housing 660 are fastened with the bolt 664.

Step 3 Final Assembly of Product In FIG. 42, the supply side housing 660 and the master disk 686 are removed from the state of the above Step 2 and the flapper 666 is fixed to the exhaust side housing 655 by the positioning pin 679. Next, the supply-side housing 660 is attached to the exhaust-side housing 655, and the outer peripheral portions of both members 660 and 655 are welded, adhered, or fastened with bolts or the like.

By the assembly process of Step 1 to Step 3, the above-mentioned problems relating to the pneumatic servo valve of the present invention, (1) it is difficult to secure the coaxial accuracy of the two micropiston portions composed of a plurality of members, (2) the flapper is deformed An elastic body that is easy to do, above (1)
The problem (2) can be solved.

The two radial gaps described above, that is, the gap 681 (δ ₁ ) between the central shaft outer frame portion 652 and the exhaust side housing 655, and the radial gap provided between the supply flow path component 661 and the supply side housing 660. The size of 661 (δ ₂ ) may be determined according to the processing accuracy of each component forming the present valve. In the embodiment, if the radius gap is set to 0.05 mm or more, it is practically sufficient. Desirably, if it is set to 0.1 mm or more, a sufficient margin is maintained and the problems (1) and (2) are solved. We were able to.

Other embodiments will be described. It should be noted that this embodiment is not shown. For example, a fluid servo valve of the present invention, a gas spring that supports an object to be isolated from a foundation, an acceleration sensor that detects a vibration state of the object to be isolated from the vibration, and the information based on information from the acceleration sensor. By adjusting the fluid servo valve, the fluid servo valve is combined with an active controller that applies gas pressure to the gas spring to reduce the vibration of the vibration isolation target, and the fluid servo valve is configured to move the gas from the supply side to the gas spring. The fluid servo device may be configured so that the gas spring is supplied and exhausted from the gas spring to the exhaust side, and the primary natural vibration of the flapper is 200 Hz or more. With such a structure, excellent response characteristics and vibration isolation characteristics can be realized, and a sufficient vibration suppressing effect with respect to disturbance input can be obtained.

[Supplement]
(1) Utilization of Magnetic Saturation Phenomenon As described above, the reason why the valve of this embodiment can greatly reduce the air consumption flow rate is that the fitting state of the convex portions on both sides of the bidirectional flapper and the orifices on each nozzle side is This is because it can be adjusted by axial movement. For that purpose, it is preferable that the flapper is driven with a stroke as large as possible from the viewpoint of the structure and the processing of the member. However, in the case of an actuator using the Maxwell stress, the maximum value of the magnetic gap between the magnetic pole and the flapper that can effectively utilize the magnetic attraction is usually on the order of 0.05 to 0.20 mm. The characteristic of the magnetic attraction force with respect to the air gap is non-linear, and when it exceeds the above-mentioned maximum value, the magnetic attraction force usually decreases significantly. However, in the process of this research, the displacement characteristic of the flapper with respect to the current can be obtained with excellent linearity (linearity) by using a suitable magnetic material and a thin disk for the movable part corresponding to the flapper. I was able to find it. By more positively utilizing this magnetic saturation phenomenon, the stroke of the flapper can be significantly increased without losing the linearity of the displacement characteristic of the flapper with respect to the current. This result has already been proposed by the present inventor in Japanese Patent Application No. 2015-024794.

FIG. 43 shows an example of the displacement characteristics of the flapper (disk) with respect to the current value. In the graph, the specification Type(II) of the electromagnet and the disk shape used in the first to third embodiments is shown in comparison with another specification Type(I). The outer diameter of the Type(II) electromagnet is twice that of Type(I), and the number of coil turns is three times. When the current value I=40 mA, the flapper displacement X is about 0.12 mm in Type(I), whereas the flapper displacement X=0.68 mm is obtained in Type(II).

Here, a structure of a servo valve composed of element parts such as an electromagnet, a nozzle, and a flapper is assumed. At this time, the shape of each component, the overall configuration of the valve, etc. are arbitrary. Whether or not the magnetic saturation phenomenon is used in the servo valve of the present invention can be evaluated by the following method.

