Fluid control valve and fluid control device

Abstract

The present invention is based on a finding that, in a two-way nozzle flapper valve which is composed of a forward-direction nozzle and a reverse-direction nozzle, forming both of the nozzles from annular nozzles enables obtainment of an effect of solving the following factors which inhibit flow rate increase. (i) Reduction in effective displacement of a flapper due to fail-safe (ii) Phenomenon of flapper attachment to a discharge nozzle According to (i), the fail-safe displacement can be reduced as compared with a single hole nozzle, and thus the effective stroke of the flapper can be increased and the supply flow rate can be increased. In addition, according to (ii), the phenomenon of attachment to the discharge nozzle can be reduced, and thus the discharge flow rate can be increased. The spring rigidity of a movable body can be maintained at the same level as that of a small flow rate valve, and a large supply flow rate / large discharge flow rate valve can be realized merely by partial modification of a nozzle part.

Description

Fluid control valve and fluid control device

[0001] The present invention relates to a fluid control device for controlling the pressure and flow rate of a fluid, and more particularly to a fluid control valve that controls the pressure and flow rate of a fluid by causing relative movement between a movable body and a nozzle via an electromechanical conversion unit, and also to a fluid control device equipped with this fluid control valve.

[0002] 1. Trends in Fluid Power Systems Fluid power is a general term for the field of technology that uses pressurized fluids to transmit and control power and drive loads. Previously, it was called hydro-pneumatics because it only used oil or air as a medium. However, it has now expanded to include rapidly developing hydraulics and functional fluids that change their flow state in response to electromagnetic fields, transforming it into a diverse academic and industrial field. Servomechanisms that utilize fluid power, particularly hydraulic servos, have the longest history. Due to their high rigidity, fast response, and lighter weight than electric servos, hydraulic servos have been widely used in industrial machinery such as aircraft, machine tools, heavy-duty robots, and injection molding machines. In recent years, servomechanisms using fluids such as air and hydraulics have also been put to practical use in fields that take advantage of their unique characteristics. The core technology is the fluid pressure servo system, which receives electrical signals to arbitrarily control flow rate and pressure. The performance of this fluid pressure servo system is determined by the servo valve, the system's heart. Fluid power systems are characterized by a higher power-to-weight ratio than electric systems, but servo valves used in fluid power systems must strive for even greater power generation, faster response, smaller size, and lower cost (easy to manufacture and simple structure). Pneumatic servos have the following characteristics: (i) they do not generate heat or magnetism, (ii) they are clean and easy to maintain, and (iii) they have a simple and compact structure. Taking advantage of these characteristics, they have been used in applications such as pneumatic vibration isolation systems in the semiconductor manufacturing industry, which isolate the precision stage that carries semiconductor wafers from vibrations on the floor. Below, we will discuss new issues that have recently emerged regarding pneumatic servos used in the semiconductor manufacturing industry.

[0003] 2. Trends in the Semiconductor Manufacturing Field - Demands from the Product Side The use of vibration control to block and suppress minute vibrations is becoming more widespread in various fields, including semiconductor manufacturing processes, liquid crystal manufacturing processes, and precision machining. Microfabrication and inspection equipment used in these processes, such as scanning electron microscopes and semiconductor exposure equipment (steppers), require strict vibration tolerance conditions to ensure equipment performance. In the future, as products become increasingly highly integrated and miniaturized, processing processes will become faster and equipment will become larger, leading to increasingly strict vibration tolerance conditions. Traditionally, equipment that is susceptible to the effects of these vibrations has been supported by actuators, and active-type precision vibration isolation tables have been used, which control the actuators to reduce the vibrations.

[0004] 3. Types of Actuators Several types of actuators are used to control active precision vibration isolation tables. Their respective characteristics can be summarized as follows: Linear motors (voice coil motors) generate large displacements, but suffer from issues with high heat generation and low generated force. Piezoelectric actuators are compact and highly responsive, but generate small displacements (at most a few tens of microns) and have issues with long-term durability. Giant magnetostrictive actuators have excellent response, but generate small displacements (about twice that of piezoelectric actuators), and like linear motors, suffer from issues with heat generation and magnetic flux leakage. In contrast, pneumatic actuators are less responsive than the actuators mentioned above, but have the advantage of easily increasing the generated force by selecting the piston outer diameter and supply pressure. Furthermore, the actuator itself has the effect of isolating vibrations from the floor (vibration isolation performance) due to the compressibility of air. Furthermore, linear disturbances can be controlled by controlling the air spring pressure. In other words, the ability to combine both "vibration isolation" and "vibration control" is a feature of pneumatic actuators that other types of actuators do not have.

[0005] [Correction pursuant to Rule 91, 03 / 03 / 2025] 4. Conventional Active Vibration Isolation Table: Figure 107 shows a model diagram of a conventional active vibration isolation table using pneumatic actuators. This active vibration isolation table is well known, as described in Patent Documents 1 and 2. A plurality of sets of pneumatic actuators (582a, 582b) are arranged on floor 580 to support a base plate 581. A precision device (not shown) is mounted on this base plate 581. The pneumatic actuator consists of an air chamber 583 filled with high-pressure air to support a vertical load, and a piston 585 inserted into the upper part of this air chamber via a diaphragm 584. Acceleration sensors 586, 587a, and 587b are used to detect the vertical and horizontal acceleration of base plate 581 and the relative displacement of base plate 581 relative to floor 580, respectively. Acceleration sensor 588 detects the acceleration of floor 580 (the vibration state of the foundation). The output signals from these sensors are input to the controller 589. A servo valve 591 controlled by the controller 589 is connected to the air chamber 583 via a pipe 590. The servo valve 591, which is a nozzle-flapper type electro-pneumatic converter, adjusts the flow rate of compressed air supplied to and exhausted from the air chamber 583, thereby controlling the internal pressure P a Here, the servo valve 591 controls the supply pressure P S A control signal is given by the controller 589 to control the desired gas pressure P a It is a control valve that adjusts the pressure to a value and outputs it, with a portion of the pressure being vented to the atmosphere P0. As the devices supported by vibration isolation tables tend to become larger, air spring vibration isolation tables, which take advantage of the advantages of pneumatic actuators, are now widely used to control micro-vibrations in ultra-precision equipment.

[0006] [Correction based on Rule 91 03.03.2025] 5. Conventional Pneumatic Servo Valves: In active vibration isolation tables, which are pneumatic servo systems using pneumatic actuators, nozzle flapper valves have primarily been used to control the actuator pressure and flow rate. This nozzle flapper valve is disclosed, for example, in Patent Document 3, and its operating principle is widely used. 5-1. Modeling the Operating Principle of a Nozzle Flapper-Type Servo Valve: Figure 108 shows a structural diagram modeling the operating principle of a conventional servo valve. The servo valve can be broadly divided into an actuator section A-1 and a fluid control section B-1. In the actuator section A-1, 551 is a magnet (permanent magnet), 552 is a coil, 553 is a body that houses the coil, 554 is a flapper, 555a and 555b are a pair of yokes attached with their tips facing each other, and 556 is the tip of the flapper on the actuator side. Reference numeral 557 denotes a leaf spring which also serves as a sealing member, and 558 denotes the support center of the leaf spring. In the fluid control section B-1, 560 denotes a forward direction nozzle, 561 denotes a reverse direction nozzle, and 562 denotes the tip of the flapper on the fluid control section side. 563 denotes a supply port, 564 denotes an exhaust port, 565 denotes a load port (control port), and 566 denotes a control chamber. Supply pressure P S The gas is supplied to the control chamber 566 through the forward nozzle 560. At the same time, the gas in the control chamber 566 is discharged to the atmosphere through the reverse nozzle. The difference between the amount of gas flowing in from the forward nozzle and the amount of gas flowing out from the reverse nozzle determines the control pressure P aThe flow rate from the load port 565 is determined by the force of the permanent magnet and the electromagnet. However, the servo valves actually used have a three-dimensional structure in which the magnetic circuit of the permanent magnet and the magnetic circuit of the electromagnet are arranged perpendicular to each other. The basic structure of the pneumatic servo valve described above is derived from the long-established hydraulic servo valve technology and is used as the primary control valve (pilot valve) of an electrohydraulic control valve. 5-2. Specific Structure of a Conventional Nozzle-Flapper Servo Valve. Figure 109 shows an example of a conventional nozzle-flapper type pneumatic servo valve, with Figure 109a being a front cross-sectional view and Figure 109b being a side cross-sectional view. The servo valve 501 comprises a body 502, a flapper displacement means 504 attached to the body by mounting bolts 503, and a flapper 506 that is displaced by energizing a coil 505 (described below). The body 502 is provided with: a diametrical groove 508 formed on the surface facing the flapper displacement means 504; an input port P, an output port A, and a discharge port R for compressed air opened on the opposite surface; a flapper insertion hole 509 connecting the output port A to the groove 508; flow paths 510a, 510b having one end opening into the flapper insertion hole 509 and connecting the output port A to the input port P and the discharge port R, respectively; and a pair of nozzles 511a, 511b assembled to the flow paths 510a, 510b with their tips (nozzle portions) facing each other. The flapper displacement means 504 is provided with a magnet assembly 513 and the coil 505 assembled within the magnet assembly 513. The magnet assembly 513 includes a pair of first yokes 514a, 514b arranged diametrically opposed to each other, a pair of magnets 515a, 515b (Fig. 102b) that are cylindrical in plan view and have both circumferential ends fixed to the first yokes, and a pair of second yokes 517 attached with their tips facing each other by mounting bolts 516 to the surfaces of the first yokes 514a, 514b opposite the body 502, and a bobbin 505a around which a coil 505 is wound is attached by fitting or the like between the first yokes 514a, 514b.Protrusions 518a, 518b facing the center in a direction perpendicular to the second yoke 517 are formed integrally with the first yokes 514a, 514b on the body 502 side, and the tips of the second yokes 517, 517 and the side tips of the protrusions 518a, 518b (Fig. 102b) face each other via the flapper 506. The circumferential end faces of the magnets 515a, 515b (Fig. 102b) are fixed to the first yokes 514a, 514b by brazing, adhesive, or the like, and the flapper displacement means 504 is assembled to the body 502 by the mounting bolts 503 that pass through through holes 519 that open into recessed portions of the first yokes 514a, 514b. The flat surface of the tip 506a of the flapper 506, which passes through the central hole of the bobbin 505a and the flapper insertion hole 509, faces the tips of the nozzles 511a, 511b, and the end 506b facing the second yoke 517 has a width across flats. The axially intermediate portion of the flapper 506 is fitted into the central hole of a leaf spring 520, each of whose ends is attached to a recessed groove 508 by two mounting screws 521 (FIG. 102b). This allows the axial ends 506a, 506b to be displaced in opposite directions around the mounting portion of the leaf spring 520. The openings of the flow paths 510a, 510b are closed by plugs 522, and the flapper displacement means 504 is sealed from the body 502 by an O-ring 523 fitted onto the outer peripheral surface of the flapper 506. Furthermore, cover 524, which covers flapper displacement means 504 and is fitted to body 502, has its tip attached to body 502 by set screw 525 (Fig. 102b). Reference numeral 526 denotes a lead wire insertion member through which the lead wire (not shown) of coil 505 passes. In the servo valve 501, when current is applied to coil 505, magnet assembly 513 causes flapper 506 to swing about the support portion formed by leaf spring 520, thereby moving tip 506a toward and away from the tips of nozzles 511a and 511b, adjusting the gap therebetween, and the air pressure at output port A increases or decreases according to the amount of current applied to coil 505. That is, in the servo valve 501, tip 506b of the flapper valve, which serves as the magnetic pole of an electromagnet, is positioned in the magnetic field generated by a pair of magnets (permanent magnets) 515a and 515b.The flapper displacement in response to the input current is controlled by varying the current passing through the coil 505 of the electromagnet. 5-3. Conventional Spool-Type Servo Valve: Patent Document 4 discloses a pneumatic pressure control valve that is an improved version of the spool type used as a secondary control valve for the electrohydraulic control valve. This pneumatic pressure control valve, shown in Figure 110, is broadly composed of a spool 601, a voice coil motor 602, a speed detection sensor 603, and a control unit 604. The spool 601 is supported axially movably by a sleeve 605, and the spool is formed with a left land 606, a first plunge 607, a second plunge 608, a central land 609, and a right land 610. The left land 606 and the right land 610 are supported by hydrostatic bearings in a non-contact manner relative to the sleeve 605. The sleeve 605 is formed with a supply port 611, a load port 612, and an exhaust port 613. Reference numeral 614 denotes a control pressure chamber. Voice coil motor 602 is composed of a coil 615, a magnet 616, and a moving body 617. Similarly, speed detection sensor 603 is composed of a coil 618, a magnet 619, and a moving body 620. Spool 601 is driven to move axially by voice coil motor 602. Reference numerals 621 and 622 denote balancing springs that support spool 601 at both ends. The pressure in control pressure chamber 614 changes depending on the relative positions of load port 612 and central land portion 609 of sleeve 605, but by giving spool 601 and sleeve 605 a stepped structure, it is possible to make the drive current and output pressure of the voice coil motor proportional to each other.

[0007] Japanese Patent Application Publication No. 2006-283966 Japanese Patent Application Publication No. 2007-155038 Japanese Patent Application Publication No. 11-294627 Japanese Patent No. 4636830 Japanese Patent Application Publication No. 2015-24794 Japanese Patent Application Publication No. 2006-57719

[0008] Fluid Power and Servo Technology, Journal of the Japan Society of Fluid Power Systems, Vol. 49, No. 6, (2018) Problem to be solved by the invention

[0009] 1. New challenges for active vibration isolation systems In recent years, new challenges have emerged due to trends in the semiconductor-related business. As the throughput of semiconductor-related equipment continues to increase, there is a demand for significant improvements in the vibration isolation performance of vibration isolation tables. Exposure equipment and inspection equipment are equipped with movable stages for transporting workpieces, and active vibration isolation systems equipped with movable stages now require vibration isolation control that can quickly and accurately attenuate the vibration of the surface plate caused by the acceleration and deceleration of the movable stage. Assume that the stage acceleration and deceleration time is on the order of a few milliseconds, and acceleration α is set to 1 to 3 G (1 G = 9.8 m / s 2 ). Because a reaction force of inertia F=mα, determined by the stage mass m and acceleration α, acts on the vibration isolation table, if the stage mass m=50-150 kg, vibration control is required to attenuate the impact load of, for example, F=1000-2000 N order. The key to achieving this high acceleration and vibration control is the realization of a new drive system that combines both large generating force and high responsiveness. Furthermore, this new drive system is required to have vibration isolation performance in the low frequency range, just like conventional active vibration isolation systems.

