JP7177492B2 - Fluid servo valve and fluid servo device - Google Patents

Description

The present invention relates to a fluid control device for controlling fluid pressure and flow rate. More specifically, the present invention relates to a fluid servo valve that controls fluid pressure and flow rate by relatively moving a flapper and a nozzle with an axial drive mechanism. It is about. Also, the present invention relates to 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 vibrations is spreading. Microfabrication/inspection devices such as scanning electron microscopes and semiconductor exposure devices (steppers) used in these processes require strict vibration tolerance conditions in order to guarantee the performance of the devices. In the future, along with the further integration and miniaturization of products, the processing speed will increase and the size of the equipment will increase, and the vibration tolerance conditions will tend to become more and more severe.

Conventionally, an active type precision vibration isolation table has been used in which an actuator is used to support a device susceptible to these vibrations and to control the actuator so as to dampen the vibration.

2. Types of Actuators There are several types of actuators used to control the active precision vibration isolation table, and the characteristics of each are summarized below. Linear motors (voice coil motors) generate a large amount of displacement, but have problems in that they generate a large amount of heat and generate a small amount of force. Piezo actuators are compact and have excellent responsiveness, but the generated displacement is as small as several tens of microns at most, and there are problems with long-term durability. The giant magnetostrictive actuator has excellent responsiveness, but the generated displacement is about twice as small as that of the piezo actuator.

On the other hand, pneumatic actuators are inferior in responsiveness to each of the actuators described above, but have the advantage of being able to easily increase the generated force by selecting the outer diameter of the piston and the pressure of the supply source. In addition, the actuator itself has an effect of isolating vibration from the floor surface (vibration isolation performance) due to the compressibility of air. In addition, by controlling the air spring pressure, it is possible to perform damping control of the direct-acting disturbance. That is, the pneumatic type actuator is characterized by the ability to have both "vibration isolation" and "vibration damping" functions, which actuators of other types do not have.

3. Conventional Example of Active Anti-vibration Table FIG. 68 shows a model diagram of a conventional active anti-vibration table using a pneumatic actuator. This active anti-vibration table is known as described in Patent Document 1 and Patent Document 2 as well. A plurality of sets of pneumatic actuators (582a, 582b) for supporting the surface plate 581 are arranged on the floor surface 580. As shown in FIG. A precision device (not shown) is mounted on the surface plate 581 . The pneumatic actuator consists of an air chamber 583 filled with high-pressure air for supporting a vertical load, and a piston 585 inserted through a diaphragm 584 above the air chamber. 586, 587a and 587b are acceleration sensors and displacement sensors for detecting vertical and horizontal acceleration of the surface plate 581 and relative displacement of the surface plate 581 with respect to the floor surface 580, respectively. An acceleration sensor 588 detects the acceleration of the floor surface 580 (the vibration state of the foundation). Output signals from these sensors are input to the controller 589 respectively. A servo valve 591 controlled by a controller 589 is connected to the air chamber 583 via a pipe 590 . The internal pressure Pa of the air chamber 583 is controlled by adjusting the flow rate of compressed air supplied to and exhausted from the air chamber 583 by a servo valve 591 which is a nozzle-flapper type electro-pneumatic converter. Here, the servo valve 591 has the function of supplying a gas having a supply pressure Ps from the outside, receiving a control signal from the controller 589, adjusting it to a desired gas pressure Pa and outputting it, and exhausting a part of it to the atmosphere P0. It is a control valve with Along with the trend toward larger devices supported by vibration isolation tables, air spring type vibration isolation tables that take advantage of pneumatic actuators have come to be widely used for micro-vibration control of ultra-precision equipment.

4. Conventional pneumatic valve (Part 1)
Nozzle flapper valves have been used in active vibration isolation tables, which are pneumatic servo devices using pneumatic actuators, to control the pressure of the actuators and adjust the flow rate. This nozzle flapper valve is a part mainly used as a primary control valve (pilot valve) of an electrohydraulic control valve. A nozzle flapper valve combines a torque motor and a pilot valve to rotationally drive an armature based on an input command signal. The torque generated at this time opens and closes the pilot valve by the nozzle flapper mechanism to supply the desired fluid pressure to the actuator. This pilot valve is called a nozzle flapper valve with a two-way flapper because it has a flapper that is displaced by the rotation of the armature and a pair of nozzles that are arranged facing each other on both sides of the flapper.

FIG. 69 shows an example of a nozzle flapper type conventional pneumatic servo valve, which is disclosed in Patent Document 3, and whose operating principle is generally widely used. Figure 69a is a front sectional view and Figure 69b is a side sectional view. Incidentally, this type of pneumatic servovalve used in the active vibration isolation table has been derived from the pilot valve of the hydraulic servovalve.

The servo valve 501 includes a body 502, a flapper displacing means 504 attached to the body with mounting bolts 503, and a flapper 506 displaced by energizing a coil 505, which will be described later.

The body 502 has a diametrical concave groove 508 formed on the surface on the side of the flapper displacement means 504, an input port P, an output port A and an exhaust port R for compressed air, and an output port A, which are opened on the opposite surface. groove 508, passages 510a and 510b, one end of which is open to the flapper insertion hole 509 and individually communicates the output port A with the input port P and the discharge port R; It has a pair of nozzles 511a and 511b assembled with their tips (nozzle portions) opposed to the paths 510a and 510b.

The flapper displacement means 504 comprises a magnet assembly 513 and the coil 505 assembled in the magnet assembly 513 . The magnet assembly 513 includes a pair of first yokes 514a and 514b that are arranged to face each other in the diametrical direction, and a pair of magnets 515a that are cylindrical in plan view and fixed to the first yokes at both ends in the circumferential direction. 515b (FIG. 69b) and a pair of second yokes 517 attached to the surfaces of the first yokes 514a and 514b opposite to the body 502 by means of mounting bolts 516 so that their tips face each other, and a coil 505 is wound thereon. The bobbin 505a is attached by fitting or the like between the first yokes 514a and 514b.

On the body 502 side of the first yokes 514a and 514b, projecting portions 518a and 518b are integrally formed so as to face the center in a direction orthogonal to the second yoke 517. , 518b (FIG. 69b) face each other with a flapper 506 interposed therebetween. Circumferential end surfaces of the magnets 515a and 515b (FIG. 69b) are fixed to the first yokes 514a and 514b by brazing or an adhesive, and the flapper displacement means 504 is formed by recesses of the first yokes 514a and 514b. It is attached to the body 502 by means of the mounting bolts 503 passing through the through holes 519 which are partially opened.

In the flapper 506, a flat surface of the tip 506a passing through the center hole of the bobbin 505a and the flapper insertion hole 509 faces the tips of the nozzles 511a and 511b, and the end 506b facing the second yoke 517 is made width across flats. there is The axially intermediate portion of the flapper 506 is fitted into the center hole of a leaf spring 520 whose both ends are each attached to the recessed groove 508 by two mounting screws 521 (Fig. 69b). are capable of being displaced in directions opposite to each other around the mounting portion by the plate spring 520. As shown in FIG.

The openings of the flow paths 510 a and 510 b are closed by plugs 522 respectively, and the flapper displacement means 504 is sealed from the body 502 by an O-ring 523 fitted on the outer peripheral surface of the flapper 506 . A cover 524 that covers the flapper displacing means 504 and is fitted to the body 502 is attached to the body 502 with a set screw 525 (FIG. 69b). 526 is a lead wire insertion member through which a lead wire (not shown) of the coil 505 passes.

In the servo valve 501, when the coil 505 is energized, the magnet assembly 513 causes the flapper 506 to swing about the supporting portion of the leaf spring 520, thereby causing the tip 506a to contact and separate from the tips of the nozzles 511a and 511b. is adjusted, the air pressure at the output port A increases or decreases in accordance with the amount of power supplied to the coil 505 .

That is, the servo valve 501 places the tip 506b of the flapper valve, which is the magnetic pole of the electromagnet, in the magnetic field generated by the pair of magnets (permanent magnets) 515a and 515b. By varying the current applied to the coil 505 of the electromagnet, the flapper displacement with respect to the input current is controlled.

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

1. Conditions Required for Pneumatic Servovalve The conditions required for the pneumatic servovalve, which is an important basic element constituting 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, several hundred Hz or more (3) Linearity: the flow rate and the generated pressure are linearly proportional to the valve driving current It is in

The reason for the above (1) is as follows. For example, when the stage (592 in FIG. 68) mounted on the anti-vibration table starts and stops, the drive reaction force due to the movement of mass is input to the surface plate, which is the stage installation surface, as linear motion disturbance. In this case, by applying stage feedforward control (hereinafter referred to as stage FF control) to the vibration isolator using the acceleration signal of the stage, it is possible to reduce surface plate vibration during acceleration/deceleration. High responsiveness is required for the air valve that drives the pneumatic actuator in order to quickly converge the surface plate vibration.

The reason for the above (2) is as follows. Although the responsiveness of the pneumatic active vibration isolation system is on the order of several Hz to tens of Hz, the reason why the servo valve requires a high resonance frequency of several hundred Hz is due to the unique needs of the pneumatic active vibration isolation system. It is based on In the case of a passive vibration isolation system composed only of a spring, damping, and mass, the peak at the resonance point can be reduced by increasing the damping at the resonance point determined by the spring constant and damping coefficient of the actuator. However, in the high frequency range above the resonance point, the damping performance is degraded.

In the case of the active vibration isolation system in which acceleration feedback is applied to the above passive vibration isolation system, as shown in FIG. peak can be reduced. Therefore, application of acceleration feedback control is essential for active vibration isolation table. However, when acceleration feedback control is performed, as will be described later in Supplement (2), the open-loop characteristics of the active vibration isolation system have an increased open-loop gain over a wide frequency range and lag in phase. Furthermore, 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 lags by 180 degrees or more. As a result, unless the resonance frequency of the pneumatic servo valve is set to a sufficiently high value, for example, 200 Hz or higher, the control system cannot obtain a sufficient margin for stability.

The reason for the above (3) is as follows.
Since the servo valve is one element that constitutes the control system of the fluid servo device (active anti-vibration table), the ratio of the change in flow rate to the change in current is incorporated in the open loop gain as a flow rate gain. When the flow rate characteristic of the servo valve is non-linear, the open loop gain for considering the stability margin of the entire active vibration isolator must be determined by the maximum value of the flow rate gain. However, the operating point of the servo valve is usually used near the middle position (I≈Imax/2) of the drive current range in many cases. Therefore, the more nonlinear the flow rate characteristic with respect to the current, the more excessive the gain margin is set at the operating point. In this case, the active anti-vibration table cannot exhibit its original sufficient performance.

Furthermore, 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 motion disturbance using the stage FF control, it is necessary to create a generated force waveform that faithfully reproduces the acceleration signal of the stage in opposite phase. For this purpose, the waveform of the valve driving current and the generated pressure should be similar, that is, the area where the characteristics of the generated pressure (generated force) with respect to the current value maintain linearity around the operating point of the valve driving current. should be as wide as possible.

2. Problems of Conventional Pneumatic Servovalve As an example of configuring an active vibration isolation table, four-point support active control is assumed. In this case, the pneumatic actuators are arranged at the four corners, and the installation direction of the units is such that two points in the horizontal X direction and two points in the Y direction are diagonally arranged. Each actuator also incorporates an actuator that supports loads in the Z direction. Therefore, a total of 8 pneumatic actuators are arranged and a total of 8 pneumatic servo valves are required to control each actuator.

A conventional pneumatic servo valve that requires the above (1) to (3) as necessary conditions requires a large number of high-precision parts, as shown in FIG. In addition, due to the three-dimensional arrangement of members, there is a large accumulated error in the nozzle flapper portion, which requires high accuracy, and it is difficult to obtain uniform performance. Moreover, when the above valves are mounted on a multi-axis control active vibration isolation table, there is a problem that the cost ratio of the vibration isolation table is high due to the large number of necessary valves.

Therefore, the basic proposition given to this research was how to discover the possibility of a "new-principle servovalve" that can greatly simplify the complicated structure of conventional pneumatic servovalves. To do so, it is necessary to analyze the historical background of pneumatic servovalve technology currently in use. The basic form of the current pneumatic servovalve is derived from the application of the hydraulic servovalve technology that has a long history.

FIG. 70 shows a structural diagram modeling the operating principle common to conventional servo valves. The configuration of the servo valve can be broadly divided into an actuator section A-1 and a fluid control section B-2. In the actuator part 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 attached with their tips facing each other, 556 is This is the tip of the flapper on the actuator side. Reference numeral 557 denotes a leaf spring that also serves as a sealing member, and 558 denotes a support central portion of the leaf spring. However, in the structure of the servo valve actually used, as shown in FIG. 69, the yoke members for forming each magnetic circuit of the permanent magnet and the electromagnet are arranged orthogonally in the circumferential direction. It has a three-dimensional structure that is

Here, attention is focused on the magnetic pole portion A-2 (dotted line circle) in the vicinity of the flapper tip portion 556 . 559a is the gap a between the tip of the flapper and the yoke 555a, and 559b is the gap b between the tip of the flapper and the yoke 555b. The dashed line Φ1 in the figure is the magnetic flux generated by the electromagnet between the tip of the flapper and the yokes, and the two-dot chain line Φ2 is the magnetic flux generated by the permanent magnet 551 formed between the yokes 555a and 555b. Here, if the magnetic flux Φ in the air gap between the magnetic poles, the cross-sectional area of the magnetic poles is S, and the magnetic permeability of the air is μ0, the generated Maxwell's total stress T is

したがって、フラッパ先端部に加わる力Fは、左右の磁極に発生する全応力（=応力×磁路面積）の差T1 -T2に比例する。

Therefore, the force F applied to the tip of the flapper is proportional to the difference T1-T2 between the total stresses (=stress x magnetic path area) generated in the left and right magnetic poles.

フラッパの変位を制御するために、電磁石が作る磁束Φ1を電流により可変させたときフラッパ先端部に加わる力Fは、式(2)に示すように、永久磁石が作る磁束Φ2によりアシスト（増強）されるのである。

In order to control the displacement of the flapper, the force F applied to the tip of the flapper when the magnetic flux Φ1 created by the electromagnet is varied by current is assisted (enhanced) by the magnetic flux Φ2 created by the permanent magnet, as shown in Equation (2). It is done.

In the fluid control unit B-1, 560 is a forward nozzle, 561 is a reverse 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 room. A gas having a supply pressure P S is supplied to the forward nozzle 560 and flows into the control chamber 566 at a rate proportional to the distance between the flapper 562 and the forward nozzle. On the other hand, the gas in the control chamber 566 flows out to the atmosphere through the exhaust port 564 at a flow rate proportional to the distance between the flapper 562 and the backward nozzle. The difference between the inflow from the forward nozzle 560 and the outflow from the reverse nozzle 561 determines the control pressure Pa in the control chamber 566 and the outflow from the load port 565 .

In the hydraulic servo, the reason why the actuator part A-1 and the fluid control part B-2 of the servo valve are separated is that the electromagnet that carries the current could not be immersed in oil, which is a conductor. The reason for using both a permanent magnet and an electromagnet is that in the case of a hydraulic servo, the force applied to the flapper surface by the force of the oil ejected from the nozzle is proportional to the density of the fluid. Big difference. In order to resist this force, the flapper needed a large driving force.

On the other hand, the reason why a permanent magnet is required even in this type of servo valve applied to the pneumatic active vibration isolation table is as follows. Assuming that the equivalent mass of the flapper 554 is m and the equivalent spring stiffness of the plate spring 557 is K, the resonance frequency f0 is proportional to the following expression (7). When constructing a servo system, it is necessary to set the resonance frequency f0 sufficiently high for the reason (2) above. Since there is a structural limit to reducing the mass m of the flapper 554, the spring stiffness K must be increased in order to increase the resonance frequency f0. As a result, it is necessary to increase the driving force F for driving the flapper 554, and it is necessary to use a permanent magnet to assist the driving force of the electromagnet.

Compared to hydraulic servos, which have a wide range of applications, pneumatic servos were a minor presence at first. When the need for pneumatic servos emerged due to the emergence of active vibration isolation tables as a demand of the times, the adoption of the basic structure of conventional servo valves (Fig. 69) cultivated through hydraulic servo technology was historically an inevitable choice. It is thought that

2-2. Examples of conventional proposals for pneumatic servo valves
While the above-described servo valve embodiment utilizes the magnetic attraction effect of a combination of a permanent magnet and an electromagnet, it utilizes the Lorentz force (the principle of a linear motor) acting on a current-carrying coil placed in a magnetic field. Patent document 4 proposes a servo valve for adjusting a flapper valve.

In the servo valve shown in FIG. 71, 601 is a forward flapper, 602 is a forward nozzle, 603 is a reverse flapper, 604 is a reverse nozzle, 605 is a supply port, 606 is an exhaust port, and 607 is a load port (control port). is. Gas having a supply pressure Ps from the supply port 605 is supplied to the forward nozzle 602 , and the gas ejected from the forward direction nozzle 602 according to the distance from the forward direction flapper 601 is discharged to the exhaust port 606 at the tip of the reverse direction nozzle 604 . , and the load ahead of the load port 607 are supplied in parallel. Therefore, the reverse nozzle 604 has a function of sucking gas, and is a so-called suction nozzle. The amount of gas sucked in the reverse direction nozzle 604 changes according to the distance from the reverse direction flapper 603 . In this way, by moving the linear motor, the ejection of gas from the forward direction nozzle 602 and the intake of gas into the reverse direction nozzle 604 are synchronously controlled, so that the gas pressure Pa can be adjusted.

The linear motor 608 has a magnet 610 fixed to a housing 609 , a moving body 611 , and a coil 612 installed on the moving body 611 . The connection terminal 614 is connected to a control unit (not shown) by a control cable or the like. Both ends of the moving body 611 are supported by the housing 609 via cloud springs 615 and 616 . The cloud spring is a leaf spring having, for example, cloud-shaped slits in a thin plate, and an object is connected to the center of the leaf spring, and the periphery is used as a fixed end to restrict the rotation of the object around the axis and prevent the movement in the axial direction. It functions as a radial support spring that enables and supports an object.

The flapper valve proposed above has a problem in that the above requirement (2) required for the active control anti-vibration table, that is, "the primary resonance point of the air valve is sufficiently high". Since the moving part of the linear motor consists of the forward flapper 601, the reverse flapper 603, the coil 612, and the moving body 611, the mass m of the moving part must be larger than that of the conventional pneumatic servo valve (Fig. 69, see Table 1). I can't help it. In order to obtain a high resonance point, it is necessary to increase the rigidity of the springs 615, 616. In addition, it is necessary to increase the force generated by the linear motor that drives against the high-rigidity spring. However, in the case of a linear motor using the Lorentz force, the electromechanical conversion efficiency of the generated force F with respect to the input current I is small, and a large generated force cannot be obtained. Therefore, the stiffness K of the springs 615, 616 must be reduced. As described above, since the resonance frequency is proportional to , the resonance point of the air valve cannot be made sufficiently high, and the above requirement (2) required for the active control anti-vibration table cannot be satisfied.

Specifically, the present invention includes a nozzle having a flow path communicating with a fluid supply source, a flapper provided so as to face the tip of the nozzle, and a flapper support member fixing a part of the flapper. and an electromagnet provided to generate an attraction force with respect to the flapper, the flapper being deformed by the attraction force of the electromagnet to change the separation distance between the tip of the nozzle and the flapper. It is configured as follows.

That is, in the present invention, unlike the conventional servo valve, which has a rigid flapper structure that swings around a fulcrum, the flapper itself is elastically deformed by the attractive force of the electromagnet, thereby separating the nozzle from the flapper. By varying the distance, the fluid pressure and flow rate are controlled.

Specifically, in the present invention , the fluid supply source supplies air, the electromagnet, the flapper, and the flapper fixing member constitute an actuator section, and the fluid passing through the nozzle is It is constructed so as to pass through a space formed by each wall surface of each member constituting the actuator section.

That is, in the present invention, the wall surfaces of the members that make up the actuator section are separated from the actuator section and the fluid control section in the conventional air servo valve derived from hydraulic servo technology. It is used as a flow passage.

Specifically, the present invention includes a nozzle having a flow path connected to a fluid supply source and configured such that the inner diameter of the tip portion is tapered, and a flapper provided so as to face the tip portion of the nozzle. , a flapper support member that fixedly supports a part of the flapper, an electromagnet provided to generate an attractive force with respect to the face plate portion of the flapper, and a closed loop configured to include at least the electromagnet and the flapper a magnetic circuit, wherein the flapper is deformed by the attractive force of the electromagnet to change the separation distance between the tip of the nozzle and the flapper, and the flapper and the electromagnet When the separation distance of becomes equal to or less than a predetermined value, the magnetic flux flowing through the magnetic material parts constituting the closed loop magnetic circuit is magnetically saturated, and the flapper's response to the current applied to the electromagnet is compared with the case where the magnetic material component is not magnetically saturated. In addition to reducing the rate of change in displacement, the attraction force of the electromagnet acting on the flapper and the restoring force due to the elastic deformation of the flapper in the vicinity of the center point of the flapper's movable range allow the valve operating point during control to reach a predetermined current value. It is configured to be a predetermined flapper position with respect to .

That is, in the present invention, a portion having a narrow magnetic path area or a portion having a high magnetic resistance is provided in the closed-loop magnetic circuit so that the magnetic flux density characteristic with respect to the current value (magnetizing force) should normally rise sharply. Set the magnetic saturation region. As a result, the flapper displacement (flow rate) characteristic with respect to the current value has excellent linearity and controllability.

Specifically, in the magnetic flux density characteristic with respect to the magnetizing force of the magnetic material part, the range of 0<H<Hc in which the magnetic flux density B increases in proportion to the magnetizing force H is defined as a linear region, and the magnetizing force H The range of H>Hc in which the gradient of the magnetic flux density B drops significantly with respect to the It is a magnetic force boundary value obtained from the intersection of the envelope of the BH characteristic in the magnetic saturation region, the magnetic flux density boundary value in the boundary region between the linear region and the magnetic saturation region is Bc, and the current value that energizes the electromagnet is the maximum. Let RS be the total linear magnetic resistance of the closed loop magnetic circuit at the value Imax, N be the number of turns of the electromagnet, and Φmax=N×Imax/RS be the magnetic flux, Sc be the magnetic path area of the magnetic material part, and Bmax be the magnetic flux density. It is configured such that Bmax>Bc when Φmax/Sc.