(I) Obtain the sum of the magnetic resistances of the closed loop magnetic circuit.
The magnetic resistance R _a between the nozzle flappers becomes minimum when the maximum current value I=I _max . When the distance between the nozzle flappers at this time is Δ _n and the magnetic pole area is S, R _a =Δ _n /(μ ₀ S). The sum of the linear magnetic resistances other than the magnetic resistance R _a is R _X , and the total magnetic resistance of the closed loop magnetic circuit is R _S =R _a +R _X. Linear reluctance refers to reluctance under the assumption that the magnetic permeability μ is constant and the relationship between the magnetizing force H and the magnetic flux density B is in a direct proportional relationship (B=μH).

(Ii) The maximum value of the magnetomotive force is E _max =N×I _max , where N is the number of turns of the electromagnetic coil for obtaining the maximum value of the magnetic flux generated in the closed loop magnetic circuit, and the maximum value of the magnetic flux is Φ _max =N×I _max. /R _S.

(Ii) Obtain the magnetic flux density B _max at a location where magnetic saturation is likely to occur.
In a closed-loop magnetic circuit, (1) a portion with the smallest magnetic path area, or (2) a portion using a magnetic material with the smallest saturation magnetic flux density, paying attention to the above (1) and (2), the magnetic path area the if S _c, a magnetic flux density _{B max = Φ max / S c} .

(Iv) Evaluation of Occurrence of Magnetic Saturation Phenomenon Here, the “magnetic flux density characteristic (BH characteristic) with respect to the magnetizing force” of the magnetic material used in the above (1) and (2) is used as evaluation data. The magnetic flux density boundary value B _c in the boundary region (magnetization force boundary value H _c ) between the linear region and the magnetic saturation region is compared with the magnitude of B _max . If B _max <B _c , the magnetic saturation phenomenon does not occur and the magnetic circuit is used in the linear region. If B _max >B _c , it can be seen that the magnetic saturation phenomenon occurs at the above-mentioned points (1) and (2), and the prerequisites for applying the magnetic saturation phenomenon are satisfied.

(2) Annular flow path forming structure In the above-described embodiment, an annular flow path forming structure is formed which forms a flow path having a substantially annular cross section between the supply side nozzle and the flapper and between the exhaust side nozzle. doing. More specifically, for example, in the case of the third embodiment, the annular flow path forming structure is configured such that the inner circumferential surface of the nozzle at the tip of each nozzle 264, 265 is separated from the inner circumferential surface in the radial direction. And an insert body (269 and 270) to be inserted as a result.

FIG. 44 shows a configuration in which the resonance frequency of the servo valve of the present invention can be set higher by reducing the weight of the flapper body. Incidentally, the resonance frequency of the magnetic attraction type servo valve applied to each of the embodiments of the present invention is determined by the resonance frequency f ₀ of the flapper which is a movable part. The flapper has a thin disk shape, and the movable portion is limited to the central portion. Therefore, the resonance frequency f ₀ can be made sufficiently higher than that of a valve driven by a conventional spool type (Patent Document 4) or moving coil type actuator. In each of the embodiments of the present invention, if a cavity is formed at the center of the insert (projection), the resonance frequency can be further improved. Reference numeral 701 is a supply side nozzle, 702 is an exhaust side nozzle, 703 is a flapper, 704 is a supply side convex portion, 705 is a supply side hollow portion formed in the center of the supply side convex portion, 706 is an exhaust side convex portion, 707. Is an exhaust side cavity formed in the center of the exhaust side convex portion. These cavities formed in each convex portion do not affect the supply-side and exhaust-side flow paths. In the case of the magnetic attraction type which is formed by the flapper thin disk and is deformed by magnetic force, the deformation amount is the largest in the central portion of the disk. Since the mass of the central part of the disk has a great influence on the resonance frequency of the disk, the reduction of the mass of the central part can greatly contribute to the improvement of the resonance frequency of the servo valve.