[0010] 2. Issues with Conventional Pneumatic Servo Systems (1) Nozzle Flapper Valve Conventionally, many vibration isolation systems have been applied to improve the installation environment of precision equipment by isolating vibrations from the installation floor. The maximum control flow rate of nozzle flapper valves, which have been used in active vibration isolation tables to control the pressure and flow rate of pneumatic actuators, was approximately Q = 6–35 L / min. Using this basic nozzle flapper valve structure, we envision high-acceleration and vibration-damping control, for example, a pneumatic servo with stage acceleration and deceleration times on the order of a few milliseconds and acceleration α set to 2G or greater. In this case, the maximum control flow rate required for the nozzle flapper valve is 6–11 times that of a vibration isolation servo valve, or on the order of 200–400 L / min. However, increasing the flow rate of a nozzle flapper valve poses significant challenges. Increasing the flow rate of a nozzle flapper valve requires increasing the diameter of the intake and exhaust nozzles and the stroke of the flapper (valve element), which is the moving part of the valve. This requires increasing the spring stiffness K that balances with the flapper and the electromagnet's generated force F. The force F generated by an electromagnet is proportional to the number of coil turns n and the current I. Increasing the number of coil turns n increases the coil inductance L and resistance R. Electrical time constant T E = L / R, an increase in inductance L reduces responsiveness. Conversely, an increase in resistance R reduces the time constant, but since the force generated by an electromagnet is proportional to the coil current I, the current must be increased by the amount of increase in resistance R. When the magnetizing force (H = n x I), which is determined by the number of coil turns n and the current, increases, the maximum value of the magnetic flux Φ is suppressed due to magnetic saturation of the magnetic material that makes up the magnetic circuit, and so it becomes necessary to change the shape of the magnetic material (for example, the thickness of the flapper). In other words, as the mass m of the moving part increases, the mechanical time constant T MAn increase in the number of pneumatic actuators is unavoidable. In summary, increasing the flow rate of a nozzle flapper valve requires significant modifications to the electrical, magnetic, and mechanical components that make up the entire valve. As an example of an active vibration isolation table, consider a four-point support active control system. In this case, pneumatic actuators are placed at the four corners of the unit, with two points in the horizontal X direction and two points diagonally in the Y direction. Each actuator also incorporates an actuator that supports the load in the Z direction. Therefore, a total of eight pneumatic actuators are required, and eight sets of pneumatic servo valves are required to control each actuator. Therefore, the increased cost of the individual valves significantly increases the cost ratio of the vibration isolation table due to the large number of required valves. (2) Spool Valve: Spool valves, which have a different basic driving principle from nozzle flapper valves, have the advantage of low steady-state air consumption and a high maximum controllable flow rate. However, compared to nozzle flapper valves, they have the following performance and cost issues: As shown in the example of the spool-type gas pressure control valve disclosed in Patent Document 4, a major problem when applying a spool-type valve to an active vibration isolation table is that the primary resonance frequency of the moving part cannot be increased. In Figure 103, the resonance frequency of the moving part including the spool 601 is determined by the mass m of the moving part and the spring constant K of the balancing springs 621 and 622. The primary resonance frequency is Since the spring constant K is proportional to the mass m, the smaller the mass m and the larger the spring constant K, the higher the spring constant K. However, there is a limit to how light the spool shaft, which forms the large-flow opening, can be. Furthermore, in the case of the voice coil motor 602 using the Lorentz force, the electromechanical conversion efficiency of the generated force relative to the input current is low, making it difficult to generate a large force. Therefore, the stiffness K of the springs 622, 622 must be small, making it difficult to increase the resonant frequency of the moving parts. To apply a pneumatic servo valve to an active vibration isolation table, the primary resonance point of the servo valve must be sufficiently high, at least several hundred Hz. Therefore, the spool-type control valve employs velocity feedback control using the signal from the velocity detection sensor 603. In other words, velocity feedback can reduce the natural frequency of the system from a control perspective, thereby compensating for the low-frequency transfer characteristics. In summary, the above-mentioned spool type gas pressure control valve, which incorporates a speed detection sensor 603 and a control unit 604 in addition to a voice coil motor 602 that drives a spool 601, has the problem of being complicated in structure and large in size. Furthermore, in the spool type gas pressure control valve, the spool 601 is housed in a sleeve 605 so as to be movable in the axial direction, and is driven by a voice coil motor 602. The spool is supported in a non-contact manner relative to the sleeve by a hydrostatic bearing. The reason for using a hydrostatic bearing is that a narrow radial gap (δ r This is to maintain a minimum radial clearance (5-6 μm) without metal-to-metal contact. Maintaining a narrow radial clearance reduces air leakage from high pressure to low pressure due to the viscous resistance of the air, thereby enabling the valve to function as a servo valve. Therefore, as shown in an example in Patent Document 4, in addition to the aforementioned complex structure, spool-type gas pressure control valves require high-precision component machining and assembly, resulting in higher costs compared to nozzle flapper valves. Furthermore, as with nozzle flapper valves, a total of eight pneumatic servo valves are required to control each actuator, making their application to active vibration isolation tables costly.

[0011] As mentioned above, trends in the semiconductor industry have raised new challenges for vibration control technology in active vibration isolation tables. To achieve high-acceleration and vibration-damping control, the actuator must be able to generate a force capable of reproducing the target waveform for feedforward control with an accuracy of several milliseconds or even 0.1 milliseconds. Compared to linear motors, which have limitations in the force they can generate, pneumatic systems clearly offer advantages. However, there are currently no prior examples or research examples of applying pneumatic systems to high-acceleration and vibration-damping control. To address the new challenges facing pneumatic servo systems—namely, to simultaneously achieve the aforementioned high-precision and high-acceleration vibration-damping control—the inventors have proposed and are currently filing a patent application for a pneumatic servo system in which the vibration-damping actuator and the pneumatic vibration-damping actuator are arranged separately and independently. In this case, the vibration-damping valve that draws and exhausts the gas inside the vibration-damping actuator is also used separately and independently from the vibration-damping valve. When the vibration isolation and vibration suppression actuators are configured separately, the large flow rate that flows through the vibration suppression servo valve during vibration suppression control is only momentary (several millimeters to several tens of milliseconds). Therefore, the magnitude of the average air consumption is not an issue. However, as mentioned above, in order to increase the flow rate of a nozzle flapper valve, it is necessary to make major modifications to the electrical circuit, magnetic circuit, and mechanism that make up the entire valve. If the small electrical time constant T E and the mechanical time constant T MIf a high-flow valve (e.g., 200 L / min or more) could be realized while maintaining the above characteristics, i.e., with high responsiveness and a simple structure, it would be expected to bring about an epoch-making (innovative) achievement in the history of servo valves, which have a wide range of applications, not just in the fields of vibration isolation and vibration suppression control. However, in the current field of fluid power systems, no examples of this achievement have been seen, even at the basic research level. Therefore, the inventors traced back to the basic operating principles of servo valves, which consist of a valve disc (flapper) and a valve seat (nozzle), and attempted to identify the factors that have prevented high flow rates. The present invention was discovered through a theoretical analysis of these factors and a search for solutions. Specifically, the invention of claim 1 is a fluid control valve comprising: a movable body; a first nozzle portion and a second nozzle portion; a movable body support portion for fixing the movable body; and a drive means for the movable body, wherein the drive means is configured so that the movable body approaches a tip end of one of the first nozzle portion and the second nozzle portion and moves away from a tip end of the other; and the fluid control valve controls the difference in flow rate passing through the first nozzle portion and the second nozzle portion, or the flow rate passing through the first nozzle portion and the flow rate passing through the second nozzle portion, individually, by movement of the movable body, wherein the first nozzle portion and the second nozzle portion are comprised of an inner member and an outer member that houses the inner member, and comprises: an inner valve seat formed on the movable body side of the inner member; an outer valve seat formed on the movable body side of the outer member; an annular flow path formed by peripheral walls of the outer valve seat and the inner valve seat; and a flow path connecting the annular flow path to a supply hole for working fluid or connecting the annular flow path to a discharge hole for working fluid, An annular nozzle is formed that adjusts the opening state between the annular flow path and the movable body by moving the movable body, the outer valve seat tip, and the inner valve seat tip closer to or farther away from each other.

[0012] That is, in the present invention, both the first nozzle section and the second nozzle section are configured as the annular nozzle, and the movable body is configured to move toward the tip of one of the first nozzle section and the second nozzle section and move away from the tip of the other. The annular nozzle is compared with a conventional single-hole nozzle under the same conditions of the gap between the movable body and the nozzle and the same supply pressure. It is noteworthy that the annular nozzle has a smaller fail-safe force dependent on the pressure difference and a smaller suction force applied to the movable body than the conventional single-hole nozzle, allowing for a larger flow rate. By utilizing the flow rate difference between the working fluid flowing through the first nozzle section and the second nozzle section, for example, a two-way, three-way valve with a large flow rate can be realized, in which the pressure and flow rate of the control chamber are controlled in proportion to the displacement of the movable body. By individually controlling the flow rates of the working fluid flowing through the first nozzle section and the second nozzle section, the pressure and flow rate of the control chambers connected to the nozzles change in opposite phases, thereby realizing, for example, a two-way, four-way valve capable of controlling a large flow rate over a wide flow rate range. In other words, a pneumatic servo that combines large load generation and high-speed response can be realized when an actuator with a large piston diameter is used as a control target. Specifically, the invention of claim 2 comprises a control chamber that is a space for accommodating the movable body, and a control port that communicates with this control chamber, wherein the working fluid flows from the supply hole through a first nozzle portion that is a supply-side nozzle and into the control chamber, and flows out from the control chamber through a second nozzle portion that is a discharge-side nozzle and into the discharge hole, and wherein the fluid pressure in the control chamber or the flow rate flowing in and out of the control port is controlled by the difference in flow rate passing through the first nozzle portion and the second nozzle portion.

[0013] That is, in the present invention, a bidirectional nozzle flapper valve is configured by configuring the first nozzle section as a forward nozzle and the second nozzle section as a reverse nozzle. Furthermore, it has been discovered that configuring both nozzles in the bidirectional nozzle flapper valve as the annular nozzle eliminates the following factors that inhibit increased flow rate: (i) Reduction in the effective displacement of the movable body due to the fail-safe mechanism; and (ii) Adsorption of the movable body to the exhaust nozzle. The issues with conventional control valves with single-hole nozzles can be summarized as follows: Increasing the intake nozzle diameter limits the increase in intake flow rate because the effective displacement of the movable body due to the fail-safe mechanism decreases. The fail-safe mechanism is an essential safety function that blocks the intake port of the control valve to prevent high-pressure gas from entering the actuator when the input current I applied to the control valve is zero or during an emergency such as a power outage. Furthermore, increasing the exhaust nozzle diameter to increase the exhaust flow rate causes adsorption of the movable body to the exhaust nozzle, rendering the nozzle flapper valve inoperable. The adhesive force of the movable body to the exhaust nozzle is proportional to the area of ​​the exhaust nozzle and the pressure difference between the supply pressure and atmospheric pressure. To solve the above problem, increasing the rigidity of the movable body reduces the fail-safe displacement, increases the effective displacement, and increases the intake flow rate. Furthermore, increasing the restoring force of the movable body also increases the limit value of the exhaust nozzle diameter d, which allows for an increase in the exhaust flow rate. However, increasing the rigidity of the movable body requires an increase in the electromagnet's generated force, which leads to a deterioration in responsiveness due to a decrease in the electrical and mechanical time constants, an increase in the capacity of the power supply, and other modifications to the electrical circuits, magnetic circuits, and mechanical components that make up the entire control valve, resulting in a significant increase in cost. When the annular nozzle is applied to the forward direction (intake side) nozzle, the fail-safe displacement X is significantly increased compared to the conventional single-hole nozzle under the same nozzle outer diameter (d0 = d). fel Since the effective stroke of the movable body can be increased, the intake flow rate can be significantly increased. When an annular nozzle is used for the reverse direction (exhaust side) nozzle, the attraction force F Pis much smaller than that of a conventional single-hole nozzle. Over a wide range of outer nozzle inner diameters d0, the phenomenon of movable body adhesion to the exhaust nozzle does not occur, allowing for ample increases in exhaust flow rate. When annular nozzles are used for both the forward and reverse nozzles, the flapper rigidity can be maintained at the same level as that of a small-flow valve, and a valve with large intake and exhaust flow rates can be realized simply by improving both nozzles. The basic performance required of a control valve is to shorten the time it takes for the actuator to fill with air while also shortening the time it takes to exhaust the air. To achieve this, it is necessary to increase both the maximum intake flow rate and the maximum exhaust flow rate. Therefore, when a pneumatic control system is constructed using a control valve of the present invention in which annular nozzles are used for both the forward and reverse nozzles, the rise time constant and fall time constant can be set to similarly small values.

[0014] Specifically, in the invention of claim 3, the driving means is configured to move the movable body surface, which is the opposing surface of the annular nozzle, approximately parallel to the axial direction of the annular nozzle. That is, this invention takes note of the fact that most control valves currently used in a wide range of industrial fields have a structure in which a movable body oscillates around a central fulcrum. The crucial difference between the widely used solenoid valve (ON / OFF valve) and the control valve of this invention, which is composed of a valve disc and a valve seat, is the dynamic characteristics required of the movable part, the valve disc. The requirements for applying a pneumatic control valve to an active vibration isolation table are as follows: (i) High-speed response—a rise time of several milliseconds or less; (ii) A sufficiently high primary resonance point of the movable part (valve disc), above several hundred Hz (e.g., 200 Hz). In a conventional oscillating motion servo valve, the requirements for the movable body (valve disc) to satisfy the above conditions (i) and (ii) are summarized as follows: the mass of the movable body is on the order of several grams. In order to reduce the moment of inertia J around the center of rotation, the shape of the valve disc is such that the distance L from the center of rotation to the tip of the movable body is small, and the valve disc is configured with a width of, for example, about 2 to 3 mm. To accommodate this narrow valve disc width, it seems that the "valve seat" (nozzle) facing the valve disc would inevitably be a single-hole nozzle with a small inner diameter. The focus of this invention was on the idea that if the movable body were to move in a parallel motion rather than an oscillating motion, it would be possible to free up the constraints of the nozzle outer diameter and use an annular nozzle with a large outer diameter that could achieve a large flow rate.