That is, in the present invention, by utilizing the magnetic saturation phenomenon of magnetic material parts, it is possible to obtain characteristics with excellent linearity (controllability). is magnetically saturated within the operating range of the valve is a prerequisite for applying the present invention. That is, if Bmax<Bc, the magnetic saturation phenomenon does not occur and the magnetic circuit is used within the linear region. If Bmax>Bc, the magnetic saturation phenomenon occurs at the points (1) and (2) above, which means that the preconditions for applying the present invention are satisfied.

Specifically, in the present invention, in the vicinity of the maximum value of the current that is passed through the electromagnet, the flow rate characteristic with respect to the current is configured to form an upwardly convex curve.

That is, in the present invention, by utilizing the magnetic saturation phenomenon in the region where the displacement (flow rate) characteristic of the flapper rises sharply as the current value increases, a characteristic with excellent linearity (controllability) is obtained. be able to. Therefore, the flow rate does not sharply increase in the vicinity of the maximum value of the input current, and the characteristics are suppressed with good controllability.

Specifically, in the present invention, the flapper is composed of a substantially flat plate-shaped member, and the elasticity of the flapper itself is used to give the flapper a restoring force proportional to the size of the gap between the nozzle and the flapper. It is a thing.
That is, in the present invention, if the member corresponding to the flapper of the valve is formed into a thin disc shape, the effective mass of the movable portion acting as the inertial load is only the elastic deformation portion near the tip of the nozzle. In the conventional servo valve, the flapper, which is a rigid body, is supported by a spring, whereas in the servo valve of the present invention, the flapper itself is an elastic body (spring). Therefore, the effective mass of the movable portion can be reduced, and the resonance frequency can be increased. Moreover, since the stiffness of the spring that supports the flapper can be made sufficiently small, the flapper can be driven only by the electromagnet, although the valve of this embodiment has a high resonance frequency.

Concretely, according to the present invention , the flapper is composed of the magnetic material part.

That is, in the present invention, 1. In order to improve the responsiveness of the valve flow rate to the current, it is necessary to reduce the weight of the flapper, which is the movable part, that is, to reduce the plate thickness of the flapper. 2. By reducing the plate thickness of the flapper, the magnetic path area through which the magnetic flux flows becomes smaller, and the magnetic saturation phenomenon can be utilized. 1.2 above. It utilizes the fact that the purposes of Further, even if the plate thickness of the flapper, which is the movable portion, is reduced, the strength of the valve body as a structural body is not affected.

Specifically, in the present invention , the electromagnet includes a first magnetic pole formed on an inner end face facing the flapper and a second magnetic pole formed on an outer end face facing the flapper, and the flapper is supported by the magnetic path part formed between the first magnetic pole and the second magnetic pole in the magnetic closed loop circuit formed by the electromagnet, and the flapper supporting member, and the magnetic path part is elastically and an elastic support portion for positively supporting the magnetic path, and the magnetic path portion and the elastic support portion are configured to have different flexural rigidity.

That is, in the present invention, the disk corresponding to the flapper is formed, for example, in the shape of a convex disk so that the portion (magnetic path portion) utilizing the magnetic attraction acting on the disk is located between the first magnetic pole and the second magnetic pole. and a portion (elastic support portion) for setting the spring rigidity of the disk is formed on the fixed side of the disk. By forming the disk into the above shape, for example, it is possible to break the contradictory relationship that "reducing the thickness of the disk to reduce the rigidity of the disk increases the magnetic resistance". As a result, the magnetic circuit for obtaining sufficient attractive force and the disk shape for obtaining appropriate rigidity can be individually selected.

Specifically, in the present invention , the bending rigidity of the elastic support portion is configured to be smaller than the bending rigidity of the magnetic path portion.
That is, in the present invention, since the magnetic circuit design and the disk structure design can be performed separately, a large disk displacement can be obtained while maintaining the same rigidity.

Specifically, in the present invention , the first magnetic pole faces the central portion of the magnetic path portion, and the second magnetic pole faces the vicinity of the outer edge of the magnetic path portion. .

That is, in the present invention, the magnetic circuit is designed to obtain an appropriate attractive force by providing the second magnetic pole that forms a closed loop magnetic circuit via a bypass member that serves as a magnetic path that bypasses the elastic support portion. , can separate the structural design of the flapper to obtain the proper stiffness.

Specifically, in the present invention , the nozzles are provided at two locations, one of which is provided on the fluid supply side and configured as a forward nozzle, and the other nozzle is provided on the fluid exhaust side. The forward nozzle, the reverse nozzle, and the flapper constitute a two-way nozzle flapper valve, and the working fluid supplied from the fluid supply source flows in the forward direction from the supply source side. The fluid passes through the nozzle, flows into the control chamber, which is a space in which the flapper is housed, and flows out from the control chamber through the reverse nozzle to the discharge side of the fluid. The reverse direction nozzle is arranged substantially coaxially on the opposite side of the flapper.

That is, in the present invention, the advantage of configuring the magnetic gap to directly change the air gap is that the forward nozzle (supply side nozzle), the reverse nozzle (exhaust side nozzle), and the two-way flapper are configured. It can also be applied to nozzle flapper valves with

Specifically, according to the present invention , the magnetic characteristics of the magnetic flux density with respect to the magnetizing force of the material constituting the flapper are selected according to the slope of the change region of the flapper displacement characteristics with respect to the desired current or the flapper flow characteristics with respect to the current. It is.

That is, in the present invention, in the magnetic characteristics of the magnetic flux density with respect to the magnetic force of the flapper material, the inclination angle of the current value in the linear region, the boundary value of the magnetic force transitioning from the linear region to the magnetic saturation region, and the magnetic flux density By appropriately selecting the maximum value (saturation magnetic flux density), the flapper displacement (flow rate) characteristic with respect to the current value can be obtained in accordance with the required specifications of the valve.

Specifically, the present invention connects the suction port to the fluid supply source, attaches a flow meter to the flow path leading from the control room to the atmosphere, and when the current energized to the electromagnet is the maximum value Imax (A) The flow rate measured by the flow meter is Qmax (L/min), the slope Qmax/Imax is the reference flow rate gain α, and the maximum value of the slope is the maximum flow rate gain β in the profile of the flow characteristic for the input current, and linearization It is configured so that η>0.2 when the effect index η=α/β is defined.

That is, in the present invention, by setting the linearization effect index η>0.2, practically satisfactory performances can be obtained for both the vibration isolation characteristics and the vibration damping characteristics. Insufficient performance due to the application target can be compensated for by synthesis (comprehensive design) of the entire control system.

Specifically, the present invention is constructed so that η>0.4.
That is, in the present invention, by setting the linearization effect index η>0.4, sufficient performance can be obtained for both vibration isolation and damping characteristics. The valve is highly versatile and does not depend on the application of the valve.

Specifically, in the present invention , the electromagnet is composed of a support shaft made of a magnetic material, a coil wound around the support shaft as an axis, and a cylindrical portion made of a magnetic material arranged to accommodate the coil. A closed loop magnetic circuit is formed by the support shaft, the flapper, and the cylindrical portion.

That is, in the present invention, a closed loop magnetic circuit is configured by the cylindrical portion housing the coil and the flapper with the support shaft as the center of the magnetic path. Since the magnetic circuit is axially symmetrical, the fluid flow can also be axially symmetrical. In addition, a fluid servo valve having a simple structure can be realized by constructing a housing that houses these parts with axisymmetric parts that are easy to work.

Specifically, according to the present invention , a flow passage is formed through the support shaft and communicates with a fluid supply side or an exhaust side, and the nozzle is provided at the opening end of the flow passage on the flapper side. be.

That is, in the present invention, the closed loop magnetic circuit is affected by forming a flow passage through the support shaft that houses the coil in the outer peripheral portion, and by providing a nozzle at the opening end of the flow passage on the flapper side. Therefore, a fluid servo valve can be realized with a simple configuration.

Specifically, in the present invention , the nozzles are provided at two locations, one of which is provided on the fluid supply side and configured as a forward nozzle, and the other nozzle is provided on the fluid exhaust side. The forward nozzle, the reverse nozzle, and the flapper constitute a two-way nozzle flapper valve, and the forward nozzle or the reverse nozzle passes through the support shaft. It is provided at the flapper side open end of the flow passage formed by the above.

That is, in the present invention, the support shaft is used to form a through hole (flow path) that serves as a flow path, and one of the two nozzles is provided at the opening end of the through hole on the flapper side. is easy to mount, and adjustment of the projection amount of the nozzle is easy. Alternatively, even when the nozzle is directly formed by machining on the flapper-side end face of the flow path penetrating the support shaft, the protrusion amount can be formed with high machining accuracy. With this configuration, it is possible to realize a nozzle flapper valve using a two-way flapper that allows a flow path configuration with low flow path resistance and good sealing performance (leakage prevention).

Specifically, the present invention includes a supply-side housing provided with the forward nozzle and having a flow path leading to the fluid supply source side, and a flow path provided with the reverse nozzle and leading to the fluid exhaust side. is formed, a supply-side gap that is a space formed between the flapper and the flapper-facing surface of the supply-side housing, and between the flapper and the flapper-facing surface of the exhaust-side housing and an exhaust side gap that is a space formed in the flapper, and a communication hole that connects the supply side gap and the exhaust side gap is formed in the flapper.

That is, in the present invention, by forming a communication hole in the flapper that connects the supply-side gap and the exhaust-side gap, which are independent sealed spaces, the two sealed spaces are a common space ( control room). As a result, the flapper can be configured so that an axial load due to a pressure difference is not applied.

Specifically, in the present invention , the flapper has a plate-like shape and is provided with an elastically deformable portion configured to be deformable toward the nozzle.

That is, in the present invention, the magnetic path portion is formed in the vicinity of the first magnetic pole, and the elastic support portion is formed on the flapper outer peripheral portion fixed side away from the closed loop magnetic circuit. can set the support stiffness of the flapper.

Specifically, according to the present invention , the electromagnet has a magnetic pole formed on an end face facing the flapper, the flapper has a plate-like shape, faces the magnetic pole at the central portion, and has the elastic deformation portion. is formed by a through-hole penetrating in the thickness direction formed between the central portion of the flapper and the flapper support member in the flapper.

That is, in the present invention, by applying a helical spring or a cloud-shaped spring that is easily elastically deformable to the elastic support portion, the supporting outer diameter of the flapper can be reduced, and the diameter of the valve body can be reduced.

Specifically, according to the present invention , a radial flow passage is formed in the flapper-side end surface of the support shaft to communicate with the open end of the nozzle provided on the support shaft.

That is, in the present invention, the gap between the end surface of the support shaft and the end surface of the flapper is sufficiently formed by forming radial flow passages communicating with the opening end of the nozzle in the end surface of the support shaft on the flapper side. Can be set smaller. Therefore, the amount of protrusion of the nozzle from the end face of the support shaft can be small, and the initial gap between the flapper and the magnetic pole (end face of the support shaft) can be made small, so that a large flapper maximum displacement (maximum flow rate) can be obtained with a small current. .

Specifically, according to the present invention , a bypass member and a second magnetic pole are formed in the closed loop magnetic circuit so that the magnetic flux bypasses the elastic support portion and has an outer diameter smaller than the coil diameter.

That is, in the present invention, the second magnetic pole that forms a closed loop magnetic circuit is provided in a ring shape with an inner diameter smaller than the outer diameter of the coil, for example, and is connected to a bypass member that serves as a magnetic path that bypasses the elastic support portion. As a result, a sufficient suction force can be obtained even if the outer diameter of the disk is reduced, and the size of the valve body can be reduced.

Specifically, in the present invention , an annular flow path forming structure that forms a flow path having a substantially annular cross section is formed between the nozzle and the flapper, and the annular flow path forming structure It consists of a cylindrical inner peripheral surface that forms the outer boundary of the annular flow path, and an insert that is inserted radially away from the inner peripheral surface.

That is, in the present invention, since the annular flow path forming structure is formed between the nozzle and the flapper, the axial length of the flow path formed by the annular flow path forming structure is Since it changes according to the movement of the flapper, it is possible to obtain a flow rate characteristic different from that of a normal nozzle flapper valve. For example, a valve can be constructed that has a section with a very small slope of flow versus current.

Specifically, in the present invention , the annular flow path forming structure is formed between the forward nozzle and the flapper and between the reverse nozzle and the flapper, and the fluid flows from the supply source side. It is configured to pass through the forward nozzle, flow into a control chamber, which is a space in which the flapper is housed, and flow out from the control chamber through the reverse nozzle to the discharge side of the fluid. characterized by

That is, in the present invention, for example, at a substantially intermediate position (operating point) of the movable range of the flapper, the nozzle-side end face of the supply-side insert is brought into a state close to the opening end of the supply-side nozzle, and the The fluid flow entering the control chamber can be a transition region transitioning from a viscous flow region to a potential flow region. Also, the flow of fluid from the control chamber to the exhaust-side nozzle can similarly be in the transition region. Therefore, the flow rate characteristic with respect to the flapper displacement can be a downwardly convex curve with respect to the amount of displacement of the flapper. Therefore, at the operating point of the valve, the steady-state flow rate of the fluid flowing from the supply source side to the exhaust side can be made sufficiently small.

Specifically, in the present invention , the cylindrical inner peripheral surface is the inner peripheral surface of the tip portion of the nozzle, and the insert is a substantially conical convex portion formed on the face plate portion of the flapper. characterized by being

That is, in the present invention, by tapering the flapper-side convex portion that engages with the nozzle so that the flow passage area between the nozzle flappers changes gently with respect to the current value, the flow rate characteristic with respect to the current value is excellent in linearity. characteristics.

Specifically, the present invention uses a closed loop magnetic circuit configured to include at least the electromagnet and the flapper, and sets the maximum stroke between the nozzle and the electromagnet to 0.5 mm or more.

That is, in the present invention, by further actively utilizing the magnetic saturation phenomenon, the stroke of the flapper can be greatly increased without losing the linearity of the displacement characteristic of the flapper with respect to the current. It is. Since the flapper can be driven with a large stroke, it is possible to construct and process parts that can adjust the fitting state of the convex portion of the flapper and the orifice on the nozzle side by moving the flapper in the axial direction.

Specifically, the present invention includes a nozzle having a flow path connected to a fluid supply source and having a tapered inner diameter at the tip, and a plate-like nozzle provided so as to face the tip of the nozzle. a flapper, a flapper support member that fixedly supports a part of the flapper, and an electromagnet provided to generate an attraction force with respect to a face plate portion of the flapper, and the attraction force of the electromagnet causes the flapper is deformed to change the separation distance between the tip of the nozzle and the face plate portion of the flapper, and the space for storing the flapper is partitioned by the flapper, one surface of the flapper and A control chamber in contact with the flapper and a constant-pressure chamber in contact with the other surface of the flapper and kept at a substantially constant pressure are formed. Hc or less, the magnetic flux density B converges to a predetermined value, and the magnetic flux density of the magnetic flux flowing through the magnetic material component increases in the magnetic saturation region when the current flowing through the electromagnet is increased in the displacement range of the flapper characterized by entering

That is, in the present invention, by providing a constant pressure chamber formed between the flapper and the fixed housing, which maintains a substantially constant pressure, the load due to the pressure difference applied before and after the flapper and the electromagnetic attractive force are balanced. A valve characteristic can be obtained in which the control pressure is proportional to the input current of the electromagnet.

Specifically, the present invention forms a closed loop magnetic circuit by connecting magnetic material members in a substantially polygonal shape in a cross-sectional view, and each of the magnetic material members includes an iron core around which a coil of an electromagnet is wound, a yoke member, and a flapper. It consists of

Specifically, according to the present invention , a magnetic pole is formed at one end of the magnetic material member, and the flapper is disposed on the opposing surface with a gap therebetween.

Specifically, according to the present invention , a flow passage is formed through one of the magnetic material members and communicates with a fluid supply side or an exhaust side, and the nozzle is an open end of the flow passage on the flapper side. It is set in

Specifically, the present invention includes the above -described fluid servo valve, a sensor for detecting the displacement and/or vibration state of the object to be controlled, and adjusting the fluid servo valve based on information from the sensor. It is composed of control means for applying gas pressure to the pneumatic actuator for controlling the displacement, velocity, acceleration, etc. of the object to be controlled.

In other words, in the present invention, the complicated structure of the conventional pneumatic servo valve derived from hydraulic servo technology can be greatly simplified, so that the performance can be maintained by a simple configuration in terms of structure and control. A pneumatic servo system can be realized.

Specifically, the present invention comprises a gas spring configured to have a primary natural frequency of 200 Hz or more of the flapper, and supporting an object to be vibration-isolated on a foundation, and supplying gas from a supply side to the gas spring. The fluid servo valve for exhausting air to the exhaust side, an acceleration sensor for detecting the vibration state of the vibration isolation object, and the vibration isolation object by adjusting the fluid servo valve based on information from the acceleration sensor. It is composed of active control means for applying a gas pressure to the gas spring to reduce the vibration of the object.

That is, in the present invention, the following features of the valve of the present embodiment are provided: (1) the resonance frequency can be set high, (2) the valve can be driven with a small electric power, (3) high-speed response can be obtained, and (4) ) The structure is simple, the number of parts is small, and the parts are easy to process, assemble, and adjust.The overall performance of the device is greatly improved, and the primary natural frequency can be set high, so acceleration feedback can be used more effectively. An active anti-vibration table can be realized.

Specifically, the present invention comprises an electromagnet, a disk, a supporting member for fixing the disk, the electromagnet, the disk, and the yoke material to form a closed loop magnetic circuit, and the Maxwell magnetic field generated between the magnetic pole and the disk. A portion composed of an output shaft that is movable by an attractive stress and is fixed to the disk, the electromagnet, the disk, the support member, the yoke material, and the output shaft is defined as a microactuator portion, A portion composed of a fluid suction port, a fluid discharge port, and a passage opening adjustment portion interposed between the discharge port and the suction port is defined as a fluid control section, and the output shaft of the microactuator section and the A fluid servo valve characterized by connecting a fluid control unit and operating the passage opening adjustment unit by the output shaft, wherein the magnetic characteristics of the magnetic material parts constituting the closed loop magnetic circuit are magnetized . The area below the force boundary value Hc, where the magnetic flux density B with respect to the magnetizing force H has a roughly proportional relationship , and the area above the magnetizing force boundary value Hc, where as the magnetizing force H increases, The magnetic flux density B is defined as a magnetic saturation region where the magnetic flux density B converges to a predetermined value. It is configured to enter

That is, in the present invention, by utilizing the magnetic saturation phenomenon in the above-described manner, the displacement characteristic of the output shaft with respect to the current can be obtained with excellent linearity (linearity). It utilizes the fact that the stroke of the shaft can be greatly increased. Milli-order displacement control is possible, which was not possible with conventional piezo actuators and giant magnetostrictive actuators, which were limited to about 50 to 100 μm. Furthermore, compared to voice coil motors (linear motors), the thrust constant is high and can be driven with a small amount of electric power, making it possible to significantly reduce the size of the actuator.

Specifically, in the present invention , the output shaft is provided so as to pass through the central portion of the electromagnet.

That is, in the present invention, when a current is applied to the coil, the end of the output shaft protrudes from the actuator main body. Therefore, although the present actuator is of the magnetic attraction type, it can be used in the same manner as the widely used piezoelectric and magnetostrictive actuators.

Specifically, in the present invention , a displacement/velocity/acceleration sensor is arranged at the end of the output shaft opposite to the electromagnet for detecting positional information such as displacement, speed, and acceleration of the output shaft. It is.

That is, in the present invention, the sensor can be arranged using the space on the opposite side of the output shaft end by making the central shaft pass through the outer frame portion, and the microactuator with a built-in sensor can be provided. can be configured compactly.

Specifically, according to the present invention , the output shaft is radially supported by a hydrostatic bearing.

That is, in the present invention, if a hydrostatic bearing is used for the bearing that supports the central shaft, the influence of Coulomb friction (static friction) can be avoided, so the displacement of the output shaft end can be controlled with high accuracy.

Specifically, in the present invention , in the supply-side channel connecting the suction port and the nozzle opening, the length of the rectification section provided on the upstream side of the opening is L, and the average length of the rectification section is It is constructed so that L/d>4, where d is the inner diameter.

That is, in the present invention, a high-pressure fluid force is applied to the flapper from the supply-side nozzle, and this fluid force excites the flapper, generating noise and destabilizing the flow rate. is to eliminate By setting the average value of the channel diameter closest to the nozzle opening on the supply side to Φd, forming a rectification section with a length L that does not cause a steep change in the flow velocity, and setting this length L to be sufficiently large. , it was found that the unstable phenomenon of the flapper caused by the fluid force can be eliminated.

Now, if the features of the fluid servo valve according to the present invention are listed,
(1) The resonance frequency can be set high (2) The valve can be driven with a small amount of power (3) High-speed response can be obtained (4) The structure is simple, the number of parts is small, and the parts are easy to process, assemble and adjust Conventional valves It is expected that the valve of the present invention, which largely eliminates the shortcomings of (1), will greatly accelerate the widespread use of pneumatic servo systems in the future. The effect is remarkable.