In the above-described embodiment, the convex portion is formed on the front and back surfaces of the flapper, and the cylindrical inner peripheral surface into which the convex portion is inserted is formed on the nozzle side. The reverse configuration can also provide the functions of the present invention. In FIG. 45, 711 is a supply side nozzle, 712 is an exhaust side nozzle, 713 is a flapper, 714 is a supply nozzle side convex portion, 715 is a supply side flow passage, 716 is an exhaust nozzle side convex portion, and 717 is an exhaust side flow passage. is there. Reference numeral 718 is a supply-side cylinder portion formed on the supply side of the flapper, and 719 is an exhaust-side cylinder portion formed on the exhaust side of the flapper.

In the above-described embodiment, the protrusions (inserts) formed on the front and back surfaces of the flapper have a cylindrical shape. In FIG. 46, the convex portion is formed in a tapered shape. 721 is a supply side nozzle, 722 is an exhaust side nozzle, 723 is a flapper, 724 is a supply nozzle side convex portion, 725 is a supply side flow passage, and 726 is a supply side. Tapered portion of convex portion, 728 is convex portion on exhaust nozzle side, 729 is exhaust side flow passage, 730 is tapered portion of convex portion on exhaust side, 731 is tip of supply side nozzle, 732 is control chamber, 733 is tip of exhaust side nozzle It is a department. 46a to 46c show the positional relationship between the nozzles 721 and 722 and the flapper 723 when the valve input current is changed. FIG. 46a shows the state where the valve input current I=0 (initial value), and FIG. 46b shows I≈I _max. A state of /2 (neutral point), and FIG. 46c shows a state of I=I _max (maximum value). Attention is now paid to the dashed circle D in FIG. 46a. The supply side nozzle tip portion 731 and the flapper 723 are in close contact with each other, and the radial direction flow path of the supply side nozzle tip portion 731 is in a blocked state. As a result, the flow path between the supply side flow passage 725 and the control chamber 732 is completely blocked. Also in FIG. 46c, the exhaust side nozzle tip portion 733 and the flapper 723 are in close contact with each other, and the radial direction flow path of the exhaust side nozzle tip portion 733 is blocked. As a result, the flow path between the control chamber 732 and the exhaust flow passage 729 is completely cut off. With this configuration, for example, as shown in the graph of the first embodiment (FIG. 4), the displacement of the flapper in the section where the effective area A=0 when I=0 and the current is 0<I<I _max /2. It is possible to make the intake side effective sectional area and the exhaust side effective sectional area relative to the displacement of the flapper linearly increase. Since the flow rate passing through the nozzle is proportional to the effective cross-sectional area, if the tapered portions 726 and 730 are formed in a curved shape, by selecting this curved shape, the section where the current is 0<I<I _max /2, Also, the intake/exhaust flow rate in the section of I _max /2<I<I _max can be set to an arbitrary curve. (Not shown)

Instead of forming the tapered portion on the convex portion, the inner surface of the nozzle may be tapered. Further, the convex portion may be always housed in the nozzle in the entire current region. Even in this case, the flow path may have a non-linear characteristic such that the effective area (flow rate) is convex downward with respect to the flapper displacement. (Not shown)

(3) Application of Surface-Drawing Nozzle The above-described embodiment has an annular flow passage forming structure that forms a flow passage having a substantially annular cross section between the supply side nozzle and the flapper and between the exhaust side nozzle. Had formed. In addition to this annular flow path forming structure, a non-contact type nozzle (surface throttle type nozzle) that utilizes viscous fluid resistance flowing in a narrow gap in the radial direction can also be applied.