[0015] Specifically, the invention of claim 4 is such that the movable body is composed of an elastically deformable low-rigidity portion formed on the outer periphery and a high-rigidity portion formed in the center and less susceptible to elastic deformation than the low-rigidity portion, the high-rigidity portion is configured to move approximately parallel to the axial direction of the exhaust-side annular nozzle portion by the driving means, and the area occupied by the high-rigidity portion is approximately a circle ΦD r The outer diameter of the outer valve seat tip of the exhaust side annular nozzle portion is approximated by Φd 0d As, ΦD r >Φd 0dThe movable body and the annular nozzle section are configured so that the following holds true. That is, in the present invention, the movable body is configured, for example, in a disk shape, and a groove, such as a spiral groove, is formed on the outer periphery of the disk, and the area having this groove is made into a low-rigidity portion. The central portion where no groove is formed is made into a high-rigidity portion, and if this high-rigidity portion is magnetically attracted, the high-rigidity portion moves parallel to the axis of the annular nozzle section. Region ΦD of the high-rigidity portion that moves parallel r is the outer diameter of the outer valve seat tip of the exhaust side annular nozzle portion Φd 0d It is easy to make the outer diameter Φd of the exhaust-side annular nozzle portion sufficiently larger than 0d Since the maximum flow rate is determined by the size of the valve, in this invention it is possible to realize a large flow rate valve with Q exceeding 200 L / min with a simple structure comparable to the small flow rate valves (35 to 50 L / min) used for vibration isolation control.

[0016] Specifically, the invention of claim 5 is a fluid control valve in which a closed loop magnetic circuit is formed by an electromagnet, the movable body, and a yoke material, and the attractive force due to Maxwell stress generated between the magnetic poles of the electromagnet and the movable body is used as the driving means for the movable body, wherein the electromagnet is composed of a support shaft made of a magnetic material, a coil wound around the support shaft as its axis, and a cylindrical part made of a magnetic material arranged to accommodate the outer periphery of the coil and the bottom side of the coil on the opposite side to the movable body side, and is provided with a through flow passage formed through the support shaft, one side of which is connected to an opening part of the control valve body, and the annular nozzle is provided on the movable body side of the through flow passage. That is, in the present invention, the electromagnet is composed of the support shaft made of a magnetic material, the coil wound around the support shaft as its axis, and a cylindrical portion made of a magnetic material arranged to accommodate the outer periphery of the coil and the bottom side of the coil opposite the movable body. The support shaft, the movable body, and the cylindrical portion form a closed-loop magnetic circuit, and the magnetic attraction type actuator can move the movable body using the attractive force of the electromagnet. An exhaust passage or an intake passage is formed through the support shaft, and the annular nozzle is attached to the movable body side of this through passage. Because the through passage can have a sufficiently large inner diameter, a large flow control valve can be realized using the annular nozzle and the movable body. If the movable body can move parallel to the plane perpendicular to the axis of the annular nozzle, as described above, an annular nozzle can be used, free from the constraints of the nozzle outer diameter and allowing for a large outer diameter and high flow rate. If the valve structure has an oscillating movable body (flapper) with one end fixed, the outer valve seat tip and the inner valve seat tip of the annular nozzle can be configured to be in close contact with the movable body when the movable body is tilted at the maximum angle Φ. As in the case where the present invention is applied to a conventional oscillating servo valve composed of a permanent magnet and an electromagnetic coil (described later), measures such as (i) installing the annular nozzle axis at an angle Φ, (ii) forming the movable body's surface facing the annular nozzle at an angle Φ, or (iii) forming the movable body's surface facing the annular nozzle at an angle Φ can be taken.

[0017] Specifically, in the invention of claim 6, the annular nozzle equipped on the movable body side of the through-flow passage is composed of an annular nozzle section having the annular flow passage formed by an inner member and an outer member, and a flow passage conversion section connecting the through-flow passage on the opening side, and the annular nozzle section and the flow passage conversion section are arranged in series within the through-flow passage. That is, in this invention, the flow passage formed within the through-flow passage is composed of the annular flow passage, the flow passage conversion section, and the cylindrical through-flow passage. By arranging the annular flow passage and the flow passage conversion section in series, the inner diameter of the through-flow passage can be reduced and the outer diameter of the support shaft can be made smaller. Reducing the outer diameter of the support shaft allows the diameter of the coil bobbin arranged around the support shaft to be reduced, thereby enabling the magnetic attraction type actuator and the control valve body to be made smaller.

[0018] Specifically, in the invention of claim 7, the magnetic pole is formed at the center of the end face of the support shaft on the side of the movable body, and an annular flow passage is formed on the outer periphery of the magnetic pole, with the outer valve seat and the inner valve seat as two peripheral walls. That is, in the present invention, by providing an annular nozzle on the outer periphery of the magnetic pole, the inner diameter d0 of the outer nozzle of the annular nozzle and the outer diameter d i By increasing the exhaust nozzle diameter, which is more restricted than the intake nozzle diameter, the valve rise characteristics, which are determined by the intake flow rate, can be made equivalent to the valve fall characteristics, which are determined by the exhaust flow rate.

[0019] Specifically, in the eighth aspect of the present invention, the outer valve seat and the inner valve seat are made of a non-magnetic material. That is, in this invention, it has been discovered that when a magnetic material is used for the nozzle valve seat that comes into close contact with the movable body, when the tip of the valve seat approaches the movable body (disk) made of a magnetic material, a phenomenon occurs in which the disk is rapidly attracted to the valve seat. The reason for this is that when the gap between the tip of the valve seat and the movable body becomes narrow, the magnetic flux density between the tip of the valve seat and the opposing surface increases locally. The attraction phenomenon does not occur because an appropriate gap is maintained between the magnetic pole and the movable body, which are made of a magnetic material.

[0020] Specifically, the invention of claim 9 includes a bypass passage formed through the inner member, with a first opening formed on the movable body side of the center of the inner valve seat, which is located approximately in the radial center of the inner valve seat, and a second opening formed in the control chamber or a passage connecting the control chamber. That is, in this invention, fluid flowing from the annular passage toward the movable body branches into centrifugal and centripetal directions. Furthermore, the fluid flowing in the centripetal direction flows toward the bypass passage formed through the inner member. The control chamber connected to the bypass passage is connected to a low-pressure pneumatic actuator. It has been found that the formation of the bypass passage, which has two passages in the centrifugal and centripetal directions, significantly increases the flow rate of the control valve of the present invention. Without the bypass passage formed in the inner member, fluid flowing from the annular passage toward the movable body flows only in the centrifugal direction. The reason for this is that the opening of the annular passage connected to the high-pressure source is subjected to a uniform pressure boundary condition in the circumferential direction, so no pressure gradient occurs in the centripetal (radial) direction, and as a result, the flow rate increase effect is reduced compared to the case with a bypass passage.

[0021] Specifically, the invention of claim 10 provides a valve in which the bypass passage is formed in either the first nozzle portion or the second nozzle portion, and a bypass through-hole communicating with the bypass passage is formed in the movable body surface facing the center of the inner valve seat constituting the annular nozzle on the side where the bypass passage is not formed. That is, in this invention, both bidirectional flapper valves are configured with annular nozzles, a bypass passage is formed in one of the annular nozzles, and a bypass through-hole is formed in the movable body facing the two annular nozzles. It was found that even if a bypass passage is formed in only one of the two annular nozzles, the same flow rate increase effect can be achieved for the intake flow rate and the exhaust flow rate as long as the movable body has a bypass through-hole. This is one of the important "bypass effects" discovered in this study. Since only one annular nozzle is required to form a bypass passage, the flow path configuration of the entire control valve can be simplified. The annular nozzle forming the bypass passage may be only on the intake side or only on the exhaust side. In either case, the "annular nozzle with a bypass through-hole and a bypass passage" compensates for this unbalanced combination.

[0022] Specifically, the invention of claim 11 is characterized in that the minimum outer diameter of the inner nozzle in the annular nozzle portion where the bypass flow passage is not formed is d ir The maximum value of the gap between the tip of the inner valve seat and the movable body is X eff The approximate inner diameter of the bypass through-hole is d B As, The bypass through-hole is formed so that the above formula is satisfied. That is, in the present invention, the conditions for the hole diameter required for the bypass through-hole formed in the movable body are shown. Within the range of the above formula, the approximate inner diameter d of the bypass through-hole B By setting the above, the flow rate increase effect (bypass effect) can be obtained for both the intake flow rate and the exhaust flow rate.

[0023] Specifically, in the invention of claim 12, the driving means for the movable body is disposed on the annular nozzle side where the bypass flow path is not formed. That is, this invention focuses on the fact that the flow path configuration can be simplified on the annular nozzle side where the bypass flow path is not formed in a bidirectional nozzle flapper valve. When a bypass flow path is not formed, the outer diameter of the annular nozzle can be reduced, making it easy to use the space around the annular nozzle to arrange an actuator (driving means) that magnetically attracts the movable body. As shown in the first embodiment, this actuator is composed of a support shaft, a coil bobbin that houses a coil, a yoke material that forms a closed Leeb magnetic circuit, and other components. If the bypass flow path were to be formed, an extremely complex structure would be unavoidable, such as dividing the coil bobbin, yoke material, and other components to form a radial bypass flow path. The bypass effect obtained from the "annular nozzle with a bypass through-hole and a bypass flow path" achieves a similar flow rate increase effect (exhaust flow rate increase effect in the first embodiment) even on the side where the bypass flow path is not formed, thereby significantly simplifying the entire control valve.

[0024] Specifically, the invention of claim 13 provides an annular nozzle formed of the inner member and the outer member, in which the bypass passage having a second opening directly connected to the control chamber and a through-passage connecting the annular passage to the supply side or the discharge side are formed independently of each other. That is, in this invention, the following two passages (1) and (2) are incorporated into a single nozzle unit formed of the inner member and the outer member. For example, when this invention is applied to an intake-side annular nozzle, passage (1) is a passage connecting to the control chamber via "air supply source → intake-side passage → annular passage." Passage (2) is a passage connecting to the control chamber via "annular passage → bypass passage → control chamber." These passages are configured by ingeniously designing three-dimensional passage shapes in the radial and axial directions. Because the bypass passage and intake-side passage are incorporated into a single nozzle unit, the installation process into the valve body and the valve structure can be greatly simplified, similar to conventional single-hole nozzles.

[0025] Specifically, in the invention of claim 14, the annular nozzle is composed of a through-flow passage formed perpendicular or inclined to the flow direction of the annular flow passage, an opening in this through-flow passage, and a flow passage converting portion connecting the annular flow passage and the through-flow passage. That is, in the case of a supply-side annular nozzle, for example, the flow passage configuration can be simplified by configuring two flow passages, the through-flow passage connected to the supply source and the bypass flow passage connected to the pneumatic actuator, to intersect at right angles. In addition, the openings connecting the two flow passages to the outside can be provided independently at separate positions. Specifically, the invention of claim 15 is a fluid control valve comprising: a movable body; a first nozzle portion and a second nozzle portion; a movable body support portion for fixing the movable body; and a drive means for the movable body, wherein the drive means is configured so that the movable body approaches a tip end of one of the first nozzle portion and the second nozzle portion and moves away from the tip end of the other, and the fluid control valve controls the difference in flow rate passing through the first nozzle portion and the second nozzle portion, or the flow rate passing through the first nozzle portion and the flow rate passing through the second nozzle portion individually, by movement of the movable body, wherein the first nozzle portion is comprised of an inner member and an outer member that houses the inner member, and comprises: an inner valve seat formed on the movable body side of the inner member; an outer valve seat formed on the movable body side of the outer member; an annular flow path formed by peripheral walls of the outer valve seat and the inner valve seat; and a flow path that connects the annular flow path with a supply hole for working fluid, or a flow path that connects the annular flow path with a discharge hole for working fluid, An annular nozzle is formed that adjusts the opening state between the annular flow path and the movable body by moving the movable body, the outer valve seat tip, and the inner valve seat tip toward or away from each other, and the second nozzle portion is formed from a single-hole nozzle. In other words, in this invention, by applying an annular nozzle to the supply side and a single-hole nozzle to the discharge side, the configuration of the servo valve body is further simplified. If a high flow rate is not required to meet the specifications required by the application target of this valve (pneumatic servo system), a single-hole nozzle can be applied to the discharge side. A single-hole nozzle refers to a nozzle type in which a flow path in the axial direction is formed only in the center.The second nozzle section may be configured by arranging a plurality of single-hole nozzles in parallel, and in this case, the configuration of the discharge side of the servo valve can be sufficiently simplified.

[0026] Specifically, the invention of claim 16 comprises a control chamber, which is a space for accommodating the movable body, and a control port communicating with the control chamber, wherein the working fluid flows from the supply hole through the first nozzle portion, which is a supply-side nozzle, into the control chamber, and flows from the control chamber through the second nozzle portion, which is a discharge-side nozzle, to the discharge hole; and the fluid pressure in the control chamber or the flow rate flowing in and out of the control port is controlled by the difference in flow rate through the first nozzle portion and the second nozzle portion. That is, in this invention, a bidirectional nozzle flapper valve is configured by using the first nozzle portion as a forward-direction nozzle and the second nozzle portion as a reverse-direction nozzle. Furthermore, it has been discovered that configuring both nozzles in the bidirectional nozzle flapper valve as the annular nozzle can eliminate the following factors that inhibit an increase in flow rate: (i) Reduction in the effective displacement of the movable body due to a fail-safe mechanism; and (ii) The phenomenon of the movable body adhering to the exhaust nozzle. Here, the problems with conventional control valves using a single-hole nozzle can be summarized as follows. Increasing the diameter of the intake nozzle limits the increase in intake flow rate because the effective displacement of the moving body due to the fail-safe mechanism decreases. The fail-safe mechanism is an essential safety feature that blocks the control valve's intake port to prevent high-pressure gas from entering the actuator when the input current I applied to the control valve is zero or during emergencies such as power outages. Furthermore, increasing the exhaust nozzle diameter to increase the exhaust flow rate can cause the moving body to adhere to the exhaust nozzle, rendering the nozzle flapper valve inoperable. The force of adhesion of the moving body to the exhaust nozzle is proportional to the area of ​​the exhaust nozzle and the pressure difference between the supply pressure and atmospheric pressure. To resolve this issue, increasing the rigidity of the moving body reduces the fail-safe displacement, increases the effective displacement, and increases the intake flow rate. Furthermore, increasing the restoring force of the moving body also increases the limit of the exhaust nozzle diameter d at which the exhaust flow rate can be increased. However, in order to increase the rigidity of the moving body, the force generated by the electromagnet had to be increased, which resulted in a deterioration in responsiveness due to a decrease in the electrical and mechanical time constant, and an increase in the capacity of the power supply, which required modifications to the electrical circuits, magnetic circuits, and mechanical parts that make up the entire control valve, resulting in a significant increase in costs.When the annular nozzle is applied to the forward direction (intake side) nozzle, the fail-safe displacement X is smaller than that of a conventional single-hole nozzle under the same nozzle outer diameter (d0 = d). fel Since the effective stroke of the movable body can be increased, the intake flow rate can be significantly increased. When an annular nozzle is used for the reverse direction (exhaust side) nozzle, the attraction force F P is much smaller than that of a conventional single-hole nozzle. Over a wide range of outer nozzle inner diameters d0, the phenomenon of movable body adhesion to the exhaust nozzle does not occur, allowing for ample increases in exhaust flow rate. When annular nozzles are used for both the forward and reverse nozzles, the flapper rigidity can be maintained at the same level as that of a small-flow valve, and a valve with large intake and exhaust flow rates can be realized simply by improving both nozzles. The basic performance required of a control valve is to shorten the time it takes for the actuator to fill with air while also shortening the time it takes to exhaust the air. To achieve this, it is necessary to increase both the maximum intake flow rate and the maximum exhaust flow rate. Therefore, when a pneumatic control system is constructed using a control valve of the present invention in which annular nozzles are used for both the forward and reverse nozzles, the rise time constant and fall time constant can be set to similarly small values.