1 is a front sectional view of a fluid servo valve according to Embodiment 1 of the present invention; FIG. FIG. 4 is a diagram showing the shape of a flapper, which is a disk-shaped disk in Embodiment 1; 3(a) is a top view of FIG. 3(b), and FIG. 3(b) is a front cross-sectional view. FIG. FIG. 4A is a partial enlarged view showing the positional relationship of the flapper, the supply side nozzle, and the exhaust side nozzle in Embodiment 1, FIG. illustration. The graph which shows the analysis result of the flapper displacement characteristic with respect to an electric current. 4 is a graph showing analysis results of control pressure characteristics with respect to current; 4 is a graph showing analysis results of control flow rate characteristics with respect to current when the exhaust port is blocked; 4 is a graph showing analysis results of internal leak flow rate characteristics with respect to current; 6 is a graph showing flapper displacement characteristics with respect to current, for explaining the outline of the second embodiment of the present invention; FIG. 10a is a model diagram for analytically obtaining the magnetic resistance of the magnetic flux flowing through the disk, FIG. 10b is a front cross-sectional view showing the disk and the electromagnet as a flapper, and FIG. 10c is an outflow source of the magnetic flux flowing radially through the disk. FIG. 4 shows a flux control surface; 4 is a graph showing magnetic flux density characteristics with respect to magnetizing force, as an example of the magnetic characteristics of the test material in the present example. The graph which shows the analysis result of the magnetic flux density characteristic with respect to an electric current. 4 is a graph showing analysis results of electromagnet attractive force characteristics with respect to electric current; The graph which shows the analysis result of the flapper displacement characteristic with respect to an electric current. 4 is a graph showing analysis results of control flow rate characteristics with respect to current when the exhaust port is blocked; The graph which shows the method of calculating|requiring the effect index of linearization based on the measured value of the flow rate with respect to a current value. FIG. 5 is a diagram showing an example of actual measurement of valve flow characteristics for obtaining a “linearization effect index”; FIG. 11 is a front cross-sectional view of a fluid servo valve using a convex disc-shaped disk according to Embodiment 3 of the present invention; FIG. 11 is an enlarged view of the vicinity of the flapper and the electromagnet in Embodiment 3; The graph which shows the analysis result of the flapper displacement characteristic with respect to an electric current. The graph which shows the analysis result of the generated force with respect to an electric current. 4 is a graph showing analysis results of control flow rate characteristics with respect to current when the exhaust port is blocked; FIG. 11 is an enlarged view of the vicinity of the electromagnet when the disc shape is a concave disc shape in the fluid servo valve according to the third embodiment; FIG. 5 is a front cross-sectional view of a fluid servo valve according to Embodiment 4 of the present invention; 25a is a partial enlarged view of the vicinity of the exhaust nozzle and the intake nozzle in Embodiment 4, FIG. 25a is a view on arrow AA of FIG. 25b, FIG. 25b is a view of arrow BB of FIG. 25c, and FIG. illustration. The graph which shows the analysis result of the flapper displacement characteristic with respect to an electric current. FIG. 11 is a front cross-sectional view of a fluid servo valve according to Embodiment 5 of the present invention; FIG. 11 is a front cross-sectional view of a fluid servo valve according to Embodiment 6 of the present invention; FIG. 12 is an enlarged view of the closed loop magnetic circuit of the fifth embodiment described above, in which only one magnetic pole is provided; FIG. 4 is an enlarged view of the closed-loop magnetic circuit of the present embodiment in which a second magnetic pole is provided as an auxiliary in addition to the first magnetic pole; The graph which shows the analysis result of the flapper displacement characteristic with respect to an electric current. The figure which adjusted rigidity by forming a gap part in a flapper. Fig. 33a is a view showing a spiral disk spring in the AA view of Fig. 33c, Fig. 33b is a partial enlarged view of Fig. 33a, and Fig. 33c is a front cross-sectional view of a fluid servo valve according to Embodiment 7 of the present invention. Fig. 33c is a partially enlarged view of the spiral disk spring portion of Fig. 33c; FIG. 4 is a diagram showing structural analysis results of deformation of the spiral disk spring when a load is applied to the central portion of the magnetic path; The figure which used the cloud-shaped spring for the disk in Embodiment 7 of this invention. Fig. 37a is a top view of Fig. 37b, and Fig. 37b is a sectional front view of a fluid servo valve according to Embodiment 8 of the present invention; FIG. 11 is a front cross-sectional view of a fluid servo valve according to Embodiment 9 of the present invention; 39a is a top view of FIG. 39b, and FIG. 39b is a cross-sectional front view of a fluid servo valve according to Embodiment 10 of the present invention. FIG. 11 is a front cross-sectional view of a fluid servo valve according to Embodiment 11 of the present invention; FIG. 12 is a front cross-sectional view of a fluid servo valve according to Embodiment 12 of the present invention; FIG. 21 is a front cross-sectional view of a fluid servo valve according to Embodiment 13 of the present invention; Fig. 43a, Fig. 43b, Fig. 43c are partial enlarged views of the nozzle and the flapper in the thirteenth embodiment, showing a fitting state between the nozzle and the flapper; Fig. 44a is a diagram showing the principle of the fluid servo valve in Embodiment 13, Fig. 44a is a diagram showing the fitting state between the nozzle and the flapper, and Fig. 44b is a graph showing the valve flow rate with respect to the gap between the nozzle flappers. . Fig. 45a is a graph showing the valve flow rate with respect to the gap between the nozzle flappers of the conventional valve, and Fig. 45b shows the valve flow rate with respect to the gap between the nozzle flappers of the present invention. graph. FIG. 20 is a graph showing analysis results of flapper displacement characteristics with respect to current in the thirteenth embodiment; FIG. Fig. 47a, Fig. 47b, and Fig. 47c are diagrams showing the fitting state between the nozzle and the flapper in the fluid servo valve according to the thirteenth embodiment when the flapper-side convex portion is tapered; FIG. 21 is a front cross-sectional view of a fluid servo valve according to Embodiment 14 of the present invention; 49a, 49b, and 49c are diagrams showing the fitting state between the nozzle and the flapper in the fluid servo valve according to the fourteenth embodiment; FIG. FIG. 20 is a front cross-sectional view of a fluid servo valve according to Embodiment 15 of the present invention; FIG. 51 is an enlarged view of the nozzle flapper portion of FIG. 50 in the fluid servo valve according to the fifteenth embodiment; FIG. 12 is a front cross-sectional view showing Embodiment 16 of the present invention, showing a case where the present invention is used as an electro-pneumatic converter; FIG. 14 is a front cross-sectional view showing Embodiment 17 of the present invention, and is a front cross-sectional view of a valve showing a case where the present invention is used as a pilot valve; FIG. 11 shows a four-way guide valve driven by Embodiment 17 of the present invention; FIG. 18 is a front cross-sectional view showing an eighteenth embodiment of the present invention, showing a microactuator section; FIG. 18 is a cross-sectional front view of a poppet valve coupled to a microactuator portion of embodiment 18 of the present invention; FIG. 18 is a cross-sectional front view of a spool valve coupled to a microactuator portion of embodiment 18 of the present invention; FIG. 20 is a front cross-sectional view of a fluid servo valve according to Embodiment 19 of the present invention; FIG. 20 is a reference diagram for comparison with the fluid servo valve according to Embodiment 20 of the present invention, and is an enlarged view of the nozzle and flapper portions in Embodiment 3; FIG. 12 is an enlarged view of the nozzle and flapper portions in Embodiment 6 of the fluid servo valve according to Embodiment 20 of the present invention; FIG. 4 is a diagram showing the characteristics of three types (A, B, and C) of magnetic materials having different magnetic flux density characteristics with respect to magnetic field intensity; FIG. 62 is a diagram qualitatively showing the difference in the displacement characteristics of the flapper with respect to current when the materials A, B, and C in FIG. 61 are used; FIG. 4 is a cross-sectional view showing an electromagnet in which a portion with a small magnetic path area is provided on the central axis for adjusting the magnetic saturation phenomenon; It is a model diagram explaining the skyhook theory in active control. Fig. 64a shows a state in which an object is suspended in an imaginary line stretched in the air, Fig. 64b shows an object supported by an actuator and an acceleration sensor arranged. figure. 4 is a graph showing a model of vibration isolation performance of a vibration isolator using a pneumatic actuator. FIG. 3 is a block diagram showing an example of an analysis model of an active control anti-vibration table; Fig. 67a is a diagram showing an example of analysis results of open-loop transfer characteristics of an active control anti-vibration table, Fig. 67a is a diagram showing gain characteristics with respect to frequency, and Fig. 67a is a diagram showing phase characteristics with respect to frequency; A model diagram showing an active anti-vibration table equipped with a conventional fluid servo valve. Fig. 69a is a front sectional view, Fig. 69b is a side sectional view, showing a conventional fluid servo valve; The figure which modeled the conventional fluid servo valve. FIG. 2 is a front cross-sectional view of a conventionally proposed fluid servo valve using a linear motor;

<First Embodiment>
FIG. 1 is a front cross-sectional view of a pneumatic servo valve according to Embodiment 1 of the present invention.
Reference numeral 10 denotes a cylindrical central shaft (support shaft) made of a magnetic material; 11, a bottom portion of the central shaft; 12, an outer frame formed concentrically with the central shaft; A coil bobbin made of a non-magnetic material, 14 is a coil wound around the coil bobbin. The central shaft 10, the outer frame portion 12, the coil bobbin 13, and the coil 14 constitute an electromagnetic actuator (electromagnet) that attracts a face plate portion of a flapper (described later) and controls its displacement. Reference numeral 15 denotes a cylindrical housing that houses the outer frame portion 12; 16, an exhaust-side bottom plate fastened to the side surface of the housing; 17, bolts for fastening the bottom portion 11 and the exhaust-side bottom plate 16; 18, the housing 15 and the exhaust-side bottom plate; 19 is an exhaust-side flow passage for gas (working fluid) formed in the central shaft 10; 20 is a discharge port formed in the exhaust-side bottom plate 16; 21 is a supply-side bottom plate, 22 is a gas supply-side channel formed in the center of the supply-side bottom plate, and 23 is a gas control-side channel connected to a pneumatic actuator (not shown). A disk-shaped flapper 24 is mounted between the housing 15 and the bottom plate 21 on the supply side with bolts 25 . That is, the bolt, the housing, and the bottom plate on the supply side are flapper supporting members, which sandwich and fix the outer edge of the flapper 24 so that the outer edge does not move. 26 is a supply side gap formed between the flapper 24 and the wall surface of the supply side bottom plate 21, and 27 is an exhaust side gap formed between the flapper 24 and the exhaust side wall surface (coil bobbin 13, housing 15, etc.).

In the disk-shaped flapper 24 shown in FIG. 2, 28a, 28b, 28c, and 28d are communication holes formed in the flapper that connect the supply side gap 26 and the exhaust side gap 27 (28b and 28d are not shown). 29 is a supply side nozzle (forward direction nozzle), and 30 is an exhaust side nozzle (reverse direction nozzle). Reference numeral 31 denotes a flapper-side end face of the central shaft 10 (first magnetic pole at the central shaft end face), and 32 a flapper-side end face of the outer frame portion (second magnetic pole at the outer frame end face). The space formed by the supply-side gap 26, the exhaust-side gap 27, and the control-side passage 23 is the intake port of the control chambers 33 and 34 of this valve. Incidentally, the term "flapper" generally conjures up an image of a flat plate that oscillates, as shown in a model diagram 64 of a conventional valve. In the present invention including this embodiment, a member that is arranged on the surface facing the nozzle and adjusts the flow path area of the fluid between the nozzle and the nozzle is called a flapper regardless of the shape of the member.

In the present embodiment, the exhaust-side nozzle is arranged on the center line of the magnetic poles (the first magnetic pole and the second magnetic pole) and on the side of the magnetic poles, and the flapper is interposed. A supply nozzle is arranged on the opposite side of the magnetic pole.

3a and 3b are enlarged views of the vicinity of the electromagnetic actuator, with FIG. 3a being a top view and FIG. 3b being a front sectional view. The dotted line with an arrow in the figure indicates the magnetic flux generated by energizing the coil 14. This magnetic flux causes the following sequence: first magnetic pole 31→air gap 27→flapper 24→air gap 27→second magnetic pole 32 →Outer frame portion 12→Bottom portion 11→Center shaft 10” forms a closed loop magnetic circuit. However, when the direction of the current flowing through the coil 14 is reversed, the direction of the magnetic flux is reversed. Here, Φ is the magnetic flux flowing through the magnetic circuit, S1 is the ring-shaped area of the center shaft end face 31 (first magnetic pole), S2 is the ring-shaped area of the outer frame end face 32 (second magnetic pole), and μ0 is the magnetic permeability of air. Then, the attractive force F acting on the face plate of the flapper 24 due to Maxwell's stress is

吸引力Fによりフラッパが変位して、フラッパに働く円盤ばねの反力とこの吸引力Fが平衡する。磁束Фはコイル１４に通電する電流値に比例するために、電流を可変させることにより、フラッパ変位、即ちノズルとその対向面間の隙間（離間距離）を調節できる。

The flapper is displaced by the attraction force F, and the reaction force of the disk spring acting on the flapper and this attraction force F are balanced. Since the magnetic flux Φ is proportional to the current flowing through the coil 14, the flapper displacement, that is, the gap (separation distance) between the nozzle and its opposing surface can be adjusted by varying the current.

FIG. 4 is a partial enlarged view showing the positional relationship between the flapper 24, the supply side nozzle 29, and the exhaust side nozzle 30. FIG. FIG. 4b shows the state where the coil is energized and the flapper 24 is between the supply side nozzle 29 and the exhaust side nozzle 30. FIG. 4a, X0 is the gap (initial gap) between the flapper 24 and the first magnetic pole 31, .delta.n is the amount of protrusion of the tip of the exhaust side nozzle 30 with respect to the end face of the first magnetic pole 31, and .delta.a is the tip of the exhaust side nozzle 30 and the flapper 24. It is the gap (flow path length) between the two, and is the maximum stroke of the flapper 24 . In the embodiment, the positional relationship of each member is set so that the flapper 24 shields the tip of the supply side nozzle 29 as shown in FIG. When the coil 14 is energized, the flapper 24 moves away from the tip of the supply nozzle 29, as shown in FIG. 4b. Here, the displacement X is the amount of movement of the flapper 24 from the position of the initial gap X0 toward the exhaust nozzle 30 side. Also in the following embodiments, "displacement X is defined as the amount of movement from the position of the initial gap X0".

FIG. 5 shows the flapper displacement X with respect to the current value I under the setting conditions described in the same figure. The analysis method is i. Given the gap (X0-X) between the magnetic pole and the flapper, find the attractive force F by magnetic field analysisii. Find the flapper displacement X from the attraction force F and the flapper support stiffness K iii. Considering the relationship between the magnetizing force H and the magnetic flux density B (FIG. 11) on the magnetic flux control surface, i. ii. as a coupled problem and perform convergence calculations.

Step i. ~iii. It was obtained through However, the above iii. will be described in detail later. From the graph in FIG. 5, X=0.045 mm when I=0.02A. Here, when I=0.02A, if the flapper 24 is set to shield the exhaust nozzle 30, the maximum flapper stroke Xmax=.delta.a=0.045 mm. Therefore, the nozzle protrusion amount Δn should be set to X0-Δa=0.250-0.045=0.205 mm.

Below, the pressure/flow rate characteristics of the servo valve of this embodiment are obtained when the gap between each nozzle and flapper is given. The gas mass flow rate through the nozzle of the servovalve uses the nozzle equations (4) and (5) for isentropic flow of compressible fluid. The opening area between the nozzle flappers is the area of the annular flow path formed between the nozzle tip and the flapper, where the nozzle inner diameter is d, the supply side opening area ain = dπX, the exhaust side opening area aout = dπ(δa-X) is. The mass flow rate Gin of the gas flowing into the air chamber from the supply source side is shown by the following equation. where Ps is the source pressure, Pa is the control chamber pressure of the servovalve, ρs is the source gas density, and κ is the specific heat ratio.

但し、Pa/Ps＜{2/(κ+1)}2/(κ-1) のときは

However, when Pa/Ps<{2/(κ+1)}2/(κ-1)

The mass flow rate Gout of the gas flowing out from the control chamber to the atmosphere side may be Ps → Pa, Pa → P0, ρs → ρa, aout = dπ(δa-X) in equations (4) and (5). . Vc is the volume of the control chamber 33 and R is the gas constant. Based on these mass flow rates Gin and Gout, the pressure Pa in the control chamber 33 is obtained by the following equation.

FIG. 6 shows analysis results of the control pressure in the steady state with respect to the current value in the pneumatic valve according to the first embodiment. The control pressure is the pressure Pa in the control chamber 33 formed by the supply-side gap 26 and the exhaust-side gap 27 connected to the control-side channel 23 . The analysis conditions are supply pressure PS=0.6 MPa (abs), atmospheric pressure P0=0.1 MPa (abs), and nozzle inner diameters of the supply side nozzle 29 and the exhaust side nozzle 30 are both Φ1.2 mm.

FIG. 7 shows the control flow rate characteristics with respect to the current when the exhaust port is blocked. The profile of the control flow rate characteristic versus current value substantially matches that of the flapper displacement characteristic versus current value in FIG.

FIG. 8 shows the internal leak flow rate with respect to the current value in the pneumatic servo valve. Here, the internal leak flow rate is defined as the flow rate QL from the exhaust side flow path 19 in a state where the control side flow path 23 of the valve is blocked.
In the graph of FIG. 5 of flapper displacement versus current value, current value I=0.0118 A, displacement X=0.02 mm, and flapper 27 is roughly midway between nozzle 29 on the supply side and nozzle 30 on the exhaust side. At this time, it can be seen that the internal leak flow rate QL shown in FIG. 7 exhibits the maximum value.

The features of the valve of this embodiment are listed below.
(1) The resonance frequency can be set high (2) The valve can be driven with low power (3) High-speed response can be obtained (4) The structure is simple and the number of parts is small, making it easy to process, assemble and adjust parts The reason for 1) is as follows. Here, if the flapper's movable mass is m and the spring constant that supports this flapper is K, the resonance frequency f0 is

As described above, in the conventional servo valve (FIG. 69), the flapper 506, which is a rigid body, oscillates, so the mass m of the flapper is inevitably large. Therefore, by setting the stiffness K of the spring supporting the flapper high, the resonance frequency f0 [formula (7)] is set high. If the spring stiffness K of the flapper support is large, a large force is required to drive the flapper. As mentioned above, in the case of the conventional servo valve, the force driving the flapper was F∝Φ1Φ2 as shown in equation (2). In other words, a large driving force is obtained by amplifying the magnetic flux Φ1 generated by the electromagnetic coil with the magnetic flux Φ2 generated by the permanent magnet.

However, in the case of this embodiment, only the magnetic flux Φ1 generated by the electromagnetic coil becomes the force for driving the flapper valve. The reason why a high resonance frequency is obtained in spite of this is as follows.

In this study, we focused on the fact that the effective mass m of the movable part, which acts as an inertial load, is only the elastically deformed part near the tip of the nozzle when the member corresponding to the flapper of the valve is made into a thin disc shape. That is, in the conventional servo valve, the flapper, which is a rigid body of mass m, is supported by a spring, whereas in the servo valve of the present invention, the flapper itself is an elastic body (spring). Table 1 compares the valve of this embodiment and the conventional flapper valve (an example) with respect to the effective mass of the movable portion, the spring stiffness of the flapper support, and the resonance frequency. The effective mass m of the movable portion in this embodiment was obtained from the spring stiffness K and the measured value of the resonance frequency f0 using the equation (7). Also, since the flapper (corresponding to the flapper 506 in FIG. 69), which is the movable part of the conventional servo valve, oscillates, the effective mass, which is the inertial load, was assumed to be 1/2 of the measured value (5 g).

As can be seen from Table 1, the valves according to the present invention have an effective mass of about 1/7 and a spring stiffness of about 1/4 compared to the conventional valve, although they have a resonance frequency equal to or greater than that of the conventional valve. Since the stiffness of the spring supporting the flapper can be made sufficiently small, the valve of this embodiment can drive the flapper only with the electromagnet, although it has a high resonance frequency. Conventional pneumatic servo valves that require both permanent magnets and electromagnets, as shown in Figures 69a and 69b, have yoke members arranged orthogonally to form respective magnetic circuits of the permanent magnets and electromagnets. It has a three-dimensional structure, and the actuator section and the fluid control section are configured separately. Therefore, due to the complexity of the configuration and the large number of required parts, there is a problem that it is difficult to adjust the position between the flapper and the nozzle, and the cost is high.

The reason for the above (2) is as follows. The reason why the servo valve of the present invention can be driven with a small power (small current) is that Maxwell's stress acting on the conductor surface is used as a drive source. Generally, a voice coil motor (linear motor) is used as an actuator that linearly moves a minute displacement on the order of 0.1 mm to several mm. A servo valve using the above-described voice coil motor is also devised in Patent Document 4 mentioned above. However, the voice coil motor uses the Lorentz force, and a large thrust constant (electromechanical conversion efficiency) cannot be obtained. This embodiment takes advantage of the fact that Maxwell's stress, which has a much higher thrust constant than the Lorentz force, can be used for the limited target of pneumatic servo valves. Table 2 shows an example in which the thrust constant of this embodiment is compared with that of a commercially available voice coil motor.

From Table 2, the thrust constant of the actuator of the valve of this embodiment is 20 times or more that of the voice coil motor. For the above reason, the capacity of the power supply for driving the servo valve of this embodiment is sufficiently small, and a small current is sufficient. By the way, the thrust force constant of this embodiment is calculated by using the graph of FIG. 0.02A).

The reason (3) above is inevitably derived from the features (1) and (2) of the valve of this embodiment. That is, since the inertia 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 also small. Therefore, the transfer characteristic of the flapper displacement (flow rate) with respect to the input current can obtain sufficiently high responsiveness.

The reason for the above (4) is as follows. While the conventional pneumatic servo valve (see FIGS. 69 and 70) derived from hydraulic servo technology has a separate structure for the actuator section and the fluid control section, this embodiment has an actuator section and a fluid control section. is an integral structure. The actuator section in this embodiment is composed of the electromagnetic actuator section (electromagnet), the flapper, and the flapper support member (cylindrical housing, supply side bottom plate 21). Further, the fluid control unit is composed of the flapper, the nozzle, the supply side channel 22 including the suction port, the exhaust side channel 19 including the discharge port, and the support member. 4a and 4b showing enlarged views of the nozzle portion, as described above, FIG. 4a shows the state of the current value I=0 applied to the coil, and FIG. 4b shows the valve driving state. In the present invention, the reason why the actuator section and the fluid control section can be integrated is the maximum magnetic gap X0 between the magnetic pole and the flapper that can effectively use the magnetic attraction action, and the gap between the nozzle and the flapper that can be effectively used as an air servo valve. This is based on the fact that the maximum value of the air gap δa (channel length) is on the same order as 0.05 to 0.20 mm. The characteristic of the magnetic attraction force with respect to the air gap Δm is non-linear, and the magnetic attraction force usually drops significantly beyond the above maximum value. Similarly, in the case of the nozzle flapper valve, the air gap .delta.a that can linearly vary the flow rate is usually limited to the above range. Further, in this embodiment, the center shaft 10 of the electromagnetic actuator is shaped like a cylinder to form an air exhaust side flow passage 19 . With this configuration, it is possible to greatly simplify the nozzle flapper valve using the bidirectional flapper. Incidentally, in the embodiment, the positions of the supply side nozzle 29 and the flapper 24 are set so that the supply side nozzle 29 is blocked by the flapper 24 at the coil current value I=0 in FIG. 4A. This is a safety function (fail-safe function) that cuts off the inflow of high-pressure air to the pneumatic actuator (not shown) when active control becomes impossible during a power failure.

Further, the valve of this embodiment is composed entirely of axially symmetrical parts. Therefore, all parts can be manufactured only by lathe processing, the number of parts is small, and adjustment after assembly is simplified. The reason why the valve of the present embodiment can be configured with axial symmetry is that, as described above, the exhaust side nozzle is arranged on the center line of the magnetic poles (the first magnetic pole and the second magnetic pole) and on the magnetic pole side, and the flapper is arranged. This is because the supply-side nozzle is arranged on the opposite side of the magnetic pole intervening. However, the positions of the exhaust-side nozzle and the supply-side nozzle may be reversed.