In FIG. 47, 741 is a supply side nozzle, 742 is a flapper, 743 is a supply nozzle side convex portion, 744 is a supply side flow passage, 745 is a central axis, 746 is an exhaust side nozzle, and 747 is a central axis 745 on the flapper side. The formed protrusion has an outer diameter ΦD3 that is sufficiently larger than the exhaust nozzle diameter ΦD2. The gap δ between the convex portion 747 and the flapper 742 constitutes a surface-throttle nozzle. When the gap δ is small, the surface restriction nozzle has a very large fluid resistance because the fluid flow is in the viscous region. If converted into the above-mentioned effective area A of the flow path, A is very small. When the gap δ increases and the fluid flow enters the potential region, the fluid resistance is significantly reduced and the effective area A is significantly increased. That is, the flow rate changes with a non-linear characteristic with respect to the gap δ (displacement of flapper). The non-linear characteristic is a downwardly convex curve, and if the inflection point A0 is used, the valve of the present invention can be configured in combination with the annular flow path forming structure.

FIG. 48 shows a case where the above-described surface-throttle nozzle is applied to both the supply side and the exhaust side. 851 is a supply side central axis, 852 is a supply side nozzle, 853 is a supply side flow passage, 854 is a supply side convex portion, 855 is a flapper, 856 is an exhaust side central axis, 857 is an exhaust side nozzle, and 858 is an exhaust side flow passage. , 859 are exhaust side convex portions.

(4) Shape of flapper The shape of the flapper (disk) applied to the above-described embodiment of the present invention is, for example, in the case of the first embodiment (FIG. 1), the convex portion (magnetic path portion) 64a having a large plate thickness, It was a convex disk-shaped flapper constituted by an outer peripheral portion (elastically deformable portion) 64b having a small plate thickness. A spiral disc spring, a cloud spring, or the like may be used as the disc spring used for this elastically deforming portion.

If the plate thickness of the elastically deforming portion is made as thin as possible to reduce the rigidity, the stress generated in the elastically deforming portion increases as the plate thickness becomes thinner, and the allowable stress (elasticity limit) of the flapper member is exceeded. .. By using a spiral disc spring as the elastically deformable portion, the maximum generated stress can be significantly reduced. The rigidity and the generated stress of the spiral disc spring can be selected by the spiral angle α, the number of grooves (ridges), the width ratio of the grooves to the ridges, etc., in addition to the plate thickness (not shown).

(5) Overall Valve Structure In the above-described embodiment of the present invention, the valve structure is mainly composed of axially symmetric parts. In addition to the above-mentioned axially symmetric parts, even if a magnetic circuit and a fluid circuit are formed by combining various iron cores such as a prism, a cylinder, a horseshoe shape, and an annular shape, a rectangular thin plate material, and a rectangular block, the servo valve according to the present invention can be used. realizable. For example, magnetic material members may be connected in a substantially polygonal shape in cross-section to form a closed loop magnetic circuit, and each of the magnetic material members may be an iron core around which a coil of an electromagnet is wound, a yoke material, or a flapper (not shown). No).

In each of the above-described embodiments of the present invention, the case where air is used as the working fluid is shown, but the working fluid used in the present invention may be oil, air, or various kinds of gases. For example, a portion of the electromagnetic coil may be molded (sealed) with resin so that the liquid may come into contact therewith (not shown).

Magnetic materials such as permalloy (B), electromagnetic stainless steel, and pure iron can be used for the disk (flapper) and the parts that form the closed loop magnetic circuit. A non-magnetic material may be used for other housings. When a difficult-to-process material such as permalloy is used in order to achieve low hysteresis, each part should be a simplified shape element. These elements may be joined by selecting a construction method such as adhesion, shrink fitting, press fitting, and welding. For example, in the first embodiment (FIG. 1), the central shaft 50 may be formed into a thin cylindrical shape and the central shaft bottom portion 51 may be formed into a thin cylindrical shape, and these elements may be joined by the above-described method (not shown).

Although the case where the valve of the present invention is applied to the industrial active vibration isolator has been described above, the present invention can be applied to various pneumatic servo devices. For example, if it is applied to a power assist system that uses a pneumatic artificial muscle and is capable of cooperating with a person, it is possible to realize fine position control and a soft operation that an electromagnetic motor does not have. Alternatively, a walking support device that supports dorsiflexion of the ankle joint using pneumatic rubber artificial muscles, a power assist chair driven by a pneumatic cylinder, a fine movement stage driven by a pneumatic bellows, and a pneumatic manipulator that grips an object with a complicated shape. , Pneumatic brake systems for rail vehicles, active suspensions for vehicles, and so on.