[0027] Specifically, the invention of claim 17 is such that the opening area of ​​the single-hole nozzle is A d , the support stiffness of the movable body is K0, the maximum displacement of the movable body is X ma , the source pressure is P s , the discharge pressure of the fluid control valve is P0, A d <K0X max / (P s -P0) in the range where the conditional expression is satisfied, the single-hole nozzle opening area A d That is, in the present invention, the flapper has a maximum displacement X = X max The exhaust nozzle is blocked and the control chamber is at maximum pressure, supply pressure P max =P s The inner diameter of the discharge nozzle is d, the atmospheric pressure (or discharge pressure) is P0, and the opening area of ​​the single-hole nozzle is Ad Then, the pressure difference between the control room and the atmospheric pressure is ΔP d (=P s -P0) flapper suction force F d = A d ×ΔP d The spring stiffness of the flapper is K0, and the restoring force of the flapper is F max = K0X max Therefore, within the range where the restoring force of the flapper is greater than the suction force of the flapper, the single-hole nozzle opening area A d If you set A, the adsorption phenomenon of the exhaust nozzle to the flapper can be avoided. d <K0X max / (P s -P0) within the range where the conditional expression of the discharge nozzle opening area A d If the discharge nozzle is composed of multiple single-hole nozzles, the opening area A d is the sum of the nozzle opening areas.

[0028] Specifically, in the invention of claim 18, the space whose outer peripheral wall is the inner valve seat of the first nozzle portion is defined as the supply-side gap, and a perforation portion connecting the supply-side gap and the control chamber is formed on the movable body surface at a position radially spaced from the opening of the single-hole nozzle. That is, in this invention, the small-diameter perforation portion formed in the flapper is formed at a position radially spaced from the opening of the discharge-side nozzle and facing the inner peripheral side of the supply-side gap. By forming the perforation portion, an effect of preventing an increase in the fail-safe force is obtained when the flapper is in close contact with the supply-side annular nozzle. If the perforation portion is not formed in the flapper, the pressure in the supply-side gap will be equal to the supply pressure P S The reason is that the tip of the outer valve seat is completely shielded by the flapper, so the pressure in the supply side gap adjacent to the annular flow path easily becomes the supply pressure P S The area A of the supply-side gap rsi Then, the pressure difference between the control chamber and the supply pressure ΔP s A large force F that pulls the flapper away from the supply nozzle is generated. si= A rsi ×ΔP s However, by forming the perforated portion, F si = 0, it is possible to prevent the increase of the fail-safe force.

[0029] Specifically, the invention of claim 19 includes a bypass flow path formed through the inner member, with a first opening formed on the movable body side of the center of the inner valve seat, which is provided approximately in the radial center of the inner valve seat, and a second opening formed in the control chamber or in a flow path connecting the control chamber. That is, in this invention, fluid flowing out from the annular flow path toward the movable body branches into a centrifugal direction and a centripetal direction. Furthermore, the fluid flowing out in the centripetal direction flows into the bypass flow path formed through the inner member. The control chamber connected to the bypass flow path is connected to a low-pressure pneumatic actuator. It has been found that the formation of the bypass flow path with two flow paths, one in the centrifugal direction and the other in the centripetal direction, allows the control valve of the present invention to achieve a significant increase in flow rate.

[0030] Specifically, in the invention of claim 20, the first nozzle portion and the second nozzle portion are both composed of annular nozzles having only flow paths formed therein in the same axial direction, the space having the inner valve seat of the annular nozzle on the supply side as its outer peripheral wall is an intake-side gap, and the space having the inner valve seat of the annular nozzle on the discharge side as its outer peripheral wall is a discharge-side gap, and perforations connecting the intake-side gap and the discharge-side gap to the control chamber are formed in the movable body surface. That is, in this invention, the small-diameter perforations formed in the flapper are formed to (i) prevent the flapper from adhering to the discharge nozzle, and (ii) prevent the supply-side gap 710 from increasing the fail-safe force, achieving the effects of (i) and (ii) above. The effect of (i) above is as follows. Even when the annular flow passage on the discharge side is blocked by the close contact between the movable body and the annular nozzle on the discharge side, the discharge gap formed in the center of the inner valve seat is connected to the perforated portion, so that the discharge gap can maintain the pressure in the control chamber. In this case, the opening cross-sectional area of ​​the annular flow passage on the discharge side is set to A rd , the pressure difference between the control room and the atmospheric pressure is ΔP dThen, F d = A rd ×ΔP d It is sufficient to simply form a hole in the movable body. In this case, the tip of the outer valve seat is completely sealed by the movable body, and the pressure in the gap adjacent to the annular flow path easily drops to atmospheric pressure. The inner diameter of the tip of the outer valve seat is the inner diameter of the discharge-side outer nozzle d0, and the area A rd0 =(d0 / 2) 2 If π, the pressure difference between the control room and atmospheric pressure is ΔP d The adhesive force F of the movable body d0 = A rd0 ×ΔP d A rd0 ≫A rd Therefore, the adhesive force of the movable body F d0 ≫F d Therefore, even if the drive current is changed while the movable body is in close contact with the discharge-side annular nozzle, this close contact state cannot be avoided, and the control valve becomes uncontrollable. In other words, when an annular nozzle is applied to the discharge side, the presence of the perforated portion is extremely effective in avoiding this suction phenomenon. As for the effect of (ii) above, similar to claim 17, by forming the perforated portion, a large force F that separates the flapper from the supply-side nozzle can be generated. si = 0, it is possible to prevent the increase of the fail-safe force.

[0031] Specifically, the invention of claim 21 provides that the annular nozzle comprises an inner portion and an outer portion formed on the outer periphery of the inner portion, and the annular nozzle comprises an inner valve seat formed on the movable body side of the inner portion, an outer valve seat formed on the movable body side of the outer member, an annular flow path formed by the peripheral walls of the outer valve seat and the inner valve seat, and multiple connecting portions formed to prevent separation of the inner portion and the outer portion. In other words, unlike the previously described embodiments which were composed of multiple parts, the present invention constructs the annular nozzle from a single part using a production method such as wire-cut electrical discharge machining. The annular flow path is composed of a complete slit penetrating in the axial direction and a local slit with a joint in the axial direction. This invention achieves the benefits of applying an annular nozzle (high flow rate) with a simple configuration similar to that of a conventional control valve composed of a single-hole nozzle.

[0032] Specifically, the invention of claim 22 is a fluid control valve comprising: a movable body; a nozzle portion disposed opposite the movable body; an inlet of the nozzle portion connected to a supply source of working fluid; a movable body support member for fixing the movable body; and drive means for the movable body, wherein the drive means changes the separation distance between the tip of the nozzle portion and the movable body, thereby controlling the flow rate of working fluid flowing out of the nozzle portion, wherein the nozzle portion is composed of an inner member and an outer member that houses the inner member, and comprises: an inner valve seat formed on the movable body side of the inner member; an outer valve seat formed on the movable body side of the outer member; an annular flow path formed by peripheral walls of the outer valve seat and the inner valve seat; and a flow path that connects this annular flow path with a supply hole for working fluid or a discharge hole for working fluid, wherein the movable body and the tip end of the outer valve seat and the tip end of the inner valve seat come close to or move away from each other to form an annular nozzle that adjusts the opening state between the annular flow path and the movable body, The movable body is composed of a valve body center portion, which is the opposing surface of the valve seat tip portion of the annular nozzle, and an elastically deformable portion on the outer periphery of the valve body center portion, in which a through groove is formed. The axis of the movable body is defined as the Z axis, and coordinates of the X axis and the Y axis are defined relative to this Z axis. The axial spring stiffness in the Z axis direction is defined as K Z , and Xθ Direction, Y θ The torsional spring stiffness in the direction is K θX , K. θY The axial spring stiffness and the torsional spring stiffness are set so that when the valve element central portion is pressed against the valve seat tip of the annular nozzle by the driving means, the valve element central portion, which is the surface facing the valve seat tip, can deform three-dimensionally. That is, in the present invention, even if the valve seat tip of the annular nozzle is slightly tilted relative to the disk surface due to errors in processing accuracy and assembly accuracy, when the valve element central portion is pressed against the valve seat tip of the annular nozzle, the elastic deformation portion having the through groove deforms three-dimensionally, thereby achieving a sealing effect that allows the valve element central portion, which is the surface facing the annular nozzle, to closely fit the shape of the valve seat tip...which can also be called a compliance effect. This sealing performance is effective when an annular nozzle with a large outer diameter is used. The reason for the excellent sealing performance is that the valve element central portion not only moves parallel in the Z-axis direction but also moves in the X-axis direction. θ Direction, Y θ This is because it deforms with the same high level of compliance (flexibility) in both directions. Compliance is the inverse of spring stiffness, and is a physical quantity that quantifies the elasticity and ductility of an object. For example, in the X direction, θ The torsional rigidity in the direction is K θX Then, λ θX =1 / K θX Servo valves consisting of a thin disk (movable body) and a single-hole nozzle with a small inner diameter are well known. However, a control valve structure that utilizes compliance to improve sealing performance for an annular nozzle with a large outer diameter is a new discovery. The movable body used in conventional oscillating servo valves is configured as a rigid body, and deforms only in the same angular direction as the oscillating motion, but the above-mentioned compliance effect, which requires three-dimensional deformation, cannot be obtained.

[0033] Specifically, the invention of claim 23 is such that, when the approximate radius of the central part of the valve body is r, the K θX and the aforementioned K θY Compare the values ​​and choose the larger one as K. θ Represented as The elastic deformation portion is configured to satisfy the above condition. That is, in the present invention, the torsional torque T applied to the disc causes the center point of the valve body central portion to deform Δz in the Z-axis direction and tilt at an angle Δθ, and conditions for obtaining a compliance action that improves sealing performance are sought. Details are described in Supplementary Note [1-1], but the Z-axis spring stiffness K is set to satisfy the above condition. z , the spring stiffness in the torsional direction K θ By setting the above, the central portion of the valve body is more likely to deform in the torsional direction than in the axial direction, and the sealing action of tightly sealing the tip end of the valve seat can be reliably obtained.

[0034] Specifically, the invention of claim 24 provides a fluid control valve comprising: a supply hole with a throttle; an intermediate control chamber into which the working fluid that has passed through the supply hole with the throttle flows; a control port that communicates the intermediate control chamber with the outside; a nozzle portion into which the working fluid that has passed through the intermediate control chamber flows; a movable body; a discharge chamber that is a space into which the working fluid that has passed through the nozzle portion flows and that accommodates the movable body; a fluid discharge hole that communicates with the exhaust chamber; a movable body support portion that fixes the movable body; and a drive means for the movable body, wherein the drive means is configured to move the movable body between a position close to the tip of the nozzle portion and a position away from it, and the drive means changes the distance between the tip of the nozzle portion and the movable body, thereby controlling the fluid pressure in the intermediate control chamber, or the flow rate flowing from the supply hole into the intermediate control chamber or the flow rate flowing from the intermediate control chamber to the discharge hole, The nozzle section is composed of an inner member and an outer member that houses the inner member, and is provided with an inner valve seat formed on the movable body side of the inner member, an outer valve seat formed on the movable body side of the outer member, an annular flow path formed by peripheral walls of the outer valve seat and the inner valve seat, and a flow path connecting the annular flow path to the supply hole or the annular flow path to the discharge hole, and an annular nozzle is configured that adjusts the opening state between the annular flow path and the movable body by moving the movable body and the tip end of the outer valve seat and the tip end of the inner valve seat closer to or farther away from each other. That is, in the present invention, the annular nozzle has a higher opening state than a single-hole nozzle at the same supply source pressure P S This is because the air resistance can be significantly reduced under the same conditions of clearance h. A weakness of conventional one-way three-way valves (two-way four-way valves) that are configured with a single-hole nozzle is that the flow rate control range is narrow, but the present invention makes it possible to realize a servo valve with a wide flow rate range while maintaining the basic configuration of the conventional one-way three-way valve.

[0035] Specifically, in the twenty-fifth aspect of the present invention, the throttled supply hole is configured with a rectifier for smoothing turbulent fluid flow. That is, in this invention, the single throttled intake hole is replaced with a rectifier. This rectifier is provided to suppress turbulence, vortexes, and shock waves generated by restricting the gas flow through an orifice, i.e., to smooth the turbulent gas flow. As a result, the control chamber connected to the control port can maintain a stable control pressure without noise.

[0036] Specifically, the invention of claim 26 has a bypass flow passage formed through the inner member, a first opening of this bypass flow passage formed on the movable body side of the center of the inner valve seat, and a second opening of this bypass flow passage formed in the discharge chamber or in a flow passage connecting the discharge chambers. That is, in this invention, by providing a bypass flow passage in an annular nozzle, the flow rate increase ratio η relative to a single-hole nozzle can be further increased compared to a case in which a bypass flow passage is not formed. Since the air resistance passing through the nozzle can be set to be small at the same ratio η, the flow rate can be further increased.