Further, as described above, the sum of the spaces of the supply-side gap 26, the exhaust-side gap 27, and the control-side passage 23 is the total volume Vc of the control chamber 33 connected to the pneumatic actuator. Since the size of this volume Vc has a significant effect on control performance (response) when performing active control (fluid servo), it is preferable that the volume Vc be as small as possible. Since the valve of this embodiment is composed of axially symmetrical parts, the gaps Δt1 and Δt2 can be narrowed, and a space for housing the coil is not required compared to a valve having a three-dimensional structure (described later). Therefore, the total volume Vc of the control chamber 33 can be made sufficiently small.

<Second embodiment>
Assume a device that applies a current to an electromagnet and utilizes the magnetic attraction of the moving part due to Maxwell's total stress T. In the case of graph A in Fig. 9, the displacement characteristics of the moving part with respect to the current are non-linear characteristics in which the displacement rises sharply as the current value increases, so it is used for devices that require ON/OFF functions (relays, etc.). It is often done. However, in the course of this research, if an appropriate magnetic material and a thin disk were used for the movable part corresponding to the flapper, the displacement characteristics of the flapper with respect to the current would be linear (linearity), as shown in graph B in Fig. 9. It was found that excellent characteristics of This effect was found by an accidental discovery. In order to theoretically investigate this phenomenon found in this research and to evaluate its applicability to nozzle flapper valves, the following theoretical analysis was performed.

1. Theoretical Analysis Fig. 10 is a model of the pneumatic servo valve structure (Fig. 1) in Embodiment 1. Fig. 10a is a partial sectional view of a disk (flapper), and Fig. 10b is a modeled front section of the pneumatic servo valve. FIG. 10c shows a maximum flux control surface, which will be described later. In FIG. 10b, 210 is the central shaft, 211 is the gap, 212 is the flapper, and 213 is the outer frame. In the model diagram 10b, the magnetic flux Φ generated by energization of the coil draws a closed loop through the route of "center axis 210→air gap 211→flapper 212→air gap 211→outer frame 213" as described above. Here, the magnetic resistance of the magnetic circuit radially flowing through the disk is obtained. In FIGS. 10a and 10b, the magnetic resistance ΔRe in the radial direction Δr is

式(8)において、hはディスク（フラッパ）の厚み、μ0は空気の透磁率、μsはディスク材料の比透磁率である。半径r=r1からr=r2までの全抵抗を求めると

In equation (8), h is the thickness of the disk (flapper), μ0 is the permeability of air, and μs is the relative permeability of the disk material. Finding the total resistance from radius r=r1 to r=r2

上記ディスクの磁気抵抗 以外の磁気抵抗を として、磁束Φは

The magnetic flux Φ is

In Equation (10), N is the number of turns of the coil, and I is the value of current flowing through the coil. Also, is the sum of the magnetic resistances of the two air gaps 211 between the flapper and the magnetic pole, the central axis 210, the outer frame 213, and the bottom. FIG. 11 shows an example of the magnetic properties of the test material in this example, showing magnetic flux density B properties with respect to magnetizing force (magnetic field strength) H. FIG. The range of 0<H<Hc, where the magnetic flux density B increases in proportion to the magnetizing force H, is the linear region, and the range of H>Hc, where the gradient of the magnetic flux density B decreases significantly with respect to the magnetizing force H, is the magnetic saturation region. defined as Incidentally, Hc is obtained from the intersection of the BH characteristic envelope A in the range of 0<H<Hc and the BH characteristic envelope B in the range of H>Hc. Hc is defined as the magnetic force boundary value between the linear region and the magnetic saturation region, and the magnetic flux density when H=Hc is defined as the magnetic flux density boundary value Bc. For the magnetic material properties of FIG. 11, Hc=1500AT/m and Bc=1.5Wb/m2.
If the closed-loop magnetic circuit in which the magnetic flux Φ flows has a location where the magnetic path area Sc is extremely narrow, the magnetic flux density (B=Φ/Sc) is the highest at that location. That is, at this location, if the magnetizing force H exceeds a predetermined value, the magnetic flux density B is magnetically saturated. If B=Bmax at the time of magnetic saturation, the magnitude of the magnetic flux is suppressed within the range of Φ<S·Bmax.

In FIG. 10c, note the ring-shaped side surface (magnetic path area Sc=2πr1h) with radius r=r1 and thickness h. This portion is a source (or a source) of magnetic flux radially flowing through the disk, and is a portion for adjusting the magnetic saturation phenomenon (hereinafter referred to as maximum magnetic flux control surface). The magnetic flux density at this point is

上式から、最大磁束コントロール面の磁路面積Sc（=2πr1h）が極度に小さい場合、磁化力Hが所定の値を超えれば、すなわちH>Hcならば、グラフ（図１１）の曲線に沿って磁束密度、及び磁束は磁気飽和し、フラッパ（図１０bの２１２）に働く吸引力F（式3）も抑制される。

From the above formula, when the magnetic path area Sc (=2πr1h) of the maximum magnetic flux control surface is extremely small, if the magnetizing force H exceeds a predetermined value, that is, if H>Hc, then along the curve in the graph (Fig. 11) As a result, the magnetic flux density and magnetic flux are magnetically saturated, and the attractive force F (equation 3) acting on the flapper (212 in FIG. 10b) is also suppressed.

Figures 12 to 15 show magnetic flux density Br1, electromagnet attraction force F, flapper displacement X, and flow rate Q from the control port for current values when various disk plate thicknesses h are varied, using the analysis method described above. This is the analysis result. As for the analysis conditions, as described in the graph of FIG. 12, the value of the radius r3 of the flapper support portion is set so that the support rigidity of the flapper is constant. Focusing on the characteristics of the magnetic flux density Br1 with respect to the current value in FIG. 12,

a. When the disk thickness is thick and h=0.5 mm, as the current value increases, the magnetic flux density Br1 shows a non-linear characteristic that sharply increases from the vicinity of I=Ic (=0.017 A). In the region of I>Ic, the magnetic flux density Br1 is saturated and converges to a constant value Br1=1.7Wb/m2 (see FIG. 11).
b. When the disk thickness is thin and h=0.2 mm, the magnetic flux density Br1 with respect to the current value exhibits linear characteristics over the entire region.

above b. , the magnetic flux density Br1 in the region of 0<I<Ic is the above i. is larger than in the case of
After the magnetic flux density increases in the region of I>Ic, it is gradually suppressed when the current value reaches Ic, and the magnetic flux density Br1 has already increased to the same level as magnetic saturation. is in the magnetic saturation region. Therefore, (1) the magnetic path area (disk thickness) is reduced, and the magnetic flux density is increased from the stage where the current value is small. (2) Magnetic saturation begins in the vicinity of the current value Ic at which the magnetic flux density sharply increases. Due to the above (1) and (2), the magnetic flux density Br1 characteristic with respect to the current value smoothly shifts from the linear region to the magnetic saturation region, resulting in a characteristic with extremely excellent linearity. Magnetic flux density characteristics with excellent linearity are also reflected in the linearity of displacement characteristics, flow characteristics, and the like.

Using the above knowledge obtained in this study, the characteristics of the electromagnet attractive force, flapper displacement, and flow rate characteristics with respect to the current values in FIGS. 13 to 15 will be described. FIG. 13 shows the electromagnet attractive force (=restoring force of the disk spring) with respect to the current value when the plate thickness h of the disk is varied. The curve profile of the electromagnetic attractive force characteristic with respect to the current value substantially coincides with the curve profile (FIG. 14) of the flapper displacement characteristic with respect to the current value, which will be described later. In the flapper displacement characteristics with respect to the current value in FIG. This is because X0 is set to 0.25 mm. (Refer to FIG. 4) When h=0.2 to 0.3 mm, the flapper valve displacement characteristics (and flow rate characteristics) change almost linearly with respect to the current value, and ideal characteristics are obtained from the viewpoint of controllability. As mentioned above, this condition is the third condition required for the pneumatic servo valve applied to the active vibration isolation table: (3) “Linearity: The generated pressure is linearly proportional to the valve drive current. It is something that satisfies the "in". When h>0.35, as the current value increases, the flapper displacement sharply increases from the vicinity of I=Ic (=0.017A), exhibiting nonlinear characteristics.

FIG. 15 shows the outflow flow rate of the nozzle obtained based on the analysis conditions (shown in FIG. 14) for obtaining the graph of the flapper displacement characteristics. This corresponds to the servo valve of FIG. 1 in which the exhaust side passage 20 is blocked and the control side passage is opened to the atmosphere. The analysis conditions are supply pressure PS=0.6 MPa (abs), atmospheric pressure Pa=0.1 MPa (abs), and the nozzle diameter of the supply side nozzle 29 is Φ1.2 mm. From a comparison of FIGS. 14 and 15, it can be seen that the curve profiles of the flapper displacement characteristics and the nozzle flow characteristics are nearly identical.

In summary, by making the flapper (disk) a thin elastic body structure, a. Reduce the effective mass of the moving part to increase the resonance frequency. (See Table 1)
b. By using the magnetic saturation phenomenon, the linearity of the flow rate characteristic to the current is improved.
above a. b. brings about a synergistic effect.

2. Effect index of linearization and evaluation by actual measurement Here, the “effect index of linearization” in the displacement (flow rate) characteristic of the flapper with respect to the valve driving current is defined. FIG. 16 is a model diagram showing a method of obtaining the "linearization effect index" based on the actual measurement value of the flow rate with respect to the current value. The valve is driven in the range of current value: 0 (point a) < I < Imax (point d). In the case of the valve of this embodiment, since the magnetic attraction force is non-linear with respect to the current value, the attraction force rises gently with respect to the current value in a region where the current value is small, and increases steeply as the current value increases. However, when the current value further increases and enters the magnetic saturation region, the increase in magnetic flux Φ (and attractive force F) is suppressed. As a result, the profile of the flow rate characteristic (flapper displacement characteristic) with respect to the current becomes a downwardly convex curve in the region where the current value is low and upwardly convex in the region where the current value is high. Here, the inflection point E at which the "curve that is downwardly convex" changes to the "curve that is upwardly convex" is obtained from the intersection of the two envelopes Bb and Cd. Assuming that Aa is the curve of the flow characteristic with respect to the current, Bb (one-dot chain line) is the envelope of the downwardly convex region of the curve Aa. Cd (one-dot chain line) is the envelope of the upwardly convex region of the curve Aa. The inflection point E is the intersection of the envelopes Bb and Cd. Let c be the X-axis coordinate of the inflection point E, and b be the X-axis coordinate where the envelope Bb crosses the X-axis. Let F be the intersection of the curve Aa with the Y axis at I=Imax (point d). Let Da be the straight line (dashed line) connecting the intersection F and the origin (0, 0). The slope QE/Ibc (angle β) of the envelope Bb is the maximum value of the flow rate gain (ratio of flow rate to current) of this valve. The gradient QF/Iad (angle α) of the straight line Da is used as the reference value for the flow rate gain. Here, the ratio of the "reference value of the flow rate gain" to the "maximum value of the flow rate gain" is defined as the linearization effect index η as follows.

η=1のとき、曲線Aaは直線Daと一致して、電流に対して流量は正比例の関係となり、線形性の評価はベストとなる。

When η=1, the curve Aa coincides with the straight line Da, and the flow rate is directly proportional to the current, giving the best linearity evaluation.

Now, the reason why the flow rate characteristic with respect to the current of the servo valve is required to be linear is as follows. Since the servo valve is one element that constitutes the control system of the fluid servo device (active vibration isolation table), the ratio of the change in flow rate to the change in current: KQ = δQ/δI is the flow rate gain, and the open loop gain KL incorporated within. That is, if the gain of the control element other than the servo valve is KX and each element is combined, KL=KX·KQ. For example, as an example of margin setting using the frequency response method for stability, adjustment conditions such as (1) gain margin of 10 dB or more (2) phase margin of 45 deg or more are applied at the production site. If the maximum value of the flow rate gain of the servo valve is KQMAX near the current value I=Imax, the open loop gain KL to anticipate the stability margin of the entire active vibration isolation table must be determined by the above maximum value KQMAX. . However, the operating point of the servo valve is usually used near the middle position (I≈Imax/2) of the drive current range in many cases. Therefore, the more nonlinear the flow rate characteristic with respect to the current, the more excessive the gain margin is set at the operating point. In this case, the active anti-vibration table cannot exhibit its original sufficient performance at the operating point where it is used for the longest time. Therefore, the more linear the flow rate characteristic with respect to the current of the servo valve, the more suitable the stability (gain margin, phase margin) can be set for the control system.

Supplementally, it is preferable that the characteristics of the generated pressure (generated force) with respect to the current of the servo valve have linearity. This is a requirement when applying this servo valve to an active control device. As described above, the feedforward control is established only when the disturbance is known. A known stage behavior signal is used to perform the stage FF control. Feedback control improves the convergence time of the free vibration of the surface plate, but it is difficult to reduce the instantaneous response of stage acceleration/deceleration. In order to effectively cancel the linear motion disturbance using the stage FF control, it is necessary to create a highly accurate generated force waveform that reproduces the acceleration signal of the stage in the opposite phase. For this reason, the valve driving current waveform and the generated pressure waveform must be similar, that is, the characteristics of the generated pressure (generated force) with respect to the valve driving current must have linearity over a wide range centering on the operating point. There is

FIG. 17 shows an example of actually measuring the flow rate characteristics of the valve for obtaining the "linearization effect index". As for the basic structure of the valve, it is assumed that the two-nozzle type shown in the embodiment of FIG. 1, in which the forward nozzle and the reverse nozzle are arranged facing each other, is used. The "linearization effect index" can also be obtained from the displacement characteristics of the flapper against the current, but measurement of the flapper displacement is often not easy due to the structure of the valve. However, the flow rate characteristic of a valve having substantially the same profile with respect to the current value as the flapper displacement characteristic can be obtained by the method shown in FIG. 17 without dismantling the valve body.

In FIG. 17, 230 is a valve to be measured, 231 is a power supply for driving this valve, 232 is a control port, 233 is a supply port connected to a supply pressure source 234, 235 is an exhaust port, and 236 is a flow meter. By measuring the flow rate when the control port 232 is open to the atmosphere while the exhaust port 235 is shut off, it is possible to obtain the flow rate characteristics of the valve with respect to the current value, which is necessary for evaluating the "linearization effect index." . In the case of the embodiment of FIG. 1, it is the flow rate between one nozzle (in this case, the forward nozzle) arranged opposite to the flapper and the flapper. With this method, the "linearization effect index" can be obtained regardless of the detailed structure of the valve.

3. Evaluation Method Using Magnetic Saturation Phenomenon In this embodiment, the magnetic saturation phenomenon is used in the region where the displacement (flow rate) characteristic of the flapper rises steeply as the current value increases. It utilizes the fact that excellent characteristics can be obtained. Therefore, it is a precondition for applying this embodiment that any one of the elements constituting the closed loop magnetic circuit is magnetically saturated within the operating range of the valve. Even if the upper limit of the valve current is set before the flapper displacement rises sharply with respect to the current without using the magnetic saturation phenomenon, it can be applied as a servo valve. However, a large flapper displacement (flow rate) cannot be obtained.

In addition, by utilizing the magnetic saturation phenomenon, the flow rate characteristic with respect to the current near the maximum value of the current flowing through the electromagnet has an upwardly convex curve, that is, has an inflection point (point E in the model diagram 16). The following effects can be obtained by configuring as follows. 4, the gap (initial gap) between the flapper 24 and the magnetic pole 31 can be set with a margin. If the above magnetic saturation phenomenon is not used, the flapper displacement (flow rate) will rise sharply with respect to the current due to slight variations in the processing and assembly accuracy of the parts, the characteristics of the attractive force of the electromagnet, and the magnetic characteristics of the magnetic material. As a result, unstable valve characteristics tend to occur.

Here, assume the structure of a servo valve composed of element parts such as an electromagnet, a nozzle, and a flapper. At this time, the shape of each component part, the overall configuration of the valve, etc. are arbitrary. In order to embody a servo valve that utilizes the magnetic saturation phenomenon, applicability of the present invention is evaluated by the following method.

i. Find the total reluctance of the closed-loop magnetic circuit.
The magnetic resistance Ra between the nozzle flappers becomes minimum when the current maximum value I=Imax. At this time, Ra=.delta.n/(.mu.0S) where .delta.n is the distance between the nozzle flappers (see FIG. 4) and S is the magnetic pole area. Let RX be the sum of the linear magnetic resistances other than the magnetic resistance Ra, and the sum of the magnetic resistances of the closed-loop magnetic circuit is RS=Ra+RX. The linear magnetoresistance indicates the magnetoresistance when it is assumed 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 Emax=N×Imax, and the maximum value of the magnetic flux is Φmax=N×Imax/RS, 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.

iii. Find the magnetic flux density Bmax at a location where magnetic saturation is likely to occur.
In the closed loop magnetic circuit, (1) the point with the narrowest magnetic path area, or (2) the point where a magnetic material with the smallest saturation magnetic flux density is used, focusing on the above (1) and (2), the magnetic path area is Sc, magnetic flux density Bmax=Φmax/Sc.

iv. Evaluation of Occurrence of Magnetic Saturation Here, the "magnetic flux density characteristics (BH characteristics) with respect to the magnetizing force" of the magnetic materials used in the above (1) and (2) are used as evaluation data (see FIG. 11). The magnetic flux density boundary value Bc in the boundary region (magnetic force boundary value Hc) between the linear region and the magnetic saturation region is compared with the magnitude of Bmax. If Bmax<Bc, the magnetic saturation phenomenon does not occur and the magnetic circuit is used within the linear region. If Bmax>Bc, the magnetic saturation phenomenon occurs at the points (1) and (2) above, which means that the preconditions for applying the invention of this embodiment are satisfied.

<Third Embodiment>
FIG. 18 is a front cross-sectional view of a pneumatic servo valve according to Embodiment 3 of the present invention. 1 and 2 show a valve configuration in which the location for setting the spring stiffness of the disk is separated into two.

Reference numeral 110 denotes a central axis of the tubular shape, 111 denotes a bottom portion of the central axis, 112 denotes an outer frame formed concentrically with the central axis, 113 denotes a coil bobbin mounted on the central axis, and 114 denotes the coil bobbin. It is a wound coil. The central shaft 111, the outer frame portion 112, the coil bobbin 113, and the coil 114 constitute an electromagnetic actuator that attracts a face plate portion of a flapper (described later) and controls its displacement.

Reference numeral 115 denotes a cylindrical housing that houses the outer frame portion 12; 116, an exhaust-side bottom plate fastened to the side surface of the housing; 117, bolts for fastening the bottom portion 111 and the exhaust-side bottom plate 116 together; 118, the housing 115 and the exhaust-side bottom plate. A bolt 116 is fastened, 119 is an exhaust-side flow passage for gas (working fluid) formed in the central shaft 110, and 120 is a discharge port formed in the exhaust-side bottom plate 116. FIG. 121 is a supply-side bottom plate, 122 is a gas supply-side channel formed in the center of the supply-side bottom plate, and 123 is a gas control-side channel connected to a pneumatic actuator (not shown). Reference numeral 124 denotes a convex disk-shaped flapper which is mounted between the housing 115 and the supply side bottom plate 121 with bolts 125 .

The flapper 124 is composed of a thick convex portion 124a (magnetic path portion) and a thin peripheral portion (elastic support portion) 124b. Reference numeral 126 denotes a supply-side clearance formed between the supply-side bottom plate 121 and the flapper 124; 127, an exhaust-side clearance formed between the flapper 124 and the housing; Formed flow holes (128b and 128d are not shown in FIG. 18), 129 is a supply side nozzle (forward direction nozzle), and 130 is an exhaust side nozzle (reverse direction nozzle). Reference numeral 131 denotes a flapper-side end face of the central shaft 110 (first magnetic pole at the end face of the central shaft), 132 a flapper-side end face of the outer frame portion (second magnetic pole at the end face of the outer frame portion), and 133 a suction port.

FIG. 19 is a model diagram of a closed loop magnetic circuit formed by parts such as the central shaft 110, the flapper 124, and the coil 114 in FIG. In this model diagram, the supply side bottom plate 121, the supply side nozzle 129, etc. are omitted. The flapper of the valve of this embodiment has a protrusion with a radius of r2 and a thickness of h2. is the maximum flux control surface that regulates the The thickness h1 of the outer peripheral portion (r2<r<r3) of the flapper is sufficiently thinner than h2 and easily elastically deformed. Also, the outer radius r2 of the convex portion may be r2<r4 when compared with the outer radius r4 of the outer frame portion end surface 132 (second magnetic pole). The reason for this is that when the radius r=r1 is used as the magnetic saturation adjustment point, S1<S2 should be satisfied in the two magnetic path areas S1 (2πr1h2) and S2 (2πr2h1). In this case, since the section of the elastic deformation portion (r2<r<r3) can be sufficiently long, the thickness h1 need not be set to an extremely thin value.

FIG. 20 compares the flapper displacement with respect to the current value, assuming three types of flapper shapes with different plate thickness shapes and spring stiffnesses. FIG. 21 compares the generated force of the electromagnet with respect to the current value. From FIGS. 20 and 21, it is as follows when the current value is the same.

i. The magnitude of the displacement is C>B>A
ii. The magnitude of the generated force is C>A>B
The reason why A and B are reversed due to displacement and generated force is as follows. The reason why the generated force is A>B is that A, which is thicker than B, has less magnetic saturation. However, the reason why the displacement is A<B is that the spring stiffness of the disc is proportional to the cube of the plate thickness, and thus the spring stiffness is A>>B. When comparing C and A, which have the same plate thickness at the center of the disk, the reason why the generated force is C>A is that C deforms more easily than A, and the gap between the disk and the magnetic pole at the same current value is This is because it is smaller than A. From the above results, it can be seen that the convex disk can obtain a larger displacement at the same current value than the uniform thickness disk while maintaining good linearity.

FIG. 22 compares the control flow rate with respect to the current value when the exhaust port 120 is shut off in valves having the three types of flapper shapes (the structure of FIG. 18). The analysis conditions are supply pressure PS=0.6 MPa (abs), atmospheric pressure P0=0.1 MPa (abs), and nozzle diameters of the supply side nozzle 129 and the exhaust side nozzle 130 are both Φ1.2 mm. All of the valves having the three types of flapper shapes assume that the supply side nozzle 129 is shut off by the flapper 124 when the current value I=0. In order to block the exhaust side nozzle 130 when the maximum current value I=0.025 A, the projection amount .delta.n (see FIG. 4a) of the tip portion of the exhaust side nozzle 130 should be set as follows. Where Δa is the flapper displacement when I=0.025A, Type A has Δn=X0−Δa=0.222 mm, Type B has Δn=0.207 mm, and Type C has Δn=0.135 mm.