The pneumatic servo system is 1. Clean, 2. Easy to maintain 3. Higher output/weight ratio compared to electric type 4. 4. Smooth movement due to compressibility 5. It has various features such as force control that are not found in other systems. It is no exaggeration to say that it is the servo valve, which is the heart of the system, that has the most control over the performance and cost of the pneumatic servo system. It is expected that widespread diffusion will be greatly accelerated.

(6) Micro-Piston Type and Component Accuracy As described above, the micro-piston type of the present invention uses two flow passages, an "axial flow passage" and a "radial flow passage", to obtain a valve drive current I. When =0, the supply flow rate Q _in =0, exhaust flow rate Q _out =Q _max , and when I=I _max , supply flow rate Q _in =Q _max , exhaust flow rate Q _out =0. Since the flow path is completely shielded, the value of the flow rate Q _max is not affected by leak (gap between members). The following synergistic effect can be obtained by combining the above characteristics with the pressure feedback action obtained by making the micropiston diameter asymmetric.

That is, the flow rate/pressure characteristic with respect to the current having excellent linearity can be obtained within the range of 0<I<I _max without being affected by the precision/assembly precision of each component forming the annular flow path forming structure. This effect is due to the linearization effect by pressure feedback that works in the range of 0<I<I _max . Therefore, the micro-piston type valve of the present invention can greatly reduce the processing/assembly accuracy while maintaining the valve performance. This effect cannot be obtained by a conventional spool valve or the like, which requires submicron level accuracy.

59 Fluid exhaust side 63 Flapper 65 Fluid supply side 69 Forward nozzle 70 Reverse direction nozzle 76 Control chamber 77 Supply side flow rate control section 78 Exhaust side flow rate control section

Claims (13)