[0037] Specifically, the invention of claim 27 is such that when an input signal is first applied to the driving means of the movable body, the gap between the movable body and the tip of the annular nozzle portion is h=0, and when the gap h is increased, the flow area of ​​the centrifugal flow path between the tip of the annular nozzle portion and the movable body is A R , the axial cross-sectional area of ​​the annular nozzle flow path is A Z , A R = A Z The gap where the flow direction is switched is defined as h = h0, and the gap where the valve flow rate is saturated at a constant value is defined as h = h S Then, h0≦h max <h S Maximum stroke h within the range maxis set. That is, in the present invention, as a result of consideration using computational fluid analysis, it was found that the characteristic curve of the annular nozzle valve flow rate is composed of three regions. (i) The region 0≦h<h0 is the region where the valve flow rate Q relative to the gap increases linearly. When the annular nozzle has an axisymmetric structure, the fluid flows two-dimensionally between the valve seat and the valve disc (disk) only in the radial direction. (ii) h0≦h<h S In the region, the gradient of the flow rate Q with respect to the gap h gradually decreases from point A (h=h0) and eventually reaches saturation point B (h=h S ) The air flow between the valve seat and the valve body is a three-dimensional flow in which the axial flow is added to the radial flow. This region is called the "transition region." (iii) h ≥ h S The region where the valve flow rate is saturated and becomes a constant value is shown in Figure 1. Based on the above results, the maximum stroke h of the servo valve using the annular nozzle is max in the transition region (h0≦h<h S ) Maximum stroke h max By setting within the above range, sufficient valve flow rate can be obtained while keeping the intake nozzle opening area the same, i.e., maintaining the same fail-safe force. As mentioned above, the fail-safe force is the spring restoring force of the disk required to block the intake nozzle before the valve operates. Increasing the spring stiffness to increase the spring restoring force requires increasing the power of the electromagnet, but h max As long as the above range is set, these modifications do not need to be made.

[0038] Specifically, the invention of claim 28 is an envelope a of a flow characteristic curve in the range of 0≦h<h0, an envelope b of a region where the flow rate is a constant value, and a gap h=h at the intersection of the envelope a and the envelope b. C Then, h0≦h max <h C Maximum stroke h within the range max That is, in the present invention, the maximum stroke h maxIf is set within the above range, the characteristics of the flow rate Q with respect to the gap h can be maintained to be almost linear. Therefore, in vibration control, this is advantageous for creating an accurate vibration suppression force waveform to attenuate the excitation force (disturbance) applied to the vibration isolation table.

[0039] Specifically, the invention of claim 29 provides a bypass flow passage, wherein the approximate inner diameter of the outer valve seat tip is an outer nozzle inner diameter d0, and the approximate outer diameter of the inner valve seat tip is an inner nozzle outer diameter d i Then, the flow path direction switching gap is h0 = (d0 - d i ) / 4. That is, in the present invention, the value of the gap h=h0 at which the gradient of the flow rate relative to the gap switches is determined for the case where a bypass flow path is formed. The fluid flowing out of the annular nozzle branches into two directions, toward the outer valve seat (centrifugal direction) and toward the inner valve seat (centripetal direction). The inner diameter d0 of the outer nozzle of the annular nozzle and the outer diameter d of the inner nozzle of the annular nozzle are i Assuming that the flow area A in the range of h < h0 R teeth Flow area A in the range of h>h0 Z is the cross-sectional area S(=A Z ) is assumed to be determined by A R = A Z Then, the flow path direction switching gap h = h0 = (d0 - d i ) / 4 is obtained.

[0040] Specifically, the invention of claim 30 comprises the configuration of the fluid control valve of claim 22 as a first configuration, and further comprises, similar to the fluid control valve, a second restrictor-equipped supply hole, a second control chamber, a second control port, a second nozzle portion, a second movable body, a second exhaust chamber, a second discharge hole, and a second movable body support portion as a second configuration, wherein, when driven by the driving means of the first configuration, the nozzle portion and the movable body of one of the first configuration and the second configuration are brought closer to each other, and the nozzle portion and the movable body of the other configuration are separated from each other, and by changing the separation distance between the tip of the nozzle portion and the movable body by the driving means, a fluid control valve is configured to control the fluid pressure in the control chamber, or the flow rate flowing into the control chamber from the restrictor-equipped supply hole, or the flow rate flowing out from the control chamber to the discharge hole. That is, in the present invention, the two-way (bidirectional) four-way valve has a flow path consisting of two-way nozzles (a forward nozzle and a reverse nozzle) and four openings (an intake port, an exhaust port, and two control ports). The components formed and attached to the left and right intake housings of a two-way four-way valve are often identical, and the two-way four-way valve can be considered a mirror image of the one-way three-way valve described above. The one-way three-way valve can similarly utilize the effects of increasing the flow rate by using an annular nozzle and maintaining a stable, noise-free control pressure by using a restrictor-equipped supply port as a flow rectifier. A two-way four-way valve is effective, for example, when controlling the pressure and flow rate of the left and right air chambers on either side of the piston in opposite phases in a pneumatic actuator consisting of a piston, a rod directly connected to the piston, and a cylindrical cylinder that houses them. Alternatively, independent pneumatic actuators can be placed on the floor side and the base side, and each pneumatic actuator can be driven by a two-way four-way valve.

[0041] Specifically, the invention of claim 46 is a fluid control valve comprising: a supply-side nozzle portion; a movable body; a movable body support portion for fixing the movable body; and a drive means for the movable body, wherein the drive means controls the flow rate passing through the supply-side nozzle portion by varying the distance between the movable body and the tip of the supply-side nozzle portion through movement of the movable body, and the supply-side nozzle portion is composed of a supply hole, a throttling portion for reducing the pressure of the working fluid that has passed through the supply hole, and a supply nozzle into which the working fluid that has passed through the throttling portion flows, and the supply nozzle is composed of an inner member and an outer member that houses the inner member, and comprises: an inner valve seat formed on the movable body side of the inner member; an outer valve seat formed on the movable body side of the outer member; and an annular flow path formed by the outer valve seat and the peripheral wall of the inner valve seat, and the movable body and the outer valve seat tip end move close to or apart from each other to form an annular nozzle that adjusts the opening state between the annular flow path and the movable body. In other words, the present invention proposes a basic principle for avoiding the generation of turbulence in the fluid that has passed through the variable nozzle and for achieving laminar flow. In summary, (1) the gap between the nozzle and the disk is the same as that of a single-hole nozzle, and the opening area A is large. e Focusing on an annular nozzle that can obtain the above, this annular nozzle is designated as a variable flow nozzle. (2) To make the flow rate Q the target flow rate, for example, a fixed throttle is placed upstream of the annular nozzle to lower the upstream pressure. (3) If the Reynolds number is used as an evaluation index, then "lower flow velocity = laminar flow." Flow velocity V=Q / A e and the opening area A e The "slowing down action" of the large annular nozzle can laminarize the fluid in the valve control chamber. Specifically, the working fluid is configured to flow from the supply hole through the supply-side nozzle section into the control chamber, which is the space housing the movable body. The fluid control valve controls the flow rate through the supply-side nozzle section by varying the distance between the movable body and the tip of the supply-side nozzle section. The supply-side nozzle section is composed of a throttle section formed from a narrow flow path and an annular nozzle into which the working fluid flows after passing through the throttle section. A feature of the present invention is that the supply-side nozzle section is configured in the following order: supply hole → throttle section formed from a narrow flow path → annular nozzle. The "throttle section" with a small flow path gap reduces the air flow rate, lowering the pressure upstream of the annular nozzle. Furthermore, the annular nozzle with a large flow path opening area slows down the air flow velocity. The deceleration action of the annular nozzle has the effect of laminarizing the air flow. In summary, while the conventional proposal is a "rectifier" (low-pass filter) that reduces air pulsation after passing through a fixed nozzle, the present invention is a "decelerator" that utilizes a variable nozzle, specifically an annular nozzle with a large aperture area. The effect of the present invention, which reduces the Reynolds number of the fluid ejected from the variable nozzle (annular nozzle) to laminarize or approach laminar flow, is applicable regardless of the type or shape of the valve, such as a one-way three-way valve or a two-way three-way valve. In other words, the valve of the present invention can be applied to a fluid circuit incorporating the variable nozzle as a fluid control element, and when laminar flow downstream of the variable nozzle is required. In this case, the upstream side of the supply hole does not necessarily have to be connected to a constant-pressure fluid supply source. The intake flow path, configured in the order of "the throttle section → the annular nozzle," may be incorporated into a larger fluid control system.

[0042] Specifically, the invention of claim 47 includes an exhaust-side nozzle disposed opposite the supply-side nozzle across the movable body; and the driving means is configured so that the movable body approaches the tip of one of the supply-side nozzle and the discharge-side nozzle and moves away from the tip of the other, and is configured to control the difference in flow rate passing through the supply-side nozzle and the discharge-side nozzle by movement of the movable body. That is, in this invention, the working fluid is configured to flow from the supply hole through the supply-side nozzle into a control chamber, which is a space that houses the movable body, and then flow from this control chamber through the discharge-side nozzle to the discharge hole. That is, this is a control valve that controls the fluid pressure in the control chamber or the flow rate flowing in and out of the control port based on the difference in flow rate passing through the supply-side nozzle and the discharge-side nozzle. Assume that the invention is applied to, for example, a two-way three-way valve used in a pneumatic active vibration isolation table. A characteristic of a two-way three-way valve is that the control pressure of the control chamber can be used over the full range from atmospheric pressure to supply pressure, and a large control flow rate can be obtained. An active vibration isolation table is required to have both the function of isolating vibrations from the floor surface (vibration isolation performance) and the function of controlling the vibration of linear disturbances by controlling the air spring pressure (vibration control performance), that is, the function of having both "vibration isolation" and "vibration control." Therefore, if the present invention is applied as a servo valve for an active vibration isolation table, it is possible to maintain a large flow rate during vibration control, and to achieve minute flow rate and high-resolution control that exceeds conventional limits during vibration isolation control.

[0043] Specifically, the invention of claim 48 is such that the space between the peripheral wall of the outer valve seat and the peripheral wall of the inner valve seat is constituted by an opening penetrating in the axial direction and a shielding portion having a sealing portion in the axial direction. That is, in this invention, an annular nozzle is constituted by forming discontinuous arc-shaped flow paths having an opening penetrating in the axial direction and a shielding portion having a sealing portion in the axial direction. By setting the ratio of the openings to the entirety to an appropriate value without changing the outer diameter of the annular nozzle, the maximum flow rate can be set to a desired flow rate. In the case of an annular nozzle formed with a continuous ring-shaped flow path, the maximum flow rate is determined by the inner diameter d of the outer nozzle. o and the outer diameter of the inner nozzle d iHowever, if the present invention is applied, the outer diameter of the annular nozzle (d o , d i ) is sufficiently large, the degree of freedom in the overall valve design increases, allowing the valve flow rate to be set over a wide range.

[0044] Specifically, the invention of claim 49 is such that the angle of the arc portion where the opening is formed is θ a The angle of the arc portion where the shielding portion is formed is θ b Then, the opening angle ratio η = θ a / (θ a +θ b ) is defined, and the maximum value of the distance X between the movable body and the tip of the supply-side nozzle portion is defined as the distance X max As such, X=X max Near the valve maximum flow rate Q max The aperture angle ratio η is set so that the following is obtained. That is, in the present invention, it is assumed that the angles of the annular flow path apertures and the annular flow path shielding portions are equally divided in the circumferential direction. θ a is the opening angle of the annular flow path, θ b is the shielding angle of the annular flow path, and the opening angle ratio η is defined. Gap X=X max The maximum flow rate of the valve Q max The aperture angle ratio η = η o Let η>η o In this case, the flow rate will be excessive, so the gap X < X max The control flow rate is the maximum flow rate Q max η<η o Then, the flow rate will be insufficient, so the gap X=X max The control flow rate is the maximum flow rate Q max cannot be obtained.

[0045] Specifically, in the invention of claim 50, the annular nozzle is disposed on the supply side and the single-hole nozzle is disposed on the discharge side, sandwiching the movable body, and a through-hole is formed in the movable body between the outer periphery of the single-hole nozzle opening and the inner periphery of the inner valve seat of the annular nozzle. That is, in this invention, an intake flow path is formed on the intake side using an intake-side throttle portion and annular nozzle, and a two-way three-way valve (or one-way two-way valve) is formed on the discharge side using a single-hole nozzle that is simple and can be manufactured at low cost. However, in this configuration, a through-hole is required in the flapper, and therefore an annular nozzle with a large outer diameter is required due to the constraints shown in (1) and (2) below. (1) Positional constraints of the through-hole when input current I = 0 (2) Input current I = I max In summary, in the case of an annular nozzle in which the flow passages are formed continuously in the circumferential direction, the flow rate is determined by the outer diameter d of the inner nozzle. i and the inner diameter of the outer nozzle d o The present invention solves this problem by making the flow path of the annular nozzle arc-shaped, setting the division angle ratio η between the opening portion and the shielding portion to a predetermined value, and forming a through hole between the annular nozzle and the inner peripheral side of the inner valve seat, thereby making it possible to set an optimal flow rate even when using an annular nozzle with a large outer diameter.

[0046] Specifically, in the invention of claim 51, the throttle portion of the supply-side nozzle portion is configured as a single-hole nozzle. That is, in this invention, the throttle portion of the supply-side nozzle portion is configured as a single-hole nozzle that is simple and can be manufactured at low cost. Even in the case of the single-hole nozzle with isentropic flow without energy loss, the upstream pressure of the annular nozzle can be reduced relative to the supply pressure, thereby achieving the effect of reducing the flow velocity of the annular nozzle.

[0047] Specifically, in the invention of claim 52, the throttle portion in the supply side nozzle portion is configured with a single-hole nozzle, and the supply side nozzle is configured with an annular nozzle configured with an arc-shaped flow path. That is, in this invention, a low-noise valve is configured by combining an "intake side throttle portion with a single-hole nozzle" and an "annular nozzle with an arc-shaped flow path." It has been found that when the intake side throttle portion is configured with a low-cost, simple single-hole nozzle, when the flow rate Q reaches a critical state (a constant value), the pressure Pa in the intermediate control chamber upstream of the annular nozzle cannot be sufficiently reduced relative to the supply pressure Ps. As a result, the effective flow area A eff An annular nozzle with a large arc-shaped flow path will result in an excessive flow rate. However, by using an annular nozzle configured with the arc-shaped flow path, the excessive flow rate can be suppressed.