FIG. 23 shows a pneumatic servo valve according to Embodiment 3 of the present invention, in which a disc corresponding to a flapper is formed in a concave disk shape. 140 is a central axis, 141 is a flapper, 142 is a coil, 143 is a coil outer frame portion, and 144a and 144b are flapper fixing portions. The flapper 141 is composed of a thin plate thickness 141a (magnetic path portion) and a thin plate thickness peripheral portion (elastic support portion) 141b. If the disk has a concave disk shape, the plate thickness is thin at the center of the disk, so magnetic saturation suppresses the generation of attractive force. Since the plate thickness is thick at the outer circumference of the disk, the disk is less likely to deform. Depending on the application of the pneumatic servo valve, this structure can be applied, for example, when the disk deformation amount (flow rate) may be very small. Alternatively, this structure can also be applied to a case where a large number of circulation holes are formed in the elastic support portion 141b to reduce the rigidity even if the plate thickness is large (for example, the seventh embodiment).

<Fourth Embodiment>
24 is a front cross-sectional view of a pneumatic servo valve according to Embodiment 4 of the present invention. FIG.
A valve that can obtain a large flapper displacement (flow rate) with a small current value by forming a radial flow passage between the nozzle opening and the outer peripheral portion of the central shaft at the flapper side end surface (first magnetic pole) of the central shaft. It shows the configuration. 50 is the central axis of the cylindrical portion, 51 is the bottom of this central axis, 52 is the outer frame portion, 53 is the coil bobbin, 54 is the coil, 55 is the cylindrical housing, 56 is the exhaust side bottom of the housing, and 57 is the fastening. Bolts 58 and 59 are exhaust side passages. 60 is a supply side bottom plate, 61 is a supply side channel, and 62 is a control side channel connected to a pneumatic actuator (not shown). A convex disk-shaped flapper 63 is mounted between the housing 55 and the supply side bottom plate 60 with a bolt 64 . The flapper 63 is composed of a thick convex portion 63a (magnetic path portion) and a thin peripheral portion (elastic support portion) 63b. 64 is a supply side gap, and 65 is an exhaust side gap. 66a, 66b, 66c, and 66d are flow holes formed in the disk-shaped flapper (66b and 66d are not shown in FIG. 24), 67 is the opening of the supply side nozzle (forward direction nozzle), and 68 is the exhaust side. This is the opening of the nozzle (reverse nozzle). Reference numeral 69 denotes a flapper-side end face of the central shaft 50 (first magnetic pole at the central shaft end face), and 70 a flapper-side end face of the outer frame portion (second magnetic pole at the outer frame end face).

Fig. 25 is a partial enlarged view of the vicinity of the exhaust nozzle and the intake nozzle, Fig. 25a is an AA view of Fig. 25b showing a state in which the flapper blocks the intake nozzle, Fig. 25b is a BB arrow view of Fig. 25c, and Fig. 25c is a view of the intake nozzle. It shows a state in which a current is applied to the coil. In this embodiment, the exhaust nozzle and the intake nozzle are integrally formed by using the center shaft 50 and the supply side housing 60 instead of attaching separate parts as in the embodiment of FIG. Reference numerals 71a, 71b, and 71c denote circulation grooves formed on the flapper-side end surface of the first magnetic pole 69, 72 a depression formed between the exhaust nozzle opening 68 and the first magnetic pole 69, and 73 the outer peripheral side of the exhaust nozzle opening 68. It is a tapered portion formed in the The groove depth of the circulation groove was sufficiently deep and formed to 0.3 to 0.5 mm.

In this embodiment, the exhaust nozzle opening 68 is formed so as to protrude slightly from the end surface of the first magnetic pole 69 by .delta.n=0.046 mm (.delta.n is shown in FIG. 4a). Therefore, when the flapper 63 shields the exhaust nozzle opening 68, the gap between the first magnetic pole end face 69 and the flapper 63 becomes the narrow value of .delta.n. However, even in this case, the outer peripheral side (recessed portion 72) of the exhaust nozzle opening 68 and the exhaust-side gap 64 communicate with the circulation groove having a sufficiently deep groove depth, and the outer peripheral side of the exhaust nozzle opening 68 The pressure can be maintained at the same pressure as the exhaust-side gap 64 and the intake-side gap 64 .

FIG. 26 shows flapper displacement characteristics with respect to current values, and compares the following two cases.
i. The exhaust nozzle opening 68 is formed to protrude slightly from the end face of the first magnetic pole 69 by δn (initial gap X0 = 0.15 mm, nozzle protrusion amount δn = 0.046 mm in this embodiment).
ii. Make the exhaust nozzle opening protrude with a sufficient distance from the first magnetic pole (when X0 = 0.25 mm and δn = 0.135 mm in the structure of Embodiment 3)

ii. above. The reason for setting Δn=0.135 mm is as follows. If the nozzle protrusion amount δn is reduced, the air resistance in the radial flow path increases, and the entire air gap with the total area S of the outer diameter of the first magnetic pole 69 may approach atmospheric pressure (PS = 0.1 MPa). There is In this case, force f=(Pa-PS)S proportional to the pressure difference between the left and right sides of the flapper is applied to the flapper. As a result of the experiment, it was found that the flapper was brought into close contact with the first magnetic pole surface, which hindered the current control of the flapper displacement (control pressure). Therefore, it is preferable to set the value of the protrusion amount Δn while maintaining a sufficient distance. ii. above. Under the condition of , a current value I = 0.025 A is required to reach (point B) the displacement X = 0.102 mm (control pressure Pa = 0.6 MPa at this time). above i. In this embodiment, stable flapper displacement current control is possible even if Δn is sufficiently small, and the current value at which the same displacement (same control pressure) is reached (point A) is I=0.015A.

<Fifth Embodiment>
27 is a front cross-sectional view of a pneumatic servo valve according to Embodiment 5 of the present invention,
By bringing the flapper valve-side end surface of the outer frame into close contact with the flapper surface, the second magnetic pole is omitted, and a magnetic circuit is formed so that the first magnetic pole alone can obtain the attraction to the flapper. With this configuration, the outer diameter of the elastic support portion of the convex flapper can be reduced, so that the outer diameter (ΦD) of the servo valve main body can be reduced. For example, in the case of an active vibration isolation table, multi-axis pneumatic actuators are attached to the pneumatic units that support the four corners of the stage. Since the pneumatic actuator and the control port of the servovalve need to be arranged close to each other, it is preferable to make the outside diameter (ΦD) of the servovalve body as small as possible.

Reference numeral 150 denotes a cylindrical central shaft, 151 a bottom portion of the central shaft, 152 an outer frame portion of the central shaft, 153 a coil bobbin, 154 a coil, 155 a cylindrical housing, 156 a bottom portion of the housing, and 157 a bottom portion of the housing. 158 is an exhaust-side flow passage, 159 is a discharge port, 160 is a supply-side housing, 161 is a supply-side flow passage, 162 is a control-side flow passage connected to a pneumatic actuator (not shown), and 163 is a convex disc shape. The flapper 163 is composed of a convex portion (magnetic path portion) 164 with a thick plate and a peripheral portion (elastic deformation portion) 165 with a thin plate. 166 is a supply-side gap, 167 is an exhaust-side gap, 168a, 168b, 168c, and 168d are circulation holes formed in the flapper 163 (168b and 168d are not shown), and 169 is a supply-side nozzle (forward nozzle). An opening 170 is an exhaust side nozzle (reverse direction nozzle) opening, 171 is a magnetic pole of an electromagnet, and 172 is a flapper side end surface of the outer frame portion 152 of the central shaft, which is in close contact with the flapper 163 . 173 is an intake port.

<Sixth Embodiment>
FIG. 28 is a front cross-sectional view of a pneumatic servovalve according to Embodiment 6 of the present invention. By providing a ring-shaped second magnetic pole with a diameter smaller than the outer diameter of the coil in the closed-loop magnetic circuit, the magnetic attraction force The outer diameter (ΦD) of the servo valve body is reduced while maintaining the That is, the effect of magnetic saturation (decrease in attractive force) of the elastically deformable portion due to the miniaturization of the valve body is eliminated. Reference numeral 250 denotes a cylindrical center shaft, 251 a bottom portion of the center shaft, 252 an outer frame portion of the center shaft (support shaft), 253 a coil bobbin, and 254 a coil. 255 is a cylindrical housing, 256 is the bottom of this housing, 257 is a fastening bolt, 258 is an exhaust side flow path, 259 is a discharge port, 260 is a supply side housing, 261 is a supply side flow path, and 262 is a pneumatic actuator (not shown). ) is the control-side flow path connected to Reference numeral 263 denotes a convex disk-shaped flapper, which is composed of a thick convex portion (magnetic path portion) 264 and a thin peripheral portion (elastic deformation portion) 265 . 266 is a supply side gap, and 267 is an exhaust side gap. 268a, 268b, 268c, and 268d are circulation holes (268b and 268d are not shown) formed in the flapper 263, 269 is a supply side nozzle (forward direction nozzle) opening, and 270 is an exhaust side nozzle (reverse direction nozzle) opening. 271 is the first magnetic pole of the electromagnet; 272 is the flapper side end surface of the outer frame 252 of the central axis; 273 is a fastening bolt; A magnetic pole ring 275 is held between the flapper side end face 272 of the outer frame portion 252 and the housing 255, and a second magnetic pole 276 is formed on the flapper 263 side end face of the magnetic pole ring. A portion 277 of the housing 255 interposed between the magnetic pole ring 275 and the flapper 265 is made of a non-magnetic material. 278 is an intake port.

FIG. 29 is an enlarged view of the closed loop magnetic circuit of the above-described fifth embodiment in which only one magnetic pole is provided, and FIG. be. In the case of FIG. 29, the closed loop magnetic circuit passes through the route of "center axis 150→magnetic pole 171→magnetic path portion 164→elastic deformation portion 165→outer frame portion 152".

In the case of this embodiment shown in FIG. 30, the closed loop magnetic circuit passes through the route of "center axis 250→first magnetic pole 271→magnetic path portion 264→second magnetic pole 276→magnetic pole ring 275→outer frame portion 252". The difference between the two closed loop magnetic circuits is that the magnetic circuit in FIG. 29 passes through the elastically deforming portion 165, whereas the magnetic circuit in FIG. be.

FIG. 31 compares "flapper displacement characteristics with respect to current values" assuming the above two magnetic circuits. As shown in the figure, each disk has the same shape and rigidity.
Type A is the case of this embodiment (the structure of FIG. 30), and Type B is the case of using the structure of FIG.
When the outer diameter of the disk is reduced in order to reduce the size of the valve, the radial length of the elastically deformable portion must also be reduced. Therefore, in order to maintain low rigidity, the analysis results are obtained under the condition that the plate thickness of the elastically deformed portion is reduced to h1=0.08 mm.
At a current value of I=0.025 A, the Type A flapper maximum displacement is Xmax=0.13 mm, while the Type B flapper maximum displacement is only Xmax=0.018 mm. The reason is,
i. In the case of Type B, the magnetic circuit passes through a thin plate thickness h1. As a result, however, the magnetic flux passes through a path with a narrow magnetic path area, and the maximum magnetic flux is greatly suppressed due to the influence of magnetic saturation.
ii. In the case of Type A, the magnetic flux skips the elastically deformable portion 265 and has a thin plate thickness h1, and draws a route of the magnetic path portion 264→the second magnetic pole 276→the outer frame portion 252. FIG. Therefore, the magnitude of the magnetic flux flowing through the closed loop magnetic circuit is not affected by the thin elastic deformation portion 265 .

That is, in this embodiment, the structural design for determining the spring rigidity of the disk portion and the magnetic circuit design for determining the attraction force characteristic with respect to the current value can be performed independently.

In the fifth and sixth embodiments described above, the rigidity is adjusted by changing the plate thickness h1 of the elastic deformation portion. However, the rigidity can be adjusted by forming an appropriate gap in the flapper. 32, 280 is a flapper, 281 is an elastic deformation portion, 282 is a gap portion, 283 is a fastening bolt, and 284 is a magnetic path portion.

Alternatively, as means for adjusting the rigidity of the flapper, for example, a plurality of small holes may be formed in the flapper axisymmetrically in the circumferential direction. (not shown)

<Seventh Embodiment>
Fig. 33 shows a pneumatic servo valve according to Embodiment 7 of the present invention, Fig. 33a is an AA view of Fig. 33c, Fig. 33b is a partially enlarged view of Fig. 33a, Fig. 33c is a sectional front view, and Fig. 34 is a 33c is a partially enlarged view of the spiral disk spring portion of 33c. FIG. In this embodiment, the rigidity of the elastic deformation portion of the flapper is selected not by the plate thickness but by the shape of the spiral formed in the flapper. 350 is a central shaft (supporting shaft), 351 is a bottom portion of this central shaft, 352 is an outer frame portion of the central shaft, 353 is a coil bobbin, and 354 is a coil wound around the coil bobbin. The central shaft 350, the central shaft bottom portion 351, the outer frame portion 352 of the central shaft, the coil bobbin 353, and the coil 354 constitute an electromagnetic actuator that attracts a face plate portion of a flapper (described later) and controls its displacement. 355 is a cylindrical housing, 356 is the bottom of this housing, 357 is a bolt, 358 is an exhaust side flow path, 359 is a discharge port, 360 is a supply side housing, 361 is a supply side flow path, 362 is a pneumatic actuator (not shown). ) is a control-side flow path connected to A disk-shaped flapper 363 is composed of a central magnetic path portion 364 and an elastic deformation portion 365 formed with a spiral disk spring (described later). 366 is a supply side gap formed between the supply side housing 360 and the flapper 363, and 367 is an exhaust side gap formed between the flapper 363 and the housing side. 368 is a supply side nozzle (forward direction nozzle) opening, and 369 is an exhaust side nozzle (reverse direction nozzle) opening. Reference numeral 370 denotes a flapper-side end face (center shaft end face) of the central shaft 350, and a magnetic pole of the electromagnet, and 371 a suction port.

In this embodiment, the spiral disk spring, which is the elastically deformable portion 365, is composed of eight ridges (peaks) and the same number of grooves (voids). In the partially enlarged view of the spiral disk spring portion in FIG. 34, 372a, 372b, 372c, and 372d are the air gap portions of the spiral disk spring (elastic deformation portion 365), and 372d is the air gap formed near the magnetic pole 370 and having the largest opening area. Department. Now, the spiral disk spring (elastic deformation portion 365) in this embodiment will be described as follows: i. ~iii. It also plays the role shown in .
i. Relieve generated stresses to obtain adequate flapper support stiffness ii. A groove (gap portion) is used to form a flow path connecting the supply side gap portion 366 and the exhaust side gap portion 367 iii. Let it be the magnetic path of the closed loop magnetic circuit

above i. The effect of is as follows. In order to reduce the outer diameter (ΦD) of the servo valve main body, the flapper is formed in a convex shape, the plate thickness of the elastic support portion is reduced as much as possible to reduce rigidity, and a second magnetic pole is provided (6th embodiment). In this case, as the plate thickness becomes thinner, the stress generated in the elastic support portion increases, and there is a problem that the allowable stress (elastic limit) of the disk-shaped flapper member is exceeded. By using a spiral disk spring as the convex member, it is possible to greatly reduce the maximum generated stress. The rigidity and generated stress of the spiral disk spring can be selected by the spiral angle α (Fig. 33a), the number of grooves (ridges), the width ratio of grooves and ridges, etc., in addition to the plate thickness. However, since the boundary line between the magnetic path portion 364 and the ridge forms an acute angle at the starting point of the spiral curve, stress concentration occurs. In order to reduce this stress concentration, curved surface portions 373 and 374 different from the original spiral curve are formed as shown in FIG. 33b. It has been found that the formation of this curved surface portion can significantly alleviate the stress concentration.

ii. above. In the case of , the gaps 372 a , 372 b , 372 c , and 372 d can also serve as flow passages connecting the supply-side gap 366 and the exhaust-side gap 367 . The gap 372d formed by the curved surface portions 373 and 374 to relieve stress concentration can secure the largest opening area. The flow of air is indicated by arrows (solid lines) in FIG.

above iii. utilizes the fact that even if the plate thickness of the spiral disk spring is sufficiently thick, the rigidity can be selected according to the shape. Since the plate thickness is large and the magnetic path area can be increased, magnetic saturation does not occur, and the second magnetic pole need not be formed as shown in the sixth embodiment. As a result, simplification of the valve structure main body can be achieved. As indicated by arrows (dashed lines) in FIG. 34, the flow of magnetic flux is "magnetic pole 370→ridge (peak) of spiral disk spring→outer frame 352 of center shaft". In the embodiment, the plate thickness of the elastic deformation portion 365 (spiral disk spring) is set to be the same as that of the magnetic path portion 364, but it may be set to be thicker than that of the magnetic path portion 364. In this case, as shown in FIG. 23, the profile of the entire flapper is concave. FIG. 35 shows structural analysis results of deformation of the spiral disk spring when a load F is applied to the central portion of the magnetic path portion 364 .

A known cloud spring as shown in FIG. 36 may be used as the disk spring that can be used for the elastic deformation portion. 380 is a flapper, 381 is an elastic deformation portion, 382a and 382b are arcuate gaps, 383 is a magnetic path portion, and 384 is a fastening bolt.

<Eighth embodiment>
In the embodiments of the present invention described above, the valve structure is mainly composed of axially symmetrical parts. In addition to the axially symmetrical parts described above, the servo valve according to the present invention can also be formed by combining various iron cores such as prismatic, cylindrical, horseshoe-shaped, and annular iron cores, rectangular thin plates, square blocks, and the like to form magnetic circuits and fluid circuits. realizable.

Figure 37 shows a pneumatic servo valve according to Embodiment 8 of the present invention, Figure 37a is a top view and Figure 37b is a front cross-sectional view. Reference numeral 400 denotes a support shaft; 401, an electromagnetic coil mounted on this support shaft; 402, the bottom of an L-shaped member; 406 is a bolt for fastening the L-shaped member bottom portion 402 and the support shaft 400 . The support shaft 400, the electromagnetic coil 401, the L-shaped member bottom portion 402, the L-shaped member upright member 403, and the flapper 404 constitute an electromagnetic actuator (actuator section) that attracts the flapper and controls its displacement. A closed loop magnetic circuit is formed by "support shaft 400→L-shaped member bottom portion 402→L-shaped member upright portion 403→flapper 404→support shaft 400". 406 is an exhaust-side flow passage formed in the L-shaped member upright member 403, and 407 is an exhaust-side nozzle (reverse nozzle) mounted on the flapper side of this exhaust-side flow passage. 408 is a supply-side block, 409 is a supply-side channel formed in this supply-side block, 410 is a supply-side nozzle (forward nozzle) mounted on the flapper side of this supply-side channel, and 411 is a supply-side block. 412 is a supply-side gap, 413 is an exhaust-side gap, and 414 is a control port. The two spaces of the supply-side gap 412 and the exhaust-side gap 413 are connected, and these two spaces form the control chamber 415 of this servo valve. The control chamber is connected to a pneumatic actuator (not shown) via control port 414 . The flapper, the supply-side nozzle, the exhaust-side nozzle, the supply-side channel, the exhaust-side channel, the supply-side block, and the L-shaped member upright portion constitute a fluid control section. 416 is the flapper side end face of the L-shaped member upright portion 403, which is the magnetic pole of the electromagnet. 417 is a suction port, and 418 is a discharge port.

In this embodiment, as in the case of the first embodiment, a pair of nozzles are arranged on the front and rear sides of the flapper on the center line of the magnetic pole to form a nozzle flapper valve with a two-way flapper. Further, air is used as the fluid, and the actuator section and the fluid control section are arranged in the same space through which the fluid penetrates. The flapper is composed of a flat plate-shaped member as a part of a closed loop magnetic circuit, and is proportional to the gap between the nozzle and the flapper so as to balance the attractive force of the electromagnet by utilizing the elasticity of the flapper itself. The flapper is provided with a restoring force.

Furthermore, the magnetic path area of the flapper (Sc = b x h ) is selected. Therefore, similarly to the second embodiment, it is possible to obtain a flapper displacement (flow rate) characteristic with respect to a current value with good linearity.

Instead of selecting the magnetic path area of the flapper as the maximum magnetic flux control surface for adjusting the magnetic saturation phenomenon, for example, as in the second embodiment, the outer diameter (or the thickness of the cylindrical portion) of the annular magnetic pole 415 is selected. It may be configured to be as small as possible.

In this embodiment, since the flapper has a cantilever support structure, the valve structure can be simplified. However, if the flapper has a structure that supports both ends, i. the required stiffness to balance the magnetic attraction, ii. Appropriate magnetic path area when using the magnetic saturation phenomenon, iii. Unstable vibration caused by fluid force from the nozzle, above i. ~iii. A more appropriate flapper shape can be selected for (not shown)

<Ninth Embodiment>
This embodiment is an improvement of the eighth embodiment. Two magnetic poles are provided on the flapper-side end surface of the electromagnet, and the closed loop is provided between the flapper-side end surface of the first magnetic pole and the flapper-side end surface of the second magnetic pole. A magnetic path portion serving as the main path of the magnetic circuit is formed, and an elastic support portion is provided on the fixed side of the flapper. According to this embodiment, as in the sixth embodiment, which is an axially symmetrical valve, the supporting rigidity of the flapper can be set independently of the magnetic attractive force characteristics, so that a larger flapper displacement (flow rate) can be obtained with the same current value. can.