A forward nozzle with a flow path connected to the fluid supply side,
A reverse nozzle with a flow path connected to the fluid exhaust side,
A flapper provided so as to face the tip portions of the forward direction nozzle and the reverse direction nozzle,
A flapper support member for fixing a part of the flapper,
Drive means for acting on the flapper such that the flapper moves between the forward nozzle and the reverse nozzle;
A supply side flow rate control unit formed between the forward direction nozzle and the facing surface of the flapper, or an exhaust side flow rate control unit formed between the reverse direction nozzle and the facing surface of the flapper,
The working fluid supplied from the fluid supply side passes through the forward direction nozzle, flows into a control chamber that is a space for accommodating the flapper, passes through the reverse direction nozzle from the control chamber, and passes through the fluid exhaust side. Is designed to flow out to
An effective cross-sectional area of the supply side flow rate control unit or the exhaust side flow rate control unit is configured to change with a non-linear characteristic with respect to the displacement of the flapper,
The non-linear characteristic is represented by a curved line or a polygonal line that is convex downward, the effective area at the bending point is A ₀ , when the maximum value of the effective area is A _max , 0 <A ₀ /A _max < A fluid servo valve characterized by being 0.5. The supply-side flow rate control unit, and the exhaust-side flow rate control unit is an annular flow path forming structure that forms a substantially annular flow path in cross section,
The annular flow path forming structure,
An inner peripheral surface of a tubular portion forming an outer boundary of the generally annular flow path,
The fluid servo valve according to claim 1, wherein the fluid servo valve comprises an insert body that is inserted in the inner peripheral surface so as to be spaced apart in the radial direction. The insertion part is formed by forming a convex portion at a substantially central position on the supply side and the exhaust side of the flapper, and the tubular portion is formed on the inner surfaces of the forward direction nozzle and the reverse direction nozzle. Item 2. The fluid servo valve according to Item 2. The insert .PHI.d _in the schematic outer diameter of the fluid supply side, as schematically outer diameter .PHI.d _ou fluid exhaust side, .PHI.d _in> fluid servo valve according to claim 2, characterized by being configured such that the .PHI.d _ou .. P _max is the maximum control chamber pressure, P ₀ is atmospheric pressure, the control chamber pressure at the neutral point is P _B , the effective cross-sectional area on the supply side is A ₀ , the maximum effective cross-sectional area on the supply side is A _max , η=A ₀ /A _max and the intermediate pressure ratio β = (P _B -P ₀ )/(P _max -P ₀ ), where η _c is the intersection of the envelope of the β curve with respect to η 5. The fluid servo valve according to claim 4, wherein A ₀ and A _max are set so that η>η _c . In the annular flow path forming structure composed of the tubular part and the insert, either the tubular part or a component forming the insert is attached to a housing that houses the annular flow path forming structure. The fluid servo valve according to claim 2, wherein the fluid servo valve is composed of a separate member, and is movable in a radial direction for adjusting an axis of the annular flow path in a state where a fastening member such as a bolt is removed. .. 7. The fluid servo valve according to claim 6, wherein the gap movable in the radial direction is 0.05 mm or more. 7. The annular flow path connecting from the fluid supply side to the fluid exhaust side is shielded by the close contact between the flapper and the supply side nozzle when the current applied to the drive means is zero. 2. The fluid servo valve according to 2. An electromagnet provided so as to generate an attraction force to the flapper, and the tip end portion of the forward nozzle and/or the tip end portion of the backward nozzle by deforming the flapper by the attraction force of the electromagnet. The fluid servo valve according to claim 1, wherein the drive means is configured so as to change a separation distance between the flapper and the flapper. A closed loop magnetic circuit is configured by the flapper, a magnetic pole that is provided opposite to the flapper and applies a suction action to the flapper, and a magnetic material component including the magnetic pole and the flapper.
The magnetic characteristics of this closed loop magnetic circuit are
A linear region in which the characteristics of the magnetic flux density with respect to the magnetizing force are roughly proportional,
The magnetic saturation region has a region where the inclination angle of the magnetic flux density characteristic with respect to the magnetizing force changes smaller than the linear region,
2. The fluid servo valve according to claim 1, wherein the magnetic flux density of the magnetic flux flowing through the magnetic material component is within the magnetic saturation region when the current passing through the electromagnet is increased within the displaceable range of the flapper. .. The drive means,
A coil bobbin fixed to the flapper,
An electromagnetic coil wound around this coil bobbin,
In the outer peripheral portion of the electromagnetic coil, a permanent magnet arranged with a gap is provided,
The fluid servo valve according to claim 1 or 9, wherein the coil bobbin, the electromagnetic coil, and the permanent magnet are configured as a linear actuator that utilizes a Lorentz force. A first nozzle,
A second nozzle,
A flapper that is disposed between the first nozzle and the second nozzle and that is provided so as to face the tip portions of the first nozzle and the second nozzle.
Drive means for acting on the flapper such that the flapper moves between the first nozzle and the second nozzle;
A first flow rate controller formed between the first nozzle and the flapper;
A second flow rate controller formed between the second nozzle and the flapper;
A first flow control channel having a first opening communicating with the outside, and a second flow control channel having a second opening communicating with the outside,
When the flapper approaches the first nozzle, the flow path is opened from the supply side to the first flow rate control flow path via the first flow rate control unit, and at the same time, the second flow rate control flow path is controlled from the second flow rate control flow path. The flow path is opened through the section,
When the flapper approaches the second nozzle, the flow path is opened from the supply side to the second flow rate control flow path via the second flow rate control unit, and at the same time, the first flow rate control flow path is controlled from the first flow rate control flow path. A fluid servo valve, characterized in that the flow path is opened through the portion. A fluid servo valve according to any one of claims 1 to 12,
A gas spring that supports the object to be isolated against the foundation,
An acceleration sensor for detecting the vibration state of the vibration isolation target,
By adjusting the fluid servo valve based on the information from the acceleration sensor, an active controller that applies a gas pressure to the gas spring to reduce the vibration of the vibration isolation object,
The fluid servo valve is configured to supply gas from the supply side to the gas spring and exhaust the gas from the gas spring to the exhaust side.
A fluid servo device in which the primary natural frequency of the flapper is 200 Hz or more.