[0048] Specifically, the invention of claim 53 is characterized in that the critical pressure ratio of the working fluid is γ c , source pressure P s The critical pressure for c =γ c P s The downstream pressure P of the throttle portion a Flow rate Q characteristics and downstream pressure P a C against d =dQ / dP a In characteristics, P a =P c Under the conditions of C d > 0. That is, in this invention, in a low noise valve configured with an intake flow path of "throttle section formed from a narrow flow path" → "annular nozzle", attention is paid to the nozzle characteristics of the "throttle section". The "throttle section" with a small flow path gap reduces the air flow rate, lowering the pressure upstream of the annular nozzle. Furthermore, the annular nozzle with a large flow path opening area slows down the air flow velocity. As mentioned above, this deceleration action by the annular nozzle has the effect of laminarizing the air flow (reducing noise). The method for experimentally evaluating the applicability of this invention is as follows. The downstream pressure P of the throttle section a The flow rate Q characteristic is calculated for the downstream pressure P a Rate of change of flow rate C d =dQ / dP aThe characteristics of the supply pressure P s The critical pressure for c =γ c P s When a single-hole nozzle is used in the throttle section, the air flow rate does not increase due to the choking phenomenon, and in order to maintain a constant value (critical flow rate), P a =P c Under the condition, the flow rate change rate C d = 0. Due to this choking phenomenon, the downstream pressure P a The back pressure P a Even if the nozzle characteristic is further reduced, the flow rate cannot be increased any further. A nozzle that does not cause choking is ideal, but in practice, if a sufficient pressure drop is obtained, it can be applied to the low noise valve of the present invention. a =P c Place C d If it is >0, it can be used as a throttle portion for a low noise valve.

[0049] Specifically, the invention of claim 54 describes a nozzle in which, when a pressure difference is applied between the nozzle opening and outlet, the relationship between this pressure difference and the flow rate is governed solely by the viscosity of the working fluid (e.g., characteristic A of a nozzle whose flow path is formed by a narrow gap), and a nozzle in which the relationship is not affected by the viscosity of the working fluid (e.g., characteristic B of a single-hole nozzle). A nozzle having characteristics intermediate between characteristics A and B is defined as a pseudo-viscous nozzle, and the throttle portion of the first nozzle portion is constructed using this pseudo-viscous nozzle. That is, this invention anticipates application to throttle portions in intake flow paths and focuses on nozzle characteristics in the "transition region" between inviscid flow and viscous flow. For example, a fluid passing through a long, narrow flow path with a gap of 10 to 20 μm (nozzle with characteristic A) is a viscous flow that is predominantly affected by viscous loss, resulting in a large pressure drop but a limited flow rate. To achieve both a large pressure drop and sufficient flow rate, a throttle portion composed of, for example, many thin tubes with small inner diameters is required. In this case, the overall external dimensions of the throttle portion become larger. On the other hand, a fluid passing through a simple single-hole nozzle (nozzle with characteristic B) with a short flow path length experiences isentropic flow (inviscid flow) without energy loss, but does not achieve a significant pressure drop compared to viscous flow. In this study, we refer to nozzles with the "transition region" characteristics between inviscid and viscous flow as "pseudo-viscous nozzles." This pseudo-viscous nozzle maintains the required flow rate with a simple structure and significantly reduces the pressure upstream of the annular nozzle. As a result, it reduces the annular nozzle outlet flow velocity and Reynolds number. In other words, we discovered that this method can dramatically improve noise reduction in servo valves with strict external dimensional constraints. Furthermore, when a pseudo-viscous nozzle is used, the upstream side of the annular nozzle maintains a stable pressure source state without pressure fluctuations. Pressure fluctuations in the upstream pressure of the annular nozzle result in fluctuations in the controlled flow rate and controlled pressure. It was found that when a pseudo-viscous nozzle is used in the orifice, the generation of turbulence after passing through the orifice is significantly suppressed compared to when a single-hole nozzle is used in the orifice. The damping effect of the fluid viscosity passing through the narrow gap of the pseudo-viscous nozzle has the effect of suppressing pressure fluctuations caused by turbulence.

[0050] Specifically, in the invention of claim 55, the throttle portion formed by the pseudo-viscosity nozzle is housed inside the housing of the servo valve. That is, in this invention, the pseudo-viscosity nozzle having nozzle characteristics in the "transition region" between non-viscous flow and viscous flow can be made sufficiently small, and is therefore housed inside the housing of the servo valve, which has strict restrictions on external dimensions.

[0051] Specifically, in the invention of claim 56, the annular nozzle and the throttle portion are formed coaxially. That is, this invention focuses on the fact that when the annular nozzle portion and the intake-side throttle portion are arranged apart, the time constant T = RC of the air circuit is a measure of responsiveness, where R is the fluid resistance of the intake-side throttle portion and C is the volume connecting the annular nozzle portion and the intake-side throttle portion. When high responsiveness is required of a servo valve, the time constant T becomes non-negligible. By arranging the annular nozzle portion and the intake-side throttle portion coaxially and bringing the volume C as close to zero as possible, noise reduction can be achieved by utilizing the deceleration effect (laminar flow effect) of the annular nozzle without reducing the responsiveness of the servo valve.

[0052] Specifically, the invention of claim 57 is such that the throttle portion is externally disposed to a housing that houses main components such as the movable body, the supply nozzle, and the drive means for the movable body, and the opening portion of this throttle portion on the fluid supply source side is used as the supply hole. In other words, this invention shows that by externally disposing the supply-side throttle portion on the servo valve, it is possible to free from dimensional constraints on the flow path shape and realize a more ideal low-noise valve.

[0053] Specifically, the invention of claim 58 provides a pseudo-viscous nozzle configured by arranging multiple narrow gap flow paths in parallel. In other words, the present invention utilizes a pseudo-viscous nozzle that lies somewhere between a viscous nozzle, which is governed solely by the viscosity of the working fluid, and a non-viscous nozzle, which is not affected by the viscosity of the working fluid. When a nozzle is configured with a single flow path, the narrower the gap, the closer its characteristics become to a viscous flow, resulting in a large pressure drop but limiting the flow rate. The present invention focuses on the fact that providing multiple narrow flow paths in parallel can compensate for insufficient flow rate.

[0054] Specifically, the invention of claim 59 relates to a servo valve having an unprecedented low noise level, in which the supply hole, the throttle portion, the annular nozzle, a flow passage connecting the throttle portion and the annular nozzle, a control port, etc. are housed in a supply-side housing, the outer member constituting the annular nozzle is integrally formed with the supply-side housing, and the inner member is housed and disposed in the center of the outer member. That is, in this invention, by focusing on the axisymmetric shape of the outer valve seat of the annular nozzle and pre-machining this outer valve seat integrally with the intake-side housing, an annular nozzle is constructed with a simple structure comparable to conventional valves. This proposal makes it possible to realize an unprecedented low-noise servo valve by simply adding one major part while maintaining the same external shape and dimensions as conventional servo valves.

[0055] According to the present invention, a high-flow rate valve or a low-noise valve can be realized with high responsiveness and a simple structure.

[0056] [Correction based on Rule 91 03.03.2025] A model diagram of a fluid control system with feedforward vibration suppression control, which is the first embodiment of the present invention. Figure 1a is a front cross-sectional view, and Figure 1b is a view taken along the arrow AA in Figure 1a. This is a front cross-sectional view of a vibration suppression air pressure control valve applied to the first embodiment. This is a diagram showing a control circuit that drives the control valve of the first embodiment and a feedforward signal input waveform. Figure 4a shows an intake side nozzle part of the control valve of the first embodiment, and Figure 4b is an enlarged view of this, and Figure 4b is a diagram showing the component configuration of the intake side annular nozzle. Figure 5a shows an enlarged view of this, and Figure 5b is an exhaust side annular nozzle part of the control valve of the first embodiment. 6a shows the state where the movable body is blocking the intake nozzle, FIG. 6b shows the state where the movable body is in a position halfway between the intake nozzle and the exhaust nozzle, and FIG. 6c shows the state where the movable body is blocking the exhaust nozzle. Graph comparing the flow rate of a valve equipped with the "annular nozzle with bypass flow path" according to the present invention and a valve with a conventional "single-hole nozzle." Graph 8a shows a model of the positional relationship between the single-hole intake nozzle and the movable body, and shows the input current 0<I<I. 1 When the input current I is in the range of I > I, Figure 8b shows 1 Graph showing the relationship between valve flow rate Q and movable body stiffness K with respect to the single-hole nozzle inner diameter d Graph showing maximum valve flow rate with respect to stiffness increase ratio K / K0 in the case of a single-hole nozzle Fail-safe displacement X for a single-hole nozzle and an annular nozzle with a bypass channel fel This graph shows the relationship between the nozzle inner diameter d and the outer nozzle inner diameter d0 of the annular nozzle with bypass passage, and the relationship between the valve flow rate Q and the flapper rigidity ratio K / K0, and compares it with the characteristics of a conventional single-hole nozzle. This graph compares the maximum flow rate that can be obtained with the annular nozzle of the present invention and a conventional single-hole nozzle, and shows the maximum flow rate Q for the rigidity ratio K / K0. max Graph of the adsorption force F to the exhaust nozzle versus the single-hole nozzle inner diameter d P Graph showing the adsorption force F of the annular nozzle to the exhaust nozzle relative to the outer nozzle inner diameter d0, compared with a conventional single-hole nozzle. PThis graph explains the factors that hindered the achievement of a large flow rate in conventional nozzle flapper type control valves. Figure 16a is a diagram showing the basic structure of an oscillating servo valve, Figure 16b is a model diagram extracting the movable body and nozzle section, and Figure 16c is a diagram in which an annular nozzle is applied to the above control valve. Figure 17a is a front cross-sectional view and Figure 17b is a top view of Figure 17a, which extracts only the electromagnetic actuator section of the pneumatic control valve that is the first embodiment of the present invention. This graph compares the case with a bypass flow path and the case without a bypass flow path, and calculates the effective flow area for the gap h of the annular nozzle. 