38 is a front cross-sectional view of a pneumatic servo valve according to Embodiment 9 of the present invention. FIG.
430 is the support shaft, 431 is the electromagnetic coil, 432 is the bottom of the L-shaped member, 433 is the upright portion of the L-shaped member, 434 is the flapper, 435 is the fastening bolt, 436 is the exhaust side flow passage, and 437 is the exhaust side nozzle (reverse direction nozzle). ). 438 is a supply-side block, 439 is a supply-side channel, 440 is a supply-side nozzle (forward nozzle), 441 is a fastening bolt, 442 is a supply-side gap, 443 is an exhaust-side gap, 444 is a control port, and 445 is a control port. Control room, 446 is the first pole of the electromagnet. The point that the control chamber is connected to a pneumatic actuator (not shown) via a control port 444 is the same as the above-described embodiment. 447 is a magnetic pole yoke material, 448 is a second magnetic pole formed on the flapper side end surface of this magnetic pole yoke material, 449 is a spacer made of a non-magnetic material, and 450 is the magnetic pole yoke material and the flapper via the spacer. to the supply side block 438, and 451 is an elastically deforming portion formed in the flapper between the second magnetic pole and the fixed side of the flapper. That is, concave portions are formed on the front and back surfaces of the flapper to constitute the elastic deformation portion. Further, 452 is a magnetic path portion between the supply side nozzle and the second magnetic pole in the flapper. Reference numeral 453 denotes a suction port, and 454 denotes a discharge port. A closed-loop magnetic circuit is formed by "material 447→support shaft 430".

<Tenth Embodiment>
This embodiment is an improvement over the ninth embodiment. By fixing and supporting the flapper at both ends, even if the plate thickness of the elastic deformation portion of the flapper is thin and the rigidity is small in the initial state where no current is applied to the electromagnet, , the magnetic gap and the air gap can always be kept constant without being affected by aging.

39 shows a pneumatic servo valve according to Embodiment 10 of the present invention, where FIG. 39a is a top view and FIG. 39b is a front sectional view. 460 is a support shaft, 461 is an electromagnetic coil, 462a and 462b are W-shaped member bottoms, 463a and 463b are W-shaped member upright portions, 464 is a flapper, 465a and 465b are fastening bolts, 466 is an exhaust side flow passage, and 467 is an exhaust. It is a side nozzle (reverse direction nozzle). 468 is a supply-side block, 469 is a supply-side channel, 470 is a supply-side nozzle (forward nozzle), 471a and 471b are supply-side gaps, 472a and 472b are exhaust-side gaps, 473 is a control port, and 474 is a control. Chamber 475 is the first pole of the electromagnet. The connection from the control chamber to a pneumatic actuator (not shown) via a control port 473 is the same as in the previous embodiment. 476a and 476b are yoke materials for magnetic poles, 477a and 477b are second magnetic poles formed on the flapper side end faces of the yoke materials for magnetic poles, 478a and 478b are spacers made of a non-magnetic material, and 479a and 479b are via the spacers. Bolts 480a and 480b for fastening the magnetic pole yoke material and the flapper to the supply side block 468 are elastic deformation portions formed in the flapper between the second magnetic pole and the fixed side of the flapper. That is, concave portions are formed on the front and back surfaces of the flapper to constitute the elastic deformation portion. In the flapper, 481a and 481b are magnetic path portions formed between the supply side nozzle and the second magnetic pole. 482 is a suction port, and 483 is a discharge port. The closed loop magnetic circuit of only the right half is "support shaft 460→first magnetic pole 475→magnetic path portion 481b of the flapper→second magnetic pole 477b→magnetic pole yoke material 476b→L-shaped member upright portion 463b→W-shaped member bottom portion 462b. → support shaft 460”.

<Eleventh Embodiment>
In each of the above-described embodiments, a flow path is provided in the closed-loop magnetic circuit that constitutes the actuator section to connect "the gap between the nozzle and the nozzle flapper to the control chamber". Therefore, the members of the actuator section and the fluid control section are partly shared. In this embodiment, the flapper is provided by extending from the actuator portion, and a flow path is constructed in which the nozzles are arranged to face the extended flapper surface.

FIG. 40 is a front sectional view of a pneumatic servo valve according to Embodiment 11 of the present invention.
800 is a support shaft, 801 is an electromagnetic coil, 802 is an L-shaped member bottom, 803 is an L-shaped member upright portion, 804 is a flapper, 805 is an upper support member, 806 is a bolt for fastening the support shaft and the upper support member, A bolt 807 fastens the support shaft and the flapper.

A closed loop magnetic circuit is formed by "support shaft 800→L-shaped member bottom portion 802→L-shaped member upright portion 803→flapper 804→support shaft 800". The following fluid control unit drawn in imaginary lines is provided at a location separate from the closed loop magnetic circuit. 808 is an exhaust-side block, 809 is an exhaust-side flow passage, and 810 is an exhaust-side nozzle (reverse nozzle). 811 is a supply-side block, 812 is a supply-side channel, 813 is a supply-side nozzle (forward nozzle), 814 is an intake port, and 815 is a discharge port. 816 is a control room, which communicates with a control port (not shown). 817 is a magnetic pole, but the two nozzles (810, 813) are not arranged on the center line of this magnetic pole. As in other embodiments, the magnetic pole and the nozzle may not be arranged on the same axis, but may be arranged adjacent to each other, for example. Even in this case, there is no problem in applying the present invention, and any one of them may be selected in consideration of the overall structure of the valve. In the case of this embodiment, the two nozzles (810, 813) are arranged outside the magnetic pole with respect to the fastening portion 807 (the bolt) of the flapper, so the displacement (flow rate) between the nozzle and the flapper is It can be obtained by amplification.

<Twelfth Embodiment>
In the above-described embodiments, only electromagnets form a closed loop magnetic circuit to drive the flapper. However, it is also possible to simplify the configuration of the fluid servo valve by using both an electromagnet and a permanent magnet to increase the attractive force of the flapper and by integrating the fluid control section and the actuator section. The combined use of an electromagnet and a permanent magnet can also be applied to the valves of the first to seventh embodiments, which have axially symmetrical structures.

In FIG. 41, 951 is a magnet (permanent magnet), 952 is a coil, 953 is a body that houses this coil, 954 is a flapper, 955a and 955b are a pair of yokes attached with their tips facing each other, and 956 is the tip of the flapper. is. 957 is a leaf spring which is a flapper supporting member, and 958 is a support center of the leaf spring. 959 is a supply side nozzle (forward direction nozzle), 960 is an exhaust side nozzle (reverse direction nozzle), 961 is a supply side flow path, 962 is an exhaust side flow path, 963 is a control port, 964 is a control chamber, and 965 is an intake port. , 966 are outlets. The supply-side nozzle and the supply-side channel, and the exhaust-side nozzle and the exhaust-side channel are formed using the pair of yokes.

<Thirteenth Embodiment>
This embodiment proposes a valve structure that solves the problem associated with the driving principle of the conventional "nozzle flapper valve using a two-way flapper", that is, the drawback that the air consumption flow rate is the largest at the operating point of the valve in the steady state. is.

FIG. 42 is a front cross-sectional view of a pneumatic servovalve according to Embodiment 13 of the present invention, wherein 750 is a cylindrical central shaft, 751 is the bottom of this central shaft, and 752 is concentric with the central shaft. An outer frame portion of the central shaft formed by a circle, 753 a coil bobbin mounted on the central shaft, and 754 a coil wound around the coil bobbin. The central shaft 750, the central shaft bottom portion 751, the outer frame portion 752 of the central shaft, the coil bobbin 753, and the coil 754 constitute an electromagnetic actuator that attracts a flapper (described later) and controls its displacement.

Reference numeral 755 denotes a cylindrical housing that accommodates the bottom portion 251 of the central shaft and the outer frame portion 752; 756, the bottom portion of the housing; 757, bolts for fastening the housing bottom portion 756 and the central shaft bottom portion 751; An exhaust-side flow passage 759 is a discharge port formed in the central shaft 750 . 760 is a supply-side housing, 761 is a supply-side channel formed in the center of the supply-side housing, and 762 is a control-side channel connected to a pneumatic actuator (not shown). Reference numeral 763 denotes a convex disk-shaped flapper, which is composed of a thick convex portion (magnetic path portion) 764a and a thin peripheral portion (elastic deformation portion) 764b. 765 is an intake port, 766 is a supply side gap formed between the supply side housing 260 and the flapper 763, and 767 is an exhaust side gap formed between the flapper 763 and the housing side. 768a, 768b, 768c, and 768d are circulation holes (768b and 768d are not shown) formed in the flapper 763, 769 is a supply side nozzle (forward direction nozzle), and 770 is an exhaust side nozzle (reverse direction nozzle). 771 is the first magnetic pole of the electromagnet on the flapper valve side end face (center shaft end face) of the center shaft 750 . A second magnetic pole 772 is formed on the flapper-side end surface of the outer frame portion 752 , and a non-magnetic ring 773 is held between the flapper 763 and the housing 755 . The supply-side housing 760 and the housing 755 sandwich the flapper and the non-magnetic ring and are fastened with bolts (not shown). Reference numeral 774 denotes a supply-side projection formed at the center of the flapper on the supply side, and 775 denotes an exhaust-side projection formed at the center of the flapper on the exhaust side. A control chamber 776 of the valve is formed by the supply-side gap and the exhaust-side gap.

The valve of this embodiment can sufficiently reduce the air consumption flow rate at the operating point of the valve in the steady state. This is because annular passage forming structures are provided between the supply-side nozzle 769 and the flapper 763 and between the exhaust-side nozzle 770 to form a passage having a substantially annular cross section. to cause. More specifically, the annular flow path forming structure is composed of a cylindrical inner peripheral surface at the tip of each nozzle 769, 770 and an insert inserted radially away from the inner peripheral surface. It will be. In other words, the protruding portion protruding perpendicularly to the face plate portion of the flapper 763 is used as an insert. properties can be changed. Figures 43a to 43c show the combination state between the nozzle flappers, and Figures 44a and 44b model the relationship between the valve flow rate and the gap between the nozzle flappers when focusing on one nozzle. 43a to 43c, the principle of operation of this valve will be described below while comparing the above two figures (FIGS. 44a and 44b).

43a shows the state of the valve input current I=0 (initial value), FIG. 43b shows the state of the input current I≈Imax/2 (operating point), and FIG. 43c shows the state of the input current I=Imax (maximum value). In the figure, 777 is a supply side nozzle orifice, and 778 is an exhaust side nozzle orifice. 779a, 779b (not shown), and 779c are circulation grooves formed in the flapper-side end face of the first magnetic pole 771, and have the same function as in the fourth embodiment (FIG. 25b).

At valve input current I=0 in FIG. 44a, the flow in the narrow annular gap formed by the supply-side protrusion 774 and the supply-side nozzle orifice 777 is a viscous flow. Therefore, the amount of air flowing into the control chamber of this valve from an air pressure source (not shown) is very small.

At the input current I≈Imax/2 (operating point) in FIG. FIG. 44b is view B in FIG. 44a, wherein the flow of fluid entering the control chamber from the supply nozzle is in a transition region transitioning from a viscous flow regime to a potential flow regime. Fluid flow from the control chamber into the exhaust nozzle is also in the transition region.

At the input current I=Imax in FIG. FIG. 44a, view C, the flow of fluid entering the control chamber from the supply nozzle is in the potential flow regime. Further, the exhaust-side convex portion 775 of the flapper penetrates deeply into the exhaust-side nozzle orifice 778, and the flow in the narrow annular gap formed by both members is a viscous flow. Therefore, the amount of air flowing out from the control chamber 776 to the atmosphere is very small.

Fig. 44b shows the characteristics (solid line) of the "valve flow rate Q with respect to the gap X between the nozzle and the flapper (valve input current)" of the valve of this embodiment in comparison with the characteristics of the conventional valve (chain line). In the case of the valve of this embodiment, the flow rate is sufficiently small from the viscous flow region A to the transition region B, and when entering the potential flow region C, the flow rate sharply increases. In the case of the conventional valve, since the entire region (A→B→C) is the potential flow region, the flow rate increases from the stage where the gap X is small. The difference in this flow rate characteristic becomes the difference in the air consumption flow rate at the two operating points.

45a and 45b schematically show that the valve of this embodiment can greatly reduce the air consumption flow rate at the operating point compared to the conventional valve. Assuming a constant pressure in the control chamber, the characteristics of flow rate Q with respect to gap X are: i. Inflow from the supply side to the control room (solid line), ii. The outflow from the control room to the atmosphere (dashed line) is indicated. FIG. 45a is the conventional valve, and FIG. 45b is the valve of this embodiment. Assuming that the intersection of the inflow and outflow is the operating point, it can be seen that the air consumption flow rate at the operating point of the valve of this embodiment is significantly smaller than that of the conventional valve. Near the operating point, curve i. ii. are both upwardly convex curves, and the curve i. ii. are a combination of downward convex curves. In other words, regardless of the detailed structure of the nozzle and flapper, or whether it is in the viscous flow region or the potential flow region, the characteristic of the flow rate Q with respect to the gap X (or current value) is convex downward near the operating point. It is the point of reducing the consumption air flow rate to have a valve characteristic that follows the curve of .

As mentioned above, the reason why the valve of this embodiment can greatly reduce the air consumption rate is that the fitting state between the protrusions on both sides of the two-way flapper and the orifices on the nozzle side can be adjusted by moving the flapper in the axial direction. is. For this purpose, it is preferable that the flapper is driven with a stroke that is as large as possible from the viewpoint of structure and processing of members. However, as shown in the first embodiment, for example, in the case of an actuator using Maxwell's stress, the maximum value of the magnetic gap between the magnetic pole and the flapper where the magnetic attraction can be effectively used is on the order of 0.05 to 0.20 mm. rice field. The characteristic of the magnetic attraction force with respect to the air gap is non-linear, and above the maximum value the magnetic attraction force usually falls off significantly. However, as described in the third embodiment, if an appropriate magnetic material and a thin disk are used for the movable portion corresponding to the flapper, the displacement characteristic of the flapper with respect to the current exhibits excellent linearity (linearity). I was able to find out what I could get in the process of this research.

By further actively utilizing this magnetic saturation phenomenon, the stroke of the flapper can be greatly increased without losing the linearity of the displacement characteristic of the flapper with respect to current.

Specifications of the electromagnet and disk shape used in this embodiment are shown in the graph of FIG.
[Type(II)] is shown in comparison with the specification [Type(I)] of the second embodiment described above. The type (II) electromagnet has twice the outer diameter of the type (I) and has three times the number of coil turns. When the current value I=40 mA, the flapper displacement X is approximately 0.12 mm in Type (I), while the flapper displacement X is 0.68 mm in Type (II) of the present embodiment. In the development of the valve of this embodiment, as a result of examination from the aspects of structure and performance and precision machining of parts, the maximum stroke between the nozzle and the electromagnet was set to 0.5 mm or more by using the magnetic saturation characteristics of the closed loop magnetic circuit. I found that setting it to 0 gives good performance.

FIG. 47 uses the structure of the above-described embodiment and devises the fitting state between the nozzle flappers in order to make the flow rate characteristic with respect to the current value more linear. That is, the flapper-side convex portion that fits with the nozzle is tapered so that the flow passage area between the nozzle flappers with respect to the current value changes more smoothly. 47a shows the valve input current I=0 (initial value), FIG. 47b shows the input current I≈Imax/2 (operating point), and FIG. 47c shows the input current I=Imax (maximum value). 780 is a supply-side housing, 781 is a central shaft, 782 is a flapper, 783 is a supply-side tapered portion formed in the central portion of the flapper, 784 is an exhaust-side tapered portion, and 785 is fitted with the supply-side tapered portion. Formed supply side nozzle (forward direction nozzle) orifice 786 is an exhaust side nozzle (reverse direction nozzle) orifice formed to fit with the exhaust side taper portion. Reference numeral 787 denotes a first magnetic pole, 788a, 788b (not shown), and 788c denote flow grooves formed on the end face of the first magnetic pole 787 on the flapper side, and have the same function as in the fourth embodiment (FIG. 25). 789 is a control room.

At valve input current I=0 in FIG. 47a, the supply taper 783 of the flapper penetrates deeply into the supply nozzle orifice 785 and flows from a pneumatic supply (not shown) into the control chamber 789 of the valve. The amount of air is minute.

At the input current I≈Imax/2 (operating point) in FIG. The same applies to the relationship between the exhaust-side tapered portion 784 and the opening end of the exhaust-side nozzle orifice 786 . Therefore, at this stage, the flow rate from the supply side to the control chamber 789 and the flow rate from the control chamber to the atmosphere side are both very small.

At the input current I=Imax of FIG. 47c, the flapper exhaust tapered portion 784 penetrates deeply into the exhaust nozzle orifice 786 and the amount of air flowing out of the control chamber 789 to atmosphere is very small.

In this embodiment, a convex disk-shaped flapper is used, but if a spiral disk spring is used as in the seventh embodiment, the generated stress can be relaxed despite the large axial displacement of the flapper. , and appropriate axial rigidity can be set. Alternatively, for example, cloud springs can also be applied.

In this embodiment, a supply-side protrusion 774 is formed in the central portion of the supply side of the flapper, and a supply-side nozzle orifice 777 is formed to fit into the supply-side protrusion in a non-contact manner. Further, an exhaust-side convex portion 775 is formed in the central portion of the flapper on the exhaust side, and an exhaust-side nozzle orifice 778 is formed to fit into the exhaust-side convex portion without contact. That is, on the supply side and the exhaust side, symmetrical projections and orifices for accommodating the projections were formed. The same applies to the fourteenth and fifteenth embodiments, which will be described later. However, even if the convex portion and the orifice for accommodating the convex portion are not formed symmetrically, they can be formed only on either the intake side or the exhaust side. may If only one of the protrusion and the orifice that accommodates the protrusion is provided, the protrusion penetrates deeply into the orifice to block the outflow, and the degree of penetration (valve current value) can be used to control the flow rate. can. For example, one may be provided with a protrusion and an orifice for accommodating this protrusion, and the other may be a combination of a general nozzle flapper valve (eg, see FIG. 1). (not shown)

In this embodiment, the protrusion is formed on the flapper side and the orifice for accommodating this protrusion is provided on the nozzle side. As in the embodiments described later, a cylindrical portion having an opening communicating with a supply source is formed on the nozzle side, and a concave portion is formed on the flapper side to accommodate this cylindrical portion while maintaining a narrow gap. good too. Alternatively, the tip of the tapered portion (convex portion) of the nozzle itself may be used as an insert and accommodated without contact with the inner peripheral surface of the reverse tapered portion (concave portion) formed on the flapper side. What is necessary is to provide an annular flow path forming structure in which the axial length of the flow path changes according to the relative movement of the nozzle and the flapper. (not shown)

<Fourteenth Embodiment>
FIG. 48 is a front cross-sectional view of a pneumatic servo valve according to Embodiment 14 of the present invention. It is devised.

300 is a central shaft, 301 is a bottom portion of this central shaft, 302 is an outer frame portion of the central shaft, 303 is a coil bobbin, and 304 is a coil. 305 is a cylindrical housing, 306 is the bottom of the housing, 307 is a fastening bolt, 308 is an exhaust side flow passage, 309 is a discharge port, 310 is a supply side housing, 311a is an intake port, 311b is a supply side flow passage, and 312 is a supply side flow passage. It is a control-side channel leading to a pneumatic actuator (not shown). A flapper 313 is sandwiched between the supply-side housing 310 and the housing 305 and is fixed by a bolt (not shown) that fastens the two members 305 and 310 .

Reference numeral 314 denotes a supply-side gap formed between the supply-side housing and the flapper, and 315 an exhaust-side gap formed between the flapper and the housing.

316a, 316b, 316c, and 316d are circulation holes formed in the flapper (316b and 316d are not shown), 317 is a supply side nozzle (forward direction nozzle) opening, and 318 is an exhaust side nozzle (reverse direction nozzle) opening. Department. 319 is a constant-pressure source port connected to the atmosphere, 320 is the magnetic pole of the electromagnet, and 321 is in close contact with the flapper at the flapper-side end surface of the outer frame.

49a shows the valve input current I=0 (initial value), FIG. 49b shows the input current I≈Imax/2 (operating point), and FIG. 49c shows the input current I=Imax (maximum value). In the figure, 322 is a supply side nozzle orifice, and 323 is an exhaust side nozzle orifice. 324a, 324b (not shown), and 324c are circulation grooves formed on the flapper-side end face of the magnetic pole 320, and have the same function as in the fourth embodiment (FIG. 25). 325 is a flapper supply side projection formed on the supply side of the flapper, 326 is a flapper exhaust side projection formed on the exhaust side of the flapper, and 327 is a supply housing side formed on the flapper side of the supply side housing. A projection 328 is a flapper-side depression formed on the supply side of the flapper, 329 is a control chamber, and 330 is a constant pressure chamber. Since this constant-pressure chamber is connected to a constant-pressure source port 319 connected to the atmosphere, the pressure is always maintained at a constant P=P0 (atmospheric pressure). The two members of the supply housing side projection 327 and the flapper side recess 328 are slidably fitted in the axial direction while always maintaining a narrow gap to form a non-contact seal portion 331 .

At the valve input current I=0 in FIG. 49a, the flapper supply-side protrusion 325 penetrates deeply into the supply-side nozzle orifice 322, and the flow in the narrow annular gap formed by both members is viscous. Therefore, the amount of air flowing into the control chamber 329 of this valve from an air pressure supply (not shown) is very small.

At the input current I≈Imax/2 (operating point) in FIG. The flow of fluid entering the control chamber from the supply nozzle is in a transition region transitioning from a viscous flow region to a potential flow region. Fluid flow from the control chamber into the exhaust nozzle is also in the transition region. Therefore, the flow rate of air flowing out from the air pressure source (not shown) to the exhaust side is still sufficiently small.

At the input current I=Imax in FIG. The flow of fluid entering the control chamber from the supply nozzle is in the potential flow regime. Further, the flapper exhaust-side convex portion penetrates deeply into the exhaust-side nozzle orifice 326, and the flow in the narrow annular gap formed by both members is a viscous flow. Therefore, the amount of air flowing out from the control chamber to the atmosphere is very small.

Now, when the input current is in the range of 0<I<Imax, the pressure Pa in the control chamber 329 changes in the range of P0<Pa<PS depending on the position of the flapper. However, since the constant pressure chamber 330 is shielded from the control chamber 329 by the seal portion 331, the pressure is P=P0 (constant).

The generated force (attractive force) F of the electromagnet, the load due to the pressure difference between both sides of the flapper, and the restoring force due to the spring rigidity of the flapper are balanced. Assuming that the area where the pressure in the constant pressure chamber 330 effectively acts on the flapper is S1 and the area on the supply side of the flapper covered with the non-contact seal portion 331 is S2, the total amount of control pressure Pa applied on the exhaust side of the flapper is The area is S1+S2. If the spring stiffness of the flapper (disk) is K and the displacement of the flapper is X, then

(Pa- P0)S1 ＞＞Kxとなるように、定圧室３３０のフラッパ面積S1を設定すれば、電磁石の発生力と制御圧力のゲージ圧Pa- P0は概略比例する。

If the flapper area S1 of the constant pressure chamber 330 is set so that (Pa-P0)S1>>Kx, the force generated by the electromagnet and the gauge pressure Pa-P0 of the control pressure are approximately proportional.