19a is a diagram of an intake-side annular nozzle with a bypass flow path, and 19b is a diagram of an intake-side annular nozzle without a bypass flow path. 20a is a diagram of an exhaust-side annular nozzle with a bypass flow path, and 20b is a diagram of an exhaust-side annular nozzle without a bypass flow path. 21a is a model diagram with a bypass flow path, and 21b is a model diagram without a bypass flow path. 21b is a model diagram with a bypass flow path. 21a is a model diagram with a bypass flow path, and 21b is a model diagram with a bypass flow path. 21b is a model diagram with a bypass flow path. 21b is a model diagram with a bypass flow path. 21b is a model diagram with a bypass flow path. 21a is a model diagram with a bypass flow path. ... The analysis results for the pressure distribution without a bypass flow path are shown in Figure 22a, which shows the pressure distribution for the entire analysis model, and Figure 22b, which shows an enlarged view of the nozzle section. The analysis results for the velocity vector when there is a bypass flow path are shown in Figure 23a, which shows the velocity vector for the entire analysis model, and Figure 23b, which shows an enlarged view of the nozzle section. The analysis results for the pressure and axial velocity distribution near the intake nozzle opening when there is a bypass flow path are shown in Figure 24a, which shows the analysis model, Figure 24b, which shows the pressure distribution, and Figure 24c, which shows the velocity distribution. The analysis results for the pressure distribution without a bypass flow path are shown in Figure 25a, which shows the pressure distribution for the entire analysis model, and Figure 25b, which shows an enlarged view of the nozzle section. The velocity vector when there is no bypass flow path. The results of the computational fluid analysis to verify the effect of increasing the exhaust flow rate by the through-hole and bypass flow path are shown in Figure 27a, where the pressure distribution of the entire analysis model when there is a bypass flow path is shown in Figure 27b, and the nozzle is an enlarged view. Figure 28a, where the velocity vector of the entire analysis model when there is a bypass flow path is shown in Figure 28b, and the nozzle is an enlarged view. Figure 29a, where the pressure distribution of the entire analysis model when there is no bypass flow path is shown in Figure 29b, and the nozzle is an enlarged view. Figure 30a, where the velocity vector of the entire analysis model when there is no bypass flow path is shown in Figure 27b, and the nozzle is an enlarged view.Figure 30b is an enlarged view of the nozzle portion. A graph showing the flow rate characteristics relative to the gap in the annular nozzle obtained by numerical analysis. Figure 32a shows an air pressure control valve that is a second embodiment of the present invention, with the intake nozzle in Figure 32c being a front cross-sectional view of the entire control valve in cross-sectional view BB. Figure 32b is a front cross-sectional view of the entire control valve in cross-sectional view CC. Figure 32c is a view taken along arrow AA in Figure 32a. Figure showing that, in the second embodiment of the present invention, the annular nozzle with built-in bypass flow path is configured by inserting an inner member into an outer member. Second embodiment of the present invention 34a is a front cross-sectional view of a bypass flow path-integrated annular structure, and FIG. 34b is a DD cross-sectional view of FIG. 34a. FIG. 34b is a front cross-sectional view of an air pressure control valve that is a third embodiment of the present invention. FIG. 37a is a top view of the exhaust side nozzle part, FIG. 37b is a front cross-sectional view, and FIG. 37c is a view taken along the arrow AA of FIG. 37b. FIG. 37b shows the flow of magnetic flux between exhaust side members in this embodiment. FIG. 37b is a front cross-sectional view of an air pressure control valve that is a fourth embodiment of the present invention. FIG. 40a is a model diagram when an adsorption prevention hole is formed, and FIG. 40b is a model diagram when an adsorption prevention hole is not formed. 44a is a view taken along the arrow AA in FIG. 44b, FIG. 44b is a top view of FIG. 44a, and FIG. 44c is a view taken along the arrow BB in FIG. 44b. FIG. 45a is a view taken along the arrow AA in FIG. 44b, FIG. 44b is a top view of FIG. 44a, and FIG. 45c is a view taken along the arrow BB in FIG. 44b. 45a is a front cross-sectional view, and FIG. 45b is a diagram showing the shape of the tip of the movable body. FIG. 46a is a model diagram extracting only the movable body and the annular nozzle, and FIG. 46b is a side view of the annular nozzle portion. In the sixth embodiment of the present invention, a model diagram showing when the annular nozzle is inclined at a small angle to the movable body. In the sixth embodiment of the present invention, FIG. 48a is a diagram showing the tight contact state between the tapered movable body and the annular nozzle, and FIG. 48b is a model diagram extracting only the movable body. In the sixth embodiment of the present invention, the tip of the annular nozzle is tapered, and a model diagram showing the tight contact state between the flapper and the annular nozzle. FIG. 49a and FIG. 49b are air pressure control valves which are the seventh embodiment of the present invention. FIG. 50a is a front cross-sectional view, and FIG. 50b is a view taken along the arrow AA in FIG. 50a.50c is a view showing the shape of the tip of the movable body, which is the opposing surface of the annular nozzle. FIG. 50c is a view showing the shape of the tip of the movable body, which is the opposing surface of the annular nozzle. FIG. 50c is a view showing the shape of the tip of the movable body, which is the opposing surface of the annular nozzle. FIG. 50c is a view showing the shape of the tip of the movable body, which is the opposing surface of the annular nozzle. FIG. 50c is a view showing the shape of the tip of the movable body, which is the opposing surface of the annular nozzle. 54b is a diagram showing the air flow from the supply source to the exhaust port when the CC cross section of FIG. 53b is illustrated. FIG. 54b is a diagram showing the air flow in the bypass flow path when the CC cross section of FIG. 53b is illustrated. FIG. 53 is a front cross section of the structure of FIG. 53 modified into a one-way three-way valve. FIG. 55 is a front cross section of the structure of FIG. 55 when a rectifier is attached. FIG. 58a is a model diagram of Unit (A), which is an actuator unit, and FIG. 58b is a model of a vibration damping valve. 58a and 58b are graphs of the force generated by the actuator over time. FIG. 59 shows a fluid control device according to an eleventh embodiment of the present invention, with FIG. 59a being a model diagram of Unit (A) and FIG. 59b being a model diagram of an anti-vibration servo valve and an anti-vibration actuator. FIG. 60a shows a front cross-sectional view of a pneumatic servo valve (fluid control valve) applied to a twelfth embodiment of the present invention, with FIG. 60b being an enlarged view of part AA in FIG. 60a. A graph of the analysis results of the flow velocity V of the intake nozzle nozzle holes of this embodiment. A graph showing the effect of reducing the Reynolds number of the intake nozzle nozzle holes. 64a is a front cross-sectional view of the pneumatic servo valve of the thirteenth embodiment of the present invention; the annular nozzle composed of an arc-shaped flow path is constructed from a single part; Fig. 64a is a view taken along the arrow AA in Fig. 64b, Fig. 64b is a top view of Fig. 64a, Fig. 64c is a view taken along the arrow BB in Fig. 64b; and Fig. 65a is a front cross-sectional view of this embodiment; Fig. 65b is an enlarged view of the dashed line CC in Fig. 65a; Fig. 65c is a view showing the relative positions and dimensional relationship of the annular flow path, the through hole formed in the flapper, and the exhaust nozzle.Fig. 65d is a diagram showing the arc-shaped flow path of the annular nozzle.Analysis model diagram of this embodimentGraph of the analysis results of the control flow rate characteristics with respect to the gapGraph of the upstream pressure characteristics of the annular nozzle with respect to the gapGraph of the analysis results of the control pressure characteristics with respect to the gapGraph of the analysis results of the nozzle jet hole flow velocity with respect to the gapGraph of the analysis results of the nozzle jet hole Reynolds number with respect to the gapFig. 72a is a front cross-sectional view of a pneumatic servo valve of a 14th embodiment of the present invention, and Fig. 72b is an enlarged view of the DD part of Fig. 72aGraph of the analysis results of the control flow rate characteristics with respect to the gapGraph of the analysis results of the control pressure characteristics with respect to the gapGraph of the analysis results of the upstream pressure characteristics of the annular nozzle with respect to the gapGraph of the analysis results of the radial flow velocity at the outlet of the annular nozzle with respect to the gapGraph of the analysis results of the radial Reynolds number at the outlet of the nozzle with respect to the gap 78a is a front cross-sectional view, and FIG. 78b is an enlarged view of part AA in FIG. 78a. FIG. 79a is a front cross-sectional view of a valve in which the annular nozzle part and the intake side throttle part are arranged at a distance, and FIG. 79b is a structural diagram modeling the annular nozzle and the intake side throttle part. FIG. 80a is an analytical model of a two-way three-way valve with a throttle part, and FIG. 80b is a diagram showing the electrical equivalent circuit of this analytical model. A graph of the analytical results of response. FIG. 82a is a front cross-sectional view of a pneumatic servo valve which is another form of the 15th embodiment, and FIG. 82b is an enlarged view of part CC in FIG. 82a. An oscillating type pneumatic servo valve which is the 16th embodiment of the present invention, and FIG. 83a is a front cross-sectional view, and FIG. 83b is a diagram showing the shape of the flapper tip which is the opposing surface of the annular nozzle. max85a is a front cross-sectional view, and FIG. 85b is a view taken along the arrow BB in FIG. 85a. In Supplementary Note [2-1], FIG. 86a is an overall view of the servo valve, FIG. 86b is an enlarged view of part AA in FIG. 86a, and FIG. 86c is a view showing the application of a single-hole nozzle to the intake-side throttle section. Model diagram of a two-way three-way valve consisting of an intake-side throttle, an annular nozzle (variable nozzle), and an exhaust nozzle. Simplified model diagram in which the exhaust nozzle is omitted and the annular nozzle, which is a variable nozzle, is replaced with a single-hole nozzle. Simplified model diagram identical to FIG. 88 above. Figure showing the analysis results of the flow velocity distribution in a computational fluid analysis model of the simplified model diagram above. Figure showing the analysis results of the pressure distribution. Graph of flow rate characteristics versus gap δ in a ring-shaped viscous pipe. Graph showing pressure drop ΔP versus gap δ in a ring-shaped viscous pipe. Graph showing fluid resistance R versus gap δ in a ring-shaped viscous pipe. Graph showing the characteristics of fluid resistance R versus gap δ in a ring-shaped viscous pipe when the viscous pipe length L is 2 mm. Upstream pressure P of the annular nozzle. a and gap h / h max Graph showing the relationship between nozzle downstream pressure P a Graph showing the characteristics of flow rate Q against pressure. Flow rate change rate C against pressure. d and downstream pressure P aIn Supplementary Note [1-1], a model diagram explaining the compliance effect of an annular nozzle and a thin disk, where Fig. 100a is a top view of the disk, Fig. 100b is a side view when the tip of the valve seat is not inclined relative to the disk that is the opposing surface, and Fig. 100c is a side view when it is inclined at an angle θ. In Supplementary Note [1-2], a model diagram in which each valve seat of the annular nozzle is configured as a square. A model diagram in which the annular nozzle is formed by dividing it discontinuously in the circumferential direction. A model diagram in which the annular flow path and bypass flow path of the annular nozzle are formed with a multilayer structure, where Fig. 103a is a side cross-sectional view and Fig. 103b is a top view. In Supplementary Note [1-3], a diagram showing an example of a control block diagram of an active vibration isolation device. Fig. 104 Fig. 105a shows the gain characteristics versus frequency, and Fig. 105b shows the phase characteristics versus frequency. In Supplementary Note [1-4], a front cross-sectional view of an air pressure control valve in which only the inner member through-passage is the main control flow path. A model diagram showing an active vibration isolation table equipped with a conventional fluid control valve. A diagram of a model of a conventional nozzle flapper type fluid control valve. A front cross-sectional view of a conventional nozzle flapper type fluid control valve. A side cross-sectional view of a conventional nozzle flapper type fluid control valve. A front cross-sectional view of a conventional spool type fluid control valve. A diagram of a conventional proposal for a one-way three-way valve aimed at reducing noise, controlling a micro-movement mechanism. A front cross-sectional view of the above one-way three-way valve.