By using the magnetic saturation phenomenon, for example, as shown in the graph (Type C) of FIG. 21, the force generated by the electromagnet can have a proportional relationship with the current value. Therefore, in this embodiment, it is possible to obtain a valve characteristic in which the control pressure is proportional to the input current while maintaining the feature of low air consumption.

In this embodiment, the magnetic poles of the electromagnet are provided only on the central axis, but a second magnetic pole may be provided as in the above-described embodiments in order to increase the attractive force. (not shown)

<Fifteenth Embodiment>
FIG. 50 is a front cross-sectional view of a pneumatic servovalve according to Embodiment 15 of the present invention, which maintains the characteristic of low air consumption and maintains a proportional relationship between the control pressure and the input current, as in the previous embodiment. The valve structure has been devised so that In the above embodiment, the constant pressure chamber is provided on the outer peripheral portion of the flapper, but in this embodiment, it is provided on the exhaust side at the center of the flapper.

900 is a central shaft, 901 is the bottom of this central shaft, 902 is an outer frame portion formed by a circle concentric with the axis of the central shaft, 903 is a coil bobbin, and 904 is a coil. 905 is a cylindrical supply side housing, 906 is the bottom of the supply side housing, 907 is a fastening bolt, 908 is a supply side flow passage, 909 is an intake port, 910 is an exhaust side housing, 911a is a discharge port, and 911b is an exhaust side flow. Channel 912 is a control side channel leading to a pneumatic actuator (not shown). A flapper 913 is sandwiched between the exhaust-side housing 910 and the supply-side housing 905 and fixed by a bolt (not shown) that fastens the two members 905 and 910 . Reference numeral 914 denotes a supply-side gap formed between the supply-side housing and the flapper, and 915 an exhaust-side gap formed between the flapper and the exhaust-side housing. 916a, 916b, 916c, and 916d are communication holes formed in the flapper (916b and 916d are not shown), 917 is the magnetic pole of the electromagnet, and 918 is in close contact with the flapper at the flapper-side end surface of the outer frame. .

Hereinafter, in the enlarged view 51, 919 is a supply side nozzle orifice, and 920 is a flapper supply side projection formed on the supply side of the flapper. The fluid resistance R1 of the supply-side channel is adjusted by fitting the flapper supply-side convex portion with the supply-side nozzle orifice 919 .

921 is a flapper exhaust side concave portion formed on the exhaust side of the flapper, and 922 is a housing exhaust side convex portion formed on the exhaust side housing. The fluid resistance R2 of the exhaust-side channel is adjusted by fitting the housing exhaust-side protrusion and the flapper exhaust-side recess 921 together. The supply-side gap 914 and the exhaust-side gap are connected by the communication hole, and these two gaps form a control chamber 923 in the same manner as in the above-described embodiments. The pressure Pa in the control chamber is determined by the supply source pressure PS, the fluid resistance R1 of the supply side channel, and the fluid resistance R2 of the exhaust side channel.

924 is a constant pressure chamber, which is connected to the exhaust side passage 911b connected to the atmosphere, so that the pressure is always kept constant P=P0 (atmospheric pressure). As in the above-described embodiment, the generated force (attractive force) F of the electromagnet, the load due to the pressure difference between the flapper surfaces, and the restoring force due to the spring stiffness of the flapper are balanced. S1 is the area determined by the radius r1 of the constant pressure chamber 924, and S2 is the flapper supply side area determined by the radius r2 on the opposite side of the constant pressure chamber. is S1.

式(15)において、(Pa- P0)S1 ＞＞Kxとなるように、定圧室９２４の半径r1を設定すれば、電磁石の発生力と制御圧力のゲージ圧Pa- P0は概略比例する。したがって、前述した実施例同様に、低消費空気流量の特徴を維持して、かつ入力電流に対する制御圧力が比例関係になるようなバルブ特性を得ることができる。

In equation (15), if the radius r1 of the constant pressure chamber 924 is set so that (Pa-P0)S1>>Kx, the force generated by the electromagnet and the gauge pressure Pa-P0 of the control pressure are approximately proportional. Therefore, it is possible to obtain a valve characteristic in which the control pressure is proportional to the input current while maintaining the feature of low air consumption, as in the above-described embodiment.

In the embodiment, the non-contact seal portion 331 is formed to keep the pressure in the constant pressure chamber 330 constant, but a seal member such as an O-ring may be used.

<Sixteenth Embodiment> ... Other Examples (Electropneumatic Converter)
FIG. 52 is a front sectional view of a pneumatic servo valve according to Embodiment 16 of the present invention, showing a case where the present invention is used as an electropneumatic converter. In the above-described embodiment, the configuration of the nozzle flapper valve is a combination of a bidirectional flapper and two nozzles (forward direction and reverse direction). However, the present invention can also be applied as an electro-pneumatic transducer, for example, by combining one side of the flapper and one nozzle.

650 is a central shaft, 651 is the bottom of this central shaft, 652 is an outer frame of the central shaft, 653 is a coil bobbin, and 654 is a coil wound around the coil bobbin. The central shaft 650, the central shaft bottom portion 651, the outer frame portion 652 of the central shaft, the coil bobbin 653, and the coil 654 constitute an electromagnetic actuator that attracts a flapper (described later) and controls its displacement. 655 is a cylindrical housing, 656 is the bottom of this housing, 657 is a fastening bolt, 658 is a supply side housing, 659a is an intake port, 659b is a supply side flow path, 660 is an exhaust side flow path connected to the atmosphere side, 661 is an electric current. It is a control-side flow path connected to a pressure chamber (not shown) of the air converter. A disk-shaped flapper 662 is composed of a thick projection (magnetic path portion) 663 and a thin peripheral portion (elastic deformation portion) 664 . 665 is a supply side gap formed between the supply side housing 658 and the flapper 662, and 666 is an actuator side gap formed between the flapper 662 and the electromagnetic actuator side. 667a, 667b, 667c, 667d are circulation holes formed in the flapper (667b, 667d are not shown), 668 is the first nozzle provided upstream of the supply side flow path, and 669 is the facing surface of the flapper. A second nozzle 670 is a control chamber formed between the first and second nozzles. Reference numeral 671 denotes a first magnetic pole on the flapper side end face of the central shaft 650; 672, a magnetic pole ring provided on the flapper side end face of the outer frame; 673, a second magnetic pole ring formed on the flapper 662 side end face of the magnetic pole ring magnetic poles. A non-magnetic spacer 674 is interposed between the magnetic pole ring and the flapper.

When an electric current is applied to this valve, the gap between the second nozzle 669 and the flapper changes, and the fluid resistance of the second nozzle 669 changes. Assuming that the second nozzle 669 is a variable resistor RX, the variable resistor RX decreases as the input current increases. Assuming that the fixed resistance of the first nozzle 668 is R0, the pressure in the control chamber 670 Pa={RX/(RX+R0)}×PS when expressed in an electric circuit model. Therefore, the configuration of this valve makes it possible to adjust the pressure Pa in the pressure chamber (not shown) of the electro-pneumatic converter in inverse proportion to the input current.

<Seventeenth Embodiment>
FIG. 53 is a front cross-sectional view of a fluid servo valve according to Embodiment 17 of the present invention, which is an example in which the present invention is applied as a pilot valve (primary control valve) that controls a four-way guide valve on the upstream side. 700 is a central shaft, 701 is the bottom of this central shaft, 702 is an outer frame of the central shaft, 703 is a coil bobbin, and 704 is a coil. 705 is a cylindrical rear side housing, 706 is the bottom of this housing, 707 is a bolt, 708 and 709 are first supply side flow paths, 710 is a front side housing, 711 is a second supply side flow path, and 712 is an exhaust side flow path. is the road. A disk-shaped flapper 713 is composed of a thick projection (magnetic path portion) 714 and a thick elastic deformation portion 715 . 716 is a front-side gap, and 717 is a rear-side gap. 718a, 718b, 718c, and 718d are circulation holes (718b and 718d are not shown) formed in the flapper 713, 719 is a first nozzle, and 720 is a second nozzle. 721 is a first fixed nozzle provided upstream of the first nozzle, and 722 is a second fixed nozzle provided upstream of the second nozzle. 723 is a first control chamber formed between the first nozzle and the first fixed nozzle; 724 is a second control chamber formed between the second nozzle and the second fixed nozzle; A first control channel leading to the valve spool right end face (Fig. 54), 726 is a second control channel leading to the spool left end face (Fig. 54). Reference numeral 727 denotes a first magnetic pole of the electromagnet on the flapper-side end surface (center shaft end surface) of the central shaft 700, and 728 denotes a second magnetic pole formed on the outer frame portion 702 on the flapper side.

FIG. 54 is a diagram showing a four-way guide valve driven by a pilot valve according to the invention described above. 730 is a 4-way guide valve, 731 is a spool of the 4-way guide valve, 732a and 732b are springs attached to both ends of the spool, 733 is a supply port, and 734a and 734b are return ports. Control ports 735a and 735b are connected to the left and right chambers of the fluid pressure cylinder to be controlled.

<Eighteenth Embodiment>
FIG. 55 is a front cross-sectional view of a microactuator according to embodiment 18 of the present invention. By combining the present invention with the poppet valve of FIG. 56 or the four-way guide valve of FIG. can be applied as
Reference numeral 850 denotes the entire microactuator, 851 a central shaft, 852 a coil bobbin, and 853 a coil wound around the coil bobbin. Reference numeral 854 denotes an outer frame portion that accommodates the central shaft and the coil bobbin; 855, a cylindrical coil-side housing that accommodates the outer frame portion; 856, bolts for fastening the outer frame portion and the coil-side housing; and 857, a flapper. The side housing 858 is a disk-shaped flapper, and is composed of a thick convex portion (magnetic path portion) 859 and a thin peripheral portion (elastic deformation portion) 860 .

861 is a first magnetic pole on the flapper side end face of the central shaft 851, 862 is a magnetic pole ring provided on the flapper side end face of the outer frame portion, and 863 is a second magnetic pole ring formed on the flapper side end face of the magnetic pole ring. magnetic poles. 864 is a spacer made of a non-magnetic material interposed between the magnetic pole ring and the flapper, and 865 is a bolt for fastening the flapper-side housing 857 and the coil-side housing 855 together. 866a and 866b are bearings for slidably supporting the central shaft, and 867 is the output shaft end (the right end of the central shaft) of this unit as a microactuator. 868 is a fluid control section driven by the output shaft of this unit.

A displacement sensor 869 detects the displacement of the flapper, and is mounted in the central portion of the flapper-side housing 857 . A signal from the displacement sensor 869 is input to the controller 870 to control the current flowing through the coil 853 so that the output shaft end 867 of the microactuator is at a predetermined position.

In the microactuator 850 of this embodiment, when a current is applied to the coil, the output shaft end 867 protrudes from the actuator main body. Therefore, although the present actuator is of the magnetic attraction type, it can be used in the same manner as the widely used piezoelectric and magnetostrictive actuators. This operation is made possible because the central axis penetrates the outer frame portion (coil bobbin 852). In addition, by arranging the sensor 869 utilizing the space on the side opposite to the end of the output shaft, the sensor built-in microactuator can be compactly constructed.

If hydrostatic bearings are used for the bearings (corresponding to 866a and 866b) that support the central shaft 851, the effects of Coulomb friction (static friction) can be avoided, and the displacement of the output shaft ends can be controlled with high accuracy. As a result, the fluid control unit connected to the output shaft can control high flow rate and pressure (not shown). Alternatively, the radial gap between the center shaft 851 and the outer frame portion 854 is sufficiently large, and the right end (the output shaft end side) of the center shaft 851 is supported by an elastic member corresponding to the flapper. . That is, by elastically supporting both ends of the central shaft 851 with disk-shaped springs, the central shaft 851 can be supported in a non-contact manner without being affected by Coulomb friction. As the disc-shaped spring, the spiral disc spring described above in the seventh embodiment, or the cloud-shaped spring can be applied (not shown).

FIG. 56 shows an example of a poppet valve driven by the microactuator.
880 is a tapered portion, 881 is a nozzle portion fitted with this tapered portion, 882 is a housing, 883 is a fluid supply port, and 884 is a fluid output port.

FIG. 57 is a diagram showing a four-way guide valve driven by the microactuator. 890 is a 4-way guide valve, 891 is a spool of the 4-way guide valve, 892 is a supply port, and 893a and 893b are return ports. Control ports 894a and 894b are connected to the left and right chambers of the fluid pressure cylinder to be controlled.

As mentioned above, in the case of Maxwell's stress-based actuator, the maximum value of the magnetic gap between the magnetic pole and the flapper where the magnetic attraction can be effectively used was 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 above the maximum value the magnetic attraction force usually falls off significantly. However, as described in the second embodiment, if a suitable magnetic material and a thin disk are used for the movable portion corresponding to the flapper, the displacement characteristic of the flapper with respect to the current exhibits excellent linearity (linearity). I was able to find out what I could get in the process of this research. As described above in the thirteenth embodiment [see the Type (II) graph in FIG. , the stroke of the flapper can be greatly increased. Incidentally, the stroke of piezo actuators, which have been widely used in the past, is limited to about 50 μm at most, and the limit of giant magnetostrictive actuators is about 100 μm. Therefore, by utilizing the magnetic saturation phenomenon, it becomes possible to control displacement on the millimeter order, which was not possible with conventional piezo actuators and giant magnetostrictive actuators. Furthermore, as described above in the first embodiment (see Table 2), the thrust constant is higher than that of a voice coil motor (linear motor), and it can be driven with a small amount of electric power, making it possible to significantly reduce the size of the actuator. is.

<Nineteenth Embodiment>
58 is a front cross-sectional view of a microactuator according to Embodiment 19 of the present invention. Similar to Embodiment 18, the present invention can be used with the poppet valve of FIG. 56 or the 4-way guide valve of FIG. By combining them, they can be applied as a fluid servo valve. However, while the structure of Embodiment 18 causes the output shaft to protrude by applying a current, the output shaft of this embodiment causes a retraction.

820 denotes the entire microactuator, 821 denotes a central shaft, 822 denotes a coil bobbin, 823 denotes a coil wound around the coil bobbin, 824 denotes an outer frame housing the central shaft and the coil bobbin, and 825 denotes a cylindrical coil. A side housing, 826 is a fastening bolt, 827 is a flapper side housing, and 828 is a disk-shaped flapper, which is composed of a thick convex portion (magnetic path portion) 829 and a thin outer peripheral portion (elastic deformation portion) 830. be done.

Reference numeral 831 denotes a first magnetic pole on the flapper side end face of the central shaft 821; 832, a magnetic pole ring provided on the flapper side end face of the outer frame; 833, a second magnetic pole ring formed on the flapper side end face of the magnetic pole ring; magnetic poles. 834 is a fastening bolt, 835 is an output shaft end of the unit integrated with the flapper, and 836 is a spacer made of non-magnetic material. 837 is a fluid control section driven by the output shaft of this unit.

Here, it is assumed that the microactuator of this embodiment is directly connected to the poppet valve shown in FIG. With no current applied, if the tapered portion of the poppet valve is in close contact with the nozzle portion 880, fluid flow is blocked. By applying electrical current, the taper moves away from the taper and fluid is supplied from fluid supply port 883 to fluid output port 884 . In other words, it is possible to provide a fail-safe function that shuts off the flow path in an emergency when the power is suddenly turned off. As with the above-described embodiments, the microactuator can be applied to various applications other than fluid devices.

<Twentieth Embodiment>
This embodiment improves the dynamic stability of the flapper due to the fluid force from the nozzle. 6 (FIG. 28) is an enlarged view of the nozzle and flapper portion.

In the process of this research, which pursues the ideal servo valve structure, it was found that the shape of the supply side channel has a great effect on the dynamic stability of the flapper. A high-pressure fluid force from the nozzle is applied to the flapper, and this fluid force excites the flapper. If the supporting rigidity of the flapper (disk), which is determined by the plate thickness and supporting outer diameter, is reduced, a large flapper displacement (flow rate) can be obtained even if the attractive force of the electromagnetic actuator is weak. However, the weaker the supporting rigidity of the flapper and the more the pulsating component in the fluid flow, the more the flapper vibrates at its resonance frequency due to the fluid force, generating noise and destabilizing the flow rate. I have a problem.

FIG. 59 shows a structure in which nozzles are attached in Embodiment 3. FIG. A supply side nozzle (forward direction nozzle) 129 and an exhaust side nozzle (reverse direction nozzle) 130 are press-fitted to the supply side channel 122 and the exhaust side channel 119, respectively. The nozzle mounting method described above is commonly employed in conventional servo valves (see FIGS. 69 and 71) in order to facilitate adjustment of the nozzle projection .delta.n (see FIG. 4a).

FIG. 60 shows the structure of Embodiment 6, in which the supply-side nozzle (opening 269 is shown) is integrally formed with the supply-side housing 260 . Similarly, the exhaust-side nozzle (opening 270 is shown) is also integrally formed with the central shaft 250 . A channel having a length L is formed with the diameter of the channel closest to the nozzle opening 269 being Φd. This flow path of length L will be called a rectifying section 279 . The length L of the straightening section 279 in FIG. 60 is sufficiently long compared to the case of FIG. 59 with independent nozzles 129 attached. As a result of research, it has been found that the flapper 263 is less likely to be affected by the excitation action due to the fluid force by making the flow path length L longer and preventing the flow velocity from abruptly changing. If there is an obstacle in the fluid flow, a Karman vortex will be generated, and the elastic body downstream of the obstacle will vibrate at its resonance frequency like a flag (corresponding to a disc) fluttering in the wind in the Karman vortex street. It is known to vibrate and can be explained as a similar phenomenon. The location where the channel diameter ΦD suddenly changes from the channel diameter ΦD to the channel diameter Φd (the location indicated by the dashed line A in the figure) causes a sharp change in the flow velocity and becomes an obstacle to the smooth flow of the fluid. The larger the ratio L / d, the more stable the flapper 263 against the fluid force, and compared to the case of FIG. effect is obtained. If configured to satisfy L/d>5, the flapper 263 could maintain dynamic stability even if the flapper stiffness was made sufficiently small. The channel diameter Φd in the rectification section 279 may not be uniform, and may be tapered toward the upstream side by eliminating the channel diameter step at the portion indicated by the dashed line A. In this case, the average inner diameter in section L is set to Φd so as to satisfy the above condition of L/d.

While the flapper of the conventional servo valve is composed of a rigid body, in the embodiment of the present invention, the flapper itself is a thin disk-shaped elastic body, and the spring rigidity can be reduced. is big. (See Table 1 of Embodiment 1)

Although the nozzle is integrally formed with the housing in Embodiment 6, an independent nozzle having a long flow path L may be attached. Alternatively, after adjusting the protrusion amount δn of the short nozzle 129 of Embodiment 3 (Fig. 59), a long pipe having the same inner diameter may be inserted to provide a flow path corresponding to the rectifying section of length L. Further, in the supply side channel 261 near the suction port 278, for example, a regulator in which thin orifices with a small inner diameter are bundled in a honeycomb shape, or a channel in which a plurality of narrow gaps flowing in the centripetal direction are formed in the circumferential direction. , etc. is more effective. (not shown)

In this embodiment, the rectifying section is provided in the flow path on the supply side of the fluid, but if the same rectifying section is provided in the flow path on the exhaust side, the dynamic stability of the flapper is further improved. (not shown)

Now, when the supply-side gap 266 near the supply-side nozzle 269 is set to be a narrow gap, the flow velocity of the fluid that flows out from the supply-side nozzle opening and flows in the radial direction increases, and the left and right gaps 266 and 267 Due to the difference in the dynamic pressure difference, the flapper generates a force that attracts the supply side. As a result, for example, when the current value is small, a dead zone (the flapper sticks to the wall surface) occurs in the flapper displacement (flow rate) characteristics. Therefore, in the flapper 263 in the vicinity of the outer peripheral side of the opening of the supply nozzle 269, a plurality of small-diameter through holes are formed, for example, in the circumferential direction. Since the fluid flows from the supply-side gap 266 to the exhaust-side gap 267 in the shortest distance, the dynamic pressure difference between the left and right gaps becomes small, thereby solving the above problem. This effect can be similarly obtained in the above-described embodiments including the present embodiment. Even if the gap on the exhaust side is narrow, a similar phenomenon occurs in the vicinity of the maximum value of the current, but this problem can be solved by forming the through hole with a small diameter as described above. (not shown)

Further, it is also effective to solve the above-mentioned problems related to the dynamic pressure difference by making the gap between the two gaps 266 and 267 tapered, for example, increasing in proportion to the size of the radius. (not shown)

Supplement (1)
(1) Changing the displacement characteristics of the flapper by selecting the BH characteristics of the magnetic material As shown in the second embodiment, for example, magnetic saturation is achieved by using an appropriate magnetic material and a thin disk for the movable part corresponding to the flapper. By using this phenomenon, the displacement characteristic of the flapper with respect to the current can have excellent linearity (controllability).
In the above-described embodiment, by making the shape of the flapper convex, for example, it was possible to change the displacement characteristics of the flapper with respect to the current. However, the displacement characteristics of the flapper with respect to current also change depending on the BH characteristics of the magnetic material (magnetic flux density characteristics with respect to magnetic field intensity).

FIG. 61 shows the characteristics of three types (A, B, and C) of magnetic materials having different "magnetic flux density characteristics with respect to magnetic field intensity". As described above, HC is the magnetic force boundary value between the linear region and the magnetic saturation region, BC is the magnetic flux density boundary value when H=HC, and Bmax is the maximum saturation magnetic flux density. The above HC, BC, and Bmax can be used to organize the properties of any magnetic material.

FIG. 62 qualitatively shows the difference in the displacement characteristics of the flapper with respect to current when the materials A, B, and C are used. That is, by selecting the BH characteristics (HC, BC, Bmax) of the magnetic material, the displacement (flow rate) characteristics of the flapper with respect to the current can be selected according to the characteristics required for the application target.

(2) Method of providing a magnetically saturated portion at a location other than the flapper In the above-described embodiments, the magnetic saturation is adjusted by using a thin plate flapper having a small magnetic path area. However, the magnetic saturation phenomenon can be used by using any of the elements constituting the closed loop magnetic circuit.

In FIG. 63, a portion with a small magnetic path area is provided on the central axis. 280 is a central axis, 281 is a flapper, 282 is a coil, 283 is a coil outer frame portion, and 284a and 284b are flapper fixing portions. The flapper 281 is composed of a thick plate 281a (magnetic path portion) and a thin outer peripheral portion (elastic support portion) 281b. Reference numeral 285 denotes a small-diameter portion formed in the intermediate portion of the central axis, which has the smallest magnetic path area in the closed-loop magnetic circuit. That is, the cross section of the small diameter portion 285 corresponds to the maximum magnetic flux control surface (see FIG. 10c) for controlling the magnetic saturation phenomenon. A maximum magnetic flux control surface may be provided, for example, on the outer peripheral surface of the coil outer frame portion 283 or the center shaft bottom portion 286 as a portion that minimizes the magnetic path area for adjusting the magnetic saturation phenomenon. Alternatively, 287 may be the first magnetic pole and 288 the second magnetic pole, and the areas of the flapper-side end surfaces of these magnetic poles may be locally made sufficiently small to form maximum magnetic flux control surfaces with small magnetic path areas. For example, the flapper side of the coil outer frame portion 283 may be locally tapered. (not shown)

In each of the above-described embodiments of the present invention, air is used as the working fluid, but oil, air, and various types of gas can be applied as the working fluid used in the present invention. For example, the electromagnetic coil portion may be molded (sealed) with resin so that it can be exposed to liquid.

Permalloy (B), electromagnetic stainless steel, pure iron, etc. can be used as the disk (flapper) material. Also, magnetic materials may be used for the parts constituting the closed-loop magnetic circuit, and non-magnetic materials may be used for other parts such as the housing.

For example, in the second and third embodiments, flow rate (displacement) characteristics with good linearity and large stroke current are obtained by utilizing the magnetic saturation phenomenon of the parts (disks) that make up the closed loop magnetic circuit. showed that it is possible. However, even if the magnetic saturation phenomenon is not used, the present invention can be applied using the valve structure disclosed in each embodiment. For example, in the graph of displacement characteristics against current value in FIG. 14, the upper limit of the current is set before the displacement rises sharply so that Imax=0.015A when the disk plate thickness h=0.5 mm. In this case, the displacement (flow rate) stroke is small, but the fluid servo valve can be realized in the range of 0<I<Imax.

In the second embodiment using the magnetic saturation phenomenon, the magnetic saturation phenomenon is adjusted by setting the location where the magnetic path area is the smallest as the “maximum magnetic flux control surface”. The magnetic saturation phenomenon can also be adjusted by using a magnetic material with a small magnetic permeability instead of the size of the magnetic path area. In this case, a member composed of a magnetic material having a low magnetic permeability may be locally arranged in the closed loop magnetic circuit. Also, as a result of actually measuring the flow rate characteristics with respect to the current value, as shown in the graph of FIG. This indicates that the magnetic saturation phenomenon is effectively utilized. However, even if the setting range of the maximum operating current of the valve is less than the value of the X axis at point E (I<IC), even if, for example, it is set only to a "downward convex curve" that does not include magnetic saturation, the above The valves of the examples are applicable. The reason for this is that, as shown in FIG. 15, for example, when the plate thickness of the disk is reduced, the effect of linearizing the flow characteristics can be obtained from the stage where the current is sufficiently small. In this case, the point c in FIG. 16 is set to Imax (=Iad), the current range Ibc is obtained from the envelope curve Bb having the largest gradient, and the "linearization effect index η" can be evaluated by equation (12). Alternatively, the effect index η may be obtained from the angles α and β.

Also, as is well known, even if some of the control elements are nonlinear, by applying feedback control, the "output to input" characteristic has the effect of linearizing the entire control system, and the nonlinearity of the control elements Gender can compensate. Therefore, even if the linearization effect index η is low, the utility of the valve of this embodiment is not denied depending on the application. A permanent magnet may also be placed in parallel with the electromagnet to adjust the level of magnetic saturation.

Supplement [2] Fluid Servo Device to which the Valve of the Present Invention is Applied In recent years, the performance required for the anti-vibration table used in semiconductor manufacturing equipment and inspection equipment has become higher and higher as products become more highly integrated. The stage mounted on the vibration isolation table (592 in FIG. 68) has been increasing in size and speed in recent years in order to improve productivity. It has been demanded. As is well known, it is possible to improve the vibration isolation and damping performance of a device by selecting and devising a control system (synthesis) such as speed, acceleration, pressure or pressure differential feedback, feedforward, etc. for the controlled object. . for example,
i. Acceleration feedback (using the acceleration sensor 586 in FIG. 68) is equivalent to an increase in mass m. Depending on the conditions, effects such as lowering the natural frequency and reducing the resonance peak can be obtained.
ii. If feedforward is performed using a signal from a ground motion acceleration sensor (588 in FIG. 68) arranged directly under the surface plate, the vibration isolation performance can be greatly improved over a wide frequency range.

The concept underlying the active vibration isolator is called the skyhook theory. The skyhook theory states that if an object can be moved in a state of suspension (skyhook) on an imaginary wire (wire) stretched in the air, as shown in Figure 64a, the object will always be in a stable state. The theory is that it can be preserved.

However, in reality, it is not possible to stretch wires in the air, so as shown in FIG. 64b, an object (controlled object) is supported by an actuator and an acceleration sensor is arranged above the object. A virtual skyhook can be realized by controlling the acceleration FB of the actuator so that the output of this acceleration sensor becomes α→0. FIG. 65 is a graph showing a model of vibration isolation performance of a vibration isolator using a pneumatic actuator. Graphs a, b, and c in the figure are the cases where only proportional displacement feedback is applied. Graphs a', b', and c' in the figure show improvements in vibration isolation performance by selecting a control system for the actuators a, b, and c. That is, in addition to proportional displacement feedback, acceleration feedback (above i.) and ground motion acceleration feedforward (above ii.) are applied. There is a trade-off between vibration isolation performance (vibration isolation performance) against ground motion disturbance and vibration suppression performance against linear motion disturbance. It is selected. Therefore, acceleration feedback and ground acceleration feedforward play an important role in improving vibration isolation performance, and the application of acceleration feedback is a prerequisite for active vibration isolators, mainly to reduce resonance peaks.

FIG. 66 shows an example of a control block diagram of an active vibration isolator, and part A indicated by a chain line is a controlled object including a surface plate (581 in FIG. 68).

FIG. 67 shows an example of open loop transfer characteristics obtained from the control block diagram of FIG. is shown. Using this graph (Bode diagram), we can see why the servo valve used in the active pneumatic servo vibration isolator, whose responsiveness (eigenvalue) of the entire system is several Hz to tens of Hz, has a high resonance frequency of several hundred Hz. explain what is required. As shown in the table in the graph,
(1) Curve i. shows the case where no acceleration feedback (FB) is applied, the resonance frequency of the pneumatic servo valve is low, and f0 = 100 Hz (point A in the figure).
(2) curve ii. shows the case where acceleration FB is applied, the resonance frequency of the pneumatic servo valve is low, and f0 = 100 Hz (point B in the figure).
(3) Curve iii. shows the case where acceleration FB is applied and the pneumatic servo valve of the present invention is applied, the resonance frequency is high, f0=1000 Hz (point C in the figure).

The above (1), (2), and (3) are evaluated from the viewpoint of control stability. Incidentally, point D (5.5 Hz) in the figure is an eigenvalue determined by the spring rigidity of the controlled object including the surface plate and the pneumatic actuator. As is well known, the system is stable if the following two points are satisfied on the Bode diagram of the open-loop transfer function.
(i) Positive gain margin at phase crossing (ii) Positive phase margin at gain crossing

In the case of (1) above, even if the resonance frequency of the pneumatic servo valve is f0=100 Hz, the above (i) and (ii) are satisfied, and the system is stable.
In the case of (2) above, the acceleration FB increases the gain and delays the phase by 180 degrees. Furthermore, the system becomes unstable because the gain margin is negative (gain>0) at the resonance point f0=100 Hz (point B) of the servo valve.
In the case of (3) above, the application of acceleration FB increases the gain and delays the phase by 180 degrees, as in (2) above. However, at the resonance point f0=1000 Hz of the servo valve of the present invention, the gain of the system is sufficiently lowered and there is a sufficiently large gain margin (gain<0), so the system is stable.

As a result of many experiments, if the resonance frequency of the pneumatic servo valve is set to 200Hz or more, the gain of the acceleration FB can be set at the minimum necessary level. However, it was desirably 300 Hz or more. Therefore, in the case of the first embodiment, for example, the plate thickness, outer diameter, support conditions, etc. of the disk-shaped disk (flapper) may be determined so as to satisfy the above conditions. The same applies to active vibration isolation devices in other embodiments. A similar problem arises when acceleration feedback is applied to a pneumatic servo device other than an active vibration isolator, so the flapper specifications should be determined so as to achieve a satisfactory resonance frequency.

Although the case where the valve of the present invention is applied to an 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 pneumatic artificial muscles that can cooperate with humans, fine position control and soft movements not found in electromagnetic motors can be achieved. Alternatively, a walking support device that supports ankle dorsiflexion using pneumatic rubber artificial muscles, a power-assisted chair driven by pneumatic cylinders, a fine movement stage driven by pneumatic bellows, and a pneumatic manipulator that grasps objects with complex shapes. , pneumatic braking systems for rail vehicles, active suspension for vehicles, and so on.

The pneumatic servo system i. clean, ii. Easy to maintain, iii. High power/weight ratio compared to electric type, iv. It has various features not found in other methods, such as smooth movement due to compressibility, and v. force control. It is no exaggeration to say that the servo valve, which is the heart of the system, determines the performance and cost of the pneumatic servo system the most. Widespread adoption is expected to accelerate significantly.

Other embodiments will be described. The flapper support member is for fixing a part of the flapper and deforming the flapper itself by the attractive force of the electromagnet, but for example, it supports the flapper so that it can swing and changes the attitude of the flapper. I don't mind. That is, in the present invention, the electromagnet may be arranged so that the magnetic lines of force pass through the flapper itself which is oscillatably provided by the flapper supporting member, and the separation distance between the nozzle and the flapper may be changed. In this case, the flapper is made of a magnetic material, and a current is applied to the electromagnet until the magnetic force applied to the flapper enters the magnetic saturation region. Something close to what is described can be implemented.

10 central shaft 11 central shaft bottom 12 outer frame 13 coil bobbin 14 coil 15 housing 16 exhaust side bottom plate 19 exhaust side flow path 20 discharge port 21 supply side bottom plate 22 supply side flow path 23 control side flow path 24 flapper 26 supply side gap Portion 27 Exhaust-side gaps 28a, 28b, 28c, 28d Flow hole 29 Forward nozzle 30 Reverse nozzle 31 First magnetic pole 32 Second magnetic pole 33 Control chamber 34 Suction port

Claims (30)

a nozzle configured such that the flow path communicates with the fluid supply source and the inner diameter of the tip portion is tapered;
a flapper provided to face the tip of the nozzle;
a flapper support member that fixedly supports a portion of the flapper;
an electromagnet provided to generate an attractive force with respect to the face plate portion of the flapper;
a closed loop magnetic circuit configured to include at least the electromagnet and the flapper;
The flapper is elastically deformed by the attraction force of the electromagnet to change the separation distance between the tip of the nozzle and the flapper, and
The magnetic properties of the magnetic material parts that make up the closed loop magnetic circuit are
A linear region that is a region below the magnetic force boundary value Hc and has a portion where the characteristics of the magnetic flux density B with respect to the magnetic force H are in a roughly proportional relationship,
It has a magnetic saturation region, which is a region above the magnetic force boundary value Hc and where the magnetic flux density B converges to a predetermined value as the magnetic force H increases,
The magnetic force boundary value Hc is a value obtained from the intersection of the BH characteristic envelope in the linear region and the BH characteristic envelope in the magnetic saturation region,
A fluid servo valve according to claim 1, wherein the magnetic flux density of the magnetic flux flowing through the magnetic material component enters the magnetic saturation region when the current flowing through the electromagnet is increased within the movable range of the flapper . Bc is the magnetic flux density boundary value in the boundary region between the linear region and the magnetic saturation region, RS is the sum of linear magnetic resistances of the closed loop magnetic circuit at the maximum value Imax of the current flowing through the electromagnet, and RS is the number of coil turns of the electromagnet. 2. The method according to claim 1, wherein Bmax>Bc, where N, magnetic flux is Φmax=N×Imax/RS, magnetic path area of the magnetic material part is Sc, and magnetic flux density is Bmax=Φmax/Sc. fluid servo valve. 3. A fluid servo valve according to claim 2, wherein the flow rate characteristic with respect to the current forms an upwardly convex curve in the vicinity of the maximum value of the current applied to the electromagnet. 3. The flapper is composed of a substantially flat plate-shaped member, and the flapper is provided with a restoring force proportional to the size of the gap between the nozzle and the flapper by utilizing the elasticity of the flapper itself. 2. A fluid servovalve according to claim 1. 3. A fluid servo valve according to claim 2, wherein said flapper is said magnetic material component. The electromagnet is
a first magnetic pole formed on an inner end face facing the flapper;
a second magnetic pole formed on an outer end face facing the flapper,
The flapper
a magnetic path part formed between the first magnetic pole and the second magnetic pole in a magnetic closed loop circuit formed by the electromagnet;
an elastic support portion that is supported by the flapper support member and that elastically supports the magnetic path portion;
5. A fluid servo valve according to claim 4, wherein said magnetic path portion and said elastic support portion have different flexural rigidity. 7. The fluid servo valve according to claim 6, wherein the bending rigidity of said elastic support portion is smaller than the bending rigidity of said magnetic path portion. 7. A fluid servo valve according to claim 6, wherein said first magnetic pole faces the central portion of said magnetic path portion, and said second magnetic pole faces the vicinity of the outer edge of said magnetic path portion. The suction port is connected to a fluid supply source, a flow meter is attached to the flow path from the control room to the atmosphere, and the flow meter measures when the current applied to the electromagnet reaches the maximum value Imax (A). The flow rate is Qmax (L/min), the gradient Qmax/Imax is the reference flow rate gain α, the maximum value of the gradient is the maximum flow rate gain β in the profile of the flow rate characteristic for the input current, and the linearization effect index η = α/β is 2. The fluid servovalve of claim 1, wherein .eta.>0.2 when defined. 10. The fluid servovalve of claim 9, wherein η>0.4. The electromagnet is composed of a support shaft made of a magnetic material, a coil wound around the support shaft as an axis, and a cylindrical portion made of a magnetic material arranged to accommodate the coil,
2. A fluid servo valve according to claim 1, wherein said support shaft, said flapper, and said cylindrical portion constitute a closed loop magnetic circuit. 11. A fluid passage passing through said support shaft and communicating with a fluid supply side or an exhaust side is formed, and said nozzle is provided at said opening end of said fluid passage on said flapper side. A fluid servovalve as described. The nozzles are provided at two locations, one nozzle is provided on the fluid supply side and configured as a forward nozzle, and the other nozzle is provided on the fluid exhaust side and configured as a reverse nozzle, The forward nozzle, the reverse nozzle, and the flapper constitute a two-way nozzle flapper valve,
13. A fluid servo valve according to claim 12, wherein said forward nozzle or said reverse nozzle is provided at said flapper side open end of a flow passage formed through said support shaft. a supply-side housing provided with the forward nozzle and formed with a flow path connected to a fluid supply source;
an exhaust-side housing provided with the reverse-direction nozzle and formed with a flow path leading to the exhaust side of the fluid;
a supply-side gap that is a space formed between the flapper and the flapper-facing surface of the supply-side housing;
an exhaust-side gap that is a space formed between the flapper and the flapper-facing surface of the exhaust-side housing;
14. The fluid servovalve according to claim 13, wherein the flapper is formed with a communication hole connecting the supply side gap and the exhaust side gap. 12. The fluid servo valve according to claim 11, wherein the flapper has a plate-like shape and includes an elastically deformable portion configured to be deformable toward the nozzle. The electromagnet has a magnetic pole formed on an end face facing the flapper,
The flapper has a plate-like shape and faces the magnetic pole at its central portion, and the elastically deformable portion penetrates the flapper in the thickness direction formed between the central portion of the flapper and the flapper support member. 16. The fluid servovalve of claim 15, wherein the fluid servovalve is defined by a hole. 12. The fluid servovalve according to claim 11, wherein a radial flow passage communicating with an open end of said nozzle provided on said support shaft is formed in said flapper-side end surface of said support shaft. 16. The fluid servo according to claim 15, wherein a bypass member and a second magnetic pole are formed in the closed loop magnetic circuit so that the magnetic flux bypasses the elastically deformable portion and have an outer diameter smaller than the diameter of the coil. valve. a nozzle configured such that the flow path communicates with the fluid supply source and the inner diameter of the tip portion is tapered;
a plate-shaped flapper provided to face the tip of the nozzle;
a flapper support member that fixedly supports a portion of the flapper;
an electromagnet provided to generate an attractive force with respect to the face plate portion of the flapper;
a closed loop magnetic circuit configured to include at least the electromagnet and the flapper;
The flapper is deformed by the attraction force of the electromagnet to change the separation distance between the tip of the nozzle and the face plate of the flapper,
A space containing the flapper is partitioned by the flapper to form a control chamber in contact with one surface of the flapper and a constant pressure chamber in contact with the other surface of the flapper and maintained at a substantially constant pressure. and
The magnetic properties of the magnetic material parts that make up the closed loop magnetic circuit are
A linear region that is a region below the magnetic force boundary value Hc and has a portion where the magnetic flux density B characteristic with respect to the magnetic force H has a roughly proportional relationship,
It has a magnetic saturation region which is a region above the magnetic force boundary value Hc and where the magnetic flux density B converges to a predetermined value as the magnetic force H increases,
The magnetic force boundary value Hc is a value obtained from the intersection of the BH characteristic envelope in the linear region and the BH characteristic envelope in the magnetic saturation region,
A fluid servo valve according to claim 1, wherein the magnetic flux density of the magnetic flux flowing through the magnetic material component enters the magnetic saturation region when the current applied to the electromagnet is increased within the movable range of the flapper. A closed loop magnetic circuit is formed by connecting magnetic material members in a substantially polygonal shape in a cross-sectional view, and each of the magnetic material members is an iron core around which a coil of an electromagnet is wound, a yoke member, and a flapper. Item 1. A fluid servo valve according to Item 1. 21. A fluid servo valve according to claim 20, wherein a magnetic pole is formed at one end of said magnetic material member, and said flapper is disposed on the opposing surface with a gap therebetween. A fluid passage is formed through one of the magnetic material members to communicate with a fluid supply side or an exhaust side, and the nozzle is provided at an opening end of the fluid passage on the flapper side. 21. The fluid servovalve of claim 20. a fluid servovalve as claimed in claim 1;
a pneumatic actuator connected to the fluid servo valve;
a sensor that detects the displacement and/or vibration state of the object to be controlled;
and control means for applying a gas pressure to the pneumatic actuator that controls at least one of displacement, velocity and acceleration of the controlled object by adjusting the fluid servo valve based on information from the sensor. Fluid servo device. a gas spring that configures the flapper to have a primary natural frequency of 200 Hz or more and supports an object to be isolated from a foundation;
the fluid servo valve for supplying gas from the supply side to the gas spring and exhausting it to the exhaust side;
an acceleration sensor that detects the vibration state of the object to be isolated;
24. The fluid according to claim 23, further comprising active control means for applying a gas pressure to said gas spring that reduces vibration of said vibration isolation object by adjusting said fluid servo valve based on information from said acceleration sensor. Servo device. an electromagnet;
flapper and
a flapper support member that supports the flapper;
an output shaft that forms a closed loop magnetic circuit with an electromagnet, a disk, and a yoke material, is moved by Maxwell attraction stress generated between the electromagnet and the flapper, and is fixed to the flapper;
a microactuator portion comprising the electromagnet, the disk, the support member, the yoke material, and the output shaft;
a housing, a fluid suction port and a discharge port formed in the housing, a flow control valve for adjusting the degree of opening of a flow path connecting the suction port and the discharge port, and the output shaft of the microactuator section. By driving the flow control valve,
The magnetic properties of the magnetic material parts that make up the closed loop magnetic circuit are
A linear region that is a region below the magnetic force boundary value Hc and has a portion where the characteristics of the magnetic flux density B with respect to the magnetic force H are in a roughly proportional relationship,
Defined as a magnetic saturation region, which is a region above the magnetic force boundary value Hc and where the magnetic flux density B converges to a predetermined value as the magnetic force H increases,
The magnetic force boundary value Hc is a value obtained from the intersection of the BH characteristic envelope in the linear region and the BH characteristic envelope in the magnetic saturation region,
A fluid servo valve according to claim 1, wherein the magnetic flux density of the magnetic flux flowing through the magnetic material component enters the saturation region when the current value applied to the electromagnet is increased within the operable range of the flapper. 26. A fluid servo valve according to claim 25, wherein said output shaft is provided so as to pass through the central portion of said electromagnet. 27. A fluid servo valve according to claim 26, wherein a sensor for detecting positional information such as displacement, speed and acceleration of said output shaft is arranged at the end of said output shaft opposite said electromagnet. 28. A fluid servovalve according to claim 27, wherein said output shaft is radially supported by hydrostatic bearings. In the magnetic properties of the magnetic material parts that make up the closed loop magnetic circuit,
The range of 0<H<Hc, where the magnetic flux density B increases in proportion to the magnetizing force H, is defined as the linear region, and the range of H>Hc, where the gradient of the magnetic flux density B decreases with respect to the magnetizing force H, is defined as the magnetic saturation region. and the Hc is a magnetic force boundary value obtained from the intersection of the envelope of the BH characteristics in the linear region of 0<H<Hc and the envelope of the BH characteristics in the magnetic saturation region of H>Hc,
2. The fluid servo valve according to claim 1, wherein the magnetic flux density of the magnetic flux flowing through the magnetic material component enters the magnetic saturation region when the current flowing through the electromagnet is increased within the movable range of the flapper. . The electromagnet is
a cylindrical central axis around which a coil is wound, the front end face of which faces the flapper and forms a first magnetic pole;
a disk-shaped bottom extending radially from the end opposite to the flapper on the central axis;
an outer frame portion which has a cylindrical shape and is provided so as to form a circle concentric with the central axis with respect to the outer peripheral portion of the bottom portion, and the front end surface on the flapper side faces the flapper to form a second magnetic pole; Equipped with
The closed-loop magnetic circuit comprises the first magnetic pole, the flapper, the gap portion between the flapper and the second magnetic pole, the second magnetic pole which is the front end surface of the outer frame portion on the flapper side, and the outer magnetic pole. 2. The fluid servo valve according to claim 1, wherein the fluid servo valve is configured as a circuit extending through the inside of the frame portion and the inside of the bottom portion to reach the central axis.

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