[0057] 162 Movable body (flapper) 170 Forward direction nozzle portion 171 Reverse direction nozzle portion 176 Control chamber 168 Supply hole (intake hole) 157 Discharge hole (exhaust hole) 301 Inner member 305 Outer member 308 Annular flow path 309 Outer valve seat 303 Inner valve seat

[0058] The present invention will be explained below in two steps: [I] Application to a vibration suppression control valve [II] Application to a vibration isolation control valve First, the embodiment [I] above will be explained.

[0059] First Embodiment FIG. 1 is a model diagram of a fluid control system (active vibration isolation table) equipped with feedforward vibration suppression control (FF control) according to a first embodiment of the present invention. FIG. 1a is a cross-sectional view taken along the line BB in FIG. 1b, and FIG. 1b is a view taken along the line AA in FIG. 1a. [1] Configuration of the Active Vibration Isolation Table: A plurality of pneumatic actuators are arranged on a floor 200 (foundation) to support a base 201 (vibration isolation table). A precision device (not shown) is mounted on this base 201. Reference numeral 202 denotes a stage mounted on the base 201. In this embodiment, the acceleration α during acceleration and deceleration of the stage and the reaction force of the inertial force F=mα, determined by the stage mass m, act as a vibration source that vibrates the base. Information about the excitation force applied to the base by this vibration source, i.e., information about the excitation force applied to the base as the stage moves, is input to a feedforward control controller (not shown). Reference numerals 203a and 203b denote pneumatic actuators (vibration isolation means) that support the base in the vertical Z-axis direction. These pneumatic actuators suppress vibration transmission between the floor and the surface plate, and also serve as a means for supporting the surface plate against the floor. The configuration of multiple sets of actuators supporting the horizontal X and Y directions is called a "torque balance type," and is modeled in the same figure. 204 denotes the floor-side outer frame, and 205 denotes the surface plate-side inner frame. Four sets of vibration-damping pneumatic actuators 206a, 206b, 206c, and 206d, vibration-damping air servo valves 207a, 207b, 207c, and 207d, vibration-isolating pneumatic actuators (vibration-isolating means) 208a, 208b, 208c, and 208d, and vibration-isolating air servo valves 209a, 209b, 209c, and 209d are arranged axially symmetrically at positions away from the X and Y axes. [2] Structure of the vibration suppression servo valve [2-1] Basic valve structure Figure 2 shows the vibration suppression pneumatic servo valve (fluid control valve) (207d shown by the chain line BB in Figure 1b) applied to the first embodiment of the present invention, and Figure 2 is a front cross-sectional view of it, and Figure 3 shows the control circuit that drives the servo valve and the FF signal (feedforward signal) input waveform. Figure 4 is an enlarged view of the intake side (supply side) nozzle part, and Figure 5 is an enlarged view of the exhaust side (discharge side) nozzle part.The servo valve, except for the intake and exhaust nozzles, is currently being proposed by the present inventors (see Patent Document 5), and its basic operating principle is a nozzle-flapper type. Reference numeral 150 denotes a cylindrical central shaft (support shaft), 151 denotes the bottom of the central shaft, 152 denotes an outer frame of the central shaft formed concentrically with the central shaft, 153 denotes a coil bobbin attached to the central shaft, and 154 denotes a coil wound around the coil bobbin. The central shaft 150, the bottom 151 of the central shaft, the outer frame 152 of the central shaft, the coil bobbin 153, and the coil 154 constitute an electromagnetic actuator that attracts a flapper (described below) and controls its displacement. Reference numeral 155 denotes a cylindrical exhaust-side housing that houses the bottom 151 and outer frame 152 of the central shaft, 156 denotes the bottom of the exhaust-side housing, 157 denotes an exhaust hole (opening) formed in the housing bottom 156, and 158 denotes an exhaust-side flow passage (through-flow passage) formed in the central shaft 150. Reference numeral 159 denotes an intake-side housing, 160 denotes an intake-side flow passage B formed in the intake-side housing, and 161 denotes a control-side flow passage A (control port) connected to a pneumatic actuator (described below). Reference numeral 162 denotes a disk-shaped flapper (movable body), the outer periphery of which is secured between the intake-side housing 159 and the exhaust-side housing 155 by a positioning pin 163. One end of the exhaust-side flow passage connects to the exhaust port of the control valve body, and the flapper side of the exhaust-side flow passage is equipped with an annular nozzle (described below). In the case of a two-way, three-way valve, the control valve body refers to the entire valve, including the flow passages connecting the three openings (intake port, control port, and exhaust port) and the actuator. The flapper 162 is composed of a spiral-shaped groove 162a and a ridge 162b (not shown). The term "flapper" derives from the idea of ​​a baby bird flapping its wings, and evokes an image of a swinging motion. However, in the present invention, the term "flapper" is not limited to swinging motion, but also includes a disk-shaped member (valve element) that moves linearly, as in this embodiment. Reference numeral 164 denotes a base member fastened to the intake side housing 159 by bolts 166. Reference numeral 167 denotes an intake side flow path A formed inside this base member, 168 denotes an intake hole, 169 denotes a control side flow path B, and 169d denotes a control port.Reference numeral 170 denotes an intake-side nozzle section (first nozzle section), which is a forward-direction nozzle attached to the center of the intake-side housing 159. As described below, this is an annular nozzle formed by combining two parts. Reference numeral 171 denotes an exhaust-side nozzle section (second nozzle section), which is a reverse-direction nozzle attached to the flapper side of the central shaft 150. Like the intake-side nozzle section, this is also an annular nozzle formed by combining two parts. Reference numeral 172 denotes an intake-side gap formed between the intake-side housing 159 and the flapper 162, and reference numeral 173 denotes an exhaust-side gap formed between the flapper 162 and the exhaust-side housing. Reference numeral 174 denotes the flapper-side end face (center shaft end face) of the central shaft 150, which is a first magnetic pole of an electromagnet. Reference numeral 175 denotes a second magnetic pole formed on the flapper-side end face of the outer frame section 152. The intake-side gap and the exhaust-side gap are connected by the spiral groove formed in the flapper 162, and these two gaps form the control chamber 176 of this valve. Reference numeral 177 denotes a bypass through-hole formed in the center of the flapper 162. Reference numeral 178d, indicated by an imaginary line, denotes a control-side flow path C formed in the floor-side outer frame 204 (FIG. 1) and connected to the air chamber of the vibration-damping pneumatic actuator 206d (described later in FIG. 3). In this embodiment, the flow path connecting the vibration-damping pneumatic servo valve 207d and the vibration-damping pneumatic actuator 206d is formed using the shortest linear flow path shape 178d. This configuration, which minimizes piping flow path resistance, fully demonstrates the effects of the present valve's high flow rate and high responsiveness. Reference numeral 179 denotes an outer sleeve that houses the outer peripheries of the intake-side housing 159 and the exhaust-side housing 155. To summarize the basic operating principle of the nozzle-flapper type servo valve described above, the intake nozzle and exhaust nozzle are arranged opposite the flapper, and the flapper is actuated by the attractive force due to Maxwell stress generated between the magnetic pole of the electromagnet and the flapper, thereby varying the length of the flow path formed between the flapper and the nozzle, thereby controlling the fluid pressure or flow rate in the control chamber. This servo valve is a two-way (bidirectional) three-way valve whose flow path consists of two-way nozzles (forward nozzle 170, reverse nozzle 171) and three openings (intake port 168, exhaust port 157, and control port 169d).[2-2] Control circuit for driving the valve Figure 3 shows the control circuit for driving the valve and the FF signal (feedforward signal) input waveform when the above servo valve is applied to vibration suppression control. 250 is the control circuit for driving the servo valve 207d. 251 indicates the feedforward signal (FF signal) F0(t) for offsetting the impact load caused by stage movement. In other words, to prevent excitation of the vibration isolation table caused by repeated acceleration and deceleration of the stage on the vibration isolation table, the target value of the input waveform is a force (counterforce) with the same magnitude but opposite phase as the excitation force. The FF signal is obtained from a command signal to the stage drive mechanism or from actual measurements by acceleration sensors installed on each control axis. [2-3] Structure of the Intake and Exhaust Nozzle Section (i) Intake-Side Nozzle Section Figure 4a shows an enlarged view of the intake-side nozzle section 170 (first nozzle section) in the servo valve of the first embodiment (Figure 2). Reference numeral 301 denotes an inner member, 302 denotes a bypass flow passage (detour flow passage), which is a central through-passage formed in the center of the inner member. The effect of the bypass flow passage will be described in detail later. Reference numeral 303 denotes an inner valve seat formed on the flapper side of the inner member, and 304 denotes a nozzle opening formed at the flapper-side open end of the bypass flow passage. Reference numeral 305 denotes an outer member, which is housed between the intake-side housing 159 and the inner member 301. Reference numeral 306 denotes an intake-side flow passage C formed radially within the outer member, 307 denotes a circumferential groove formed on the inner surface of the outer member, and 308 denotes an annular flow passage formed by the inner surface of the outer member 305 and the outer surface of the inner member 301. That is, high-pressure air flowing in through the intake hole 168 of the base member 164 flows through the following order: intake-side passage A 167 → intake-side passage B 160 → intake-side passage C 306 → circumferential groove 307 → annular passage 308, and then flows out into the intake-side gap 172. Reference numeral 309 denotes an outer valve seat formed on the flapper side of the outer member, and this outer valve seat and inner valve seat 303 form an intake-side annular nozzle. The bypass passage 302, which is a central through-passage, communicates with the control-side passage B 169 formed in the center of the base member 164. The upstream side of this control-side passage B 169 communicates with the control-side passage A 161 formed in the intake-side housing 159.Specifically, the fluid flowing from the intake-side annular nozzle toward the outer valve seat 309 (centrifugal direction) flows into the control chamber 176 and then into the control-side flow path A 161. The fluid flowing toward the inner valve seat 303 (centripetal direction) flows into the control-side flow path B 169 via the bypass flow path 302. In other words, the two fluids branching out from the annular nozzle merge again in the control-side flow path B 169. As shown in Figure 4a, 309c denotes the tip of the outer valve seat 309, and 303c denotes the tip of the inner valve seat 303. These tips are provided with a seal having a width of 0.1 to 0.3 mm. The flapper and the outer valve seat tip 309c and inner valve seat tip 303c are in close contact with each other to block the annular flow paths. Figure 4b shows that the intake-side annular nozzle is composed of two components, an inner member 301 and an outer member 305, and is housed in the intake-side housing 159. (ii) Exhaust-Side Nozzle Section Figure 5a shows an enlarged view of the exhaust-side nozzle section 171 (second nozzle section), with 310 as an inner member, 311 as an inner valve seat formed on the flapper side of the inner member, 312 as a conical gap, 313 as a cylindrical section formed on the opposite side of the inner member from the flapper side, and 314a, 314b, 314c, and 314d as cylindrical exhaust flow passages formed axially through the cylindrical section. The cylindrical section 313, which has an axial through-passage, serves as a flow passage converting section connecting the annular flow passage and the cylindrical flow passage (exhaust-side flow passage 158). 315 as an outer member, which is provided on the outer periphery of the inner member. 316 as an outer valve seat formed at the tip of the outer member on the flapper side. 317 as an annular flow passage formed by the inner circumferential surface of the outer member and the outer circumferential surface of the inner member. In other words, the outer valve seat and the inner valve seat together constitute an exhaust-side annular nozzle. The outer member and the inner member are press-fitted onto the cylindrical central shaft 150. 316c denotes the tip of the outer valve seat 316, and 311c denotes the tip of the inner valve seat 311, with a seal portion 0.1 to 0.3 mm wide formed at these tips. The flapper and outer valve seat tip 316c and inner valve seat tip 311c are in close contact with each other to block the exhaust-side annular flow passage 317. Figure 5b shows that the exhaust-side annular nozzle is composed of two parts: the inner member 310 and the outer member 315.To summarize the configuration of the exhaust-side annular nozzle, the flow path formed in the exhaust-side flow passage (through-flow path) 158 is composed of the annular flow path formed by the inner member and the outer member, a flow path conversion section (cylindrical section 313), and the cylindrical through-flow path. By arranging the annular flow path and the flow path conversion section in series, the inner diameter of the through-flow path can be reduced, thereby reducing the outer diameter of the central shaft. Reducing the outer diameter of the central shaft allows for a reduction in the outer diameter of the coil bobbin 154 arranged around this central shaft, thereby enabling the magnetic attraction type actuator and the servo valve body to be made smaller. Both the intake-side nozzle (i) and the exhaust-side nozzle (ii) above are annular nozzles that adjust the opening state between the annular flow path and the flapper by moving the flapper closer to or farther from the outer valve seat tip (309c, 303c) and the inner valve seat tip (303c, 311c). By applying these annular nozzles, the valve of this embodiment achieves significantly increased intake and exhaust flow rates compared to conventional servo valves consisting of single-hole valves. [2-4] Valve Operation of This Embodiment Figure 6 shows the movement of the flapper (movable body) and the air flow within the valve during the rise (intake stroke) from when power is applied to the servo valve of this embodiment until air is filled into the pneumatic actuator (not shown). The fall (exhaust stroke) during which the air in the pneumatic actuator is exhausted will be described in detail later. Figure 6a shows the state when the current applied to the valve is 0 and the flapper is blocking the intake (supply) nozzle. Figure 6b shows the state when the flapper is positioned midway between the intake and exhaust (discharge) nozzles. Figure 6c shows the state when the flapper is blocking the exhaust nozzle. In Figure 6a, the intake port 168 is connected to a high-pressure source P. SThe pressure in the control chamber 176 and the pneumatic actuator (not shown) connected to this control chamber is maintained at atmospheric pressure. The flapper is displaced due to its elastic deformation (fail-safe), and the annular flow path 308 is blocked by the flapper coming into close contact with the outer valve seat tip 309c and the inner valve seat tip 303c. The nozzle opening 304, which is the open end of the bypass flow path 302 on the flapper side, is connected to the bypass through-hole 177 and the bypass flow path 302, and therefore the pressure in the control chamber 176 is maintained at atmospheric pressure. Therefore, the fail-safe force F required to block the annular flow path is fel is the opening cross-sectional area of ​​the annular flow path, A r If the pressure difference between the supply pressure and the control room is ΔP, then F fel =A r×ΔP is sufficient. In Figure 6b, the fluid flowing from the annular passage toward the flapper side branches into the centrifugal direction and the centripetal direction. Furthermore, the fluid flowing in the centripetal direction branches into the bypass passage and the through-hole. The bypass passage acts to branch the fluid flow, significantly increasing the intake flow rate during the intake stroke. As will be described in detail later, the combination of the bypass through-hole and the bypass passage also significantly increases the exhaust flow rate during the exhaust stroke. This effect of increasing both the intake flow rate and the exhaust flow rate was discovered in this study and will be referred to hereafter as the "bypass effect." In Figure 6c, the exhaust-side annular passage 317 is blocked by the flapper and the outer valve seat tip 316c and inner valve seat tip 311c of the exhaust-side annular nozzle coming into close contact or proximity. When the exhaust flow rate is blocked, it is preferable for the flapper and the valve seat tip 316c, 311c to be in close contact. However, complete contact is not required because leakage flow is significantly reduced when the gap between the components is in the viscous fluid range (10 μm or less). In the servo valve of this embodiment, annular nozzles are used for both the forward and reverse nozzles. The basic performance required of a servo valve is to shorten the time it takes to fill the actuator with air while also shortening the time it takes to exhaust the air. To achieve this, it is necessary to increase the maximum intake flow rate and the maximum exhaust flow rate. Therefore, when a pneumatic servo system is constructed using a servo valve of the present invention in which annular nozzles are used for both the forward and reverse nozzles, the rise time constant and fall time constant can be set to the same level. In this embodiment, the exhaust-side flow passage is formed through the central axis 150, and a coil bobbin is attached around this central axis 150 to form an electromagnetic actuator. An exhaust nozzle is installed on the flapper-side end of the exhaust-side flow passage, and an intake nozzle is installed across the flapper. Common to all aspects of the present invention, the nozzle flapper valve functions even if the exhaust flow path and the intake flow path are reversed (not shown).[2-5] Effects of the Present Invention (Summary) ... Comparison of Maximum Flow Rates Between a Valve of the Present Invention and a Conventional Valve The graph in Figure 7 shows an example of a comparison of the maximum flow rates between a valve equipped with the "annular nozzle with bypass flow path" of the present invention and a valve using a conventional "single-hole nozzle." In the figure, 401 is the single-hole nozzle, 402 is the valve seat tip of this single-hole nozzle, and 403 is the flapper (movable body) when using the single-hole nozzle. Here, the inner diameter of the outer valve seat tip (309c in Figure 4a) is the outer nozzle inner diameter d0, and the outer diameter of the inner valve seat tip (303c in Figure 4a) is the inner nozzle outer diameter d. i That is, in the case of a single-hole nozzle, it shows the relationship between the nozzle inner diameter d at the valve seat tip 402 and the flow rate Q, and in the case of an annular nozzle, it shows the relationship between the outer nozzle inner diameter d0 and the flow rate Q. As shown in the figure, the analysis conditions are the maximum displacement of the flapper X0 = 180 μm, the spring stiffness of the disk (flapper) K0 = 43.9 N / mm, and the intake side pressure P S In all the annular nozzles covered by this embodiment, the inner diameter d0 of the outer nozzle and the outer diameter d i The difference Δr = d0-d i = 0.72 mm. The force required for the actuator of the servo valve to generate is the force F = K0X0 required to drive the flapper over its full stroke. Therefore, the design specifications of the valve of the present invention and the conventional valve are the same except for the nozzle section. In the same figure, in the case of the conventional single-hole nozzle, the valve flow rate Q increases in the range of the nozzle inner diameter d < 2.8 mm, and reaches a maximum value Q at d = 2.5 mm. max= 50 L / min. However, it decreases in the range of d > 2.5 mm, reaching Q = 0 at d = 4.5 mm. The reason for the flow rate suppression as the nozzle inner diameter increases will be discussed in detail below. In the case of the annular nozzle of the present invention, the flow rate Q increases dramatically as the outer nozzle inner diameter d0 increases. The flow rate suppression observed with conventional single-hole nozzles does not occur when the outer nozzle inner diameter d0 < 7.5 mm. At d = d0 = 2.5 mm, where flow rate suppression begins with the conventional single-hole nozzle, the valve of the present invention achieves a flow rate of Q = 108 L / min. Furthermore, at d0 = 7.5 mm, where flow rate suppression begins, Q = 210 L / min. In the case of the servo valve to which the present invention is applied, the outer nozzle inner diameter d0 has a significantly smaller effect on the restriction on flow rate increase compared to the conventional single-hole nozzle. Therefore, while maintaining the small actuator configuration of the conventional valve, which had a maximum flow rate of 35 to 50 L / min, the valve of the present invention achieves a flow rate more than four times that of the conventional valve. [3] Consideration of Obstacles to High Flow Rates and the Benefits of the Invention The previous section presented an example comparing the maximum flow rates of the valve of the present invention and a conventional valve. As mentioned above, the required force for the actuator of the two servo valves is the same, and the design specifications are the same except for the nozzle. The actuator specifications were based on a low-flow valve with a control flow rate of approximately 15 to 50 L / min, which is required for vibration isolation control. However, the valve of the present invention achieved a flow rate of over 200 L / min, more than four times higher. This result completely overturns the conventional wisdom that achieving high flow rates in a nozzle flapper valve requires significant modifications to the electrical and magnetic circuits and mechanical components that make up the entire valve. As mentioned above, this research was initiated by a new challenge that emerged in the vibration control technology of active vibration isolation tables due to trends in the semiconductor-related industry. In other words, to achieve high acceleration and vibration isolation control, the actuator needs to generate a large force that can reproduce the target waveform of feedforward control with an accuracy of several milliseconds or even 0.1 milliseconds. The key to achieving this is the realization of a new servo valve that combines high-speed response and high flow rate. Therefore, the inventors of the present invention traced back to the basic operating principle of a servo valve, which is composed of a valve body (flapper) and a valve seat (nozzle), and attempted to identify the factors that hindered the achievement of a large flow rate.The present invention was fortunately discovered through a theoretical analysis of these factors (1) through (3) and a search for solutions. (1) Reduction in the effective flapper displacement due to the fail-safe mechanism; (2) Adsorption of the flapper to the exhaust nozzle; and (3) Constraints imposed by the servo valve's basic operating principles. The optimal configuration of the servo valve of the present invention cannot be achieved by solving only one of these factors; all of them must be satisfied. (1) through (3) are explained below in order. [3-1-1] Factors Impeding High Flow Rates (1) Reduction in the Effective Flapper Displacement Due to the Fail-Safe Feature. One of the factors impeding the achievement of high flow rates in conventional servo valves was the fail-safe mechanism, a necessary requirement for nozzle-flapper valves. Below, we consider the impact of the fail-safe mechanism on valve flow rate, comparing it with the present invention. The fail-safe mechanism is an essential safety feature that closes the servo valve's intake port to prevent high-pressure gas from entering the actuator when the input current I applied to the servo valve is zero, or during an emergency such as a power outage. (1) Fail-safe model diagram Figure 8 shows a model of the positional relationship between a single-hole intake nozzle 401 and a flapper 403 (movable body). 404 is an exhaust nozzle arranged opposite the intake nozzle. Figure 8a shows the input current 0<I<I. 1 8b shows the state in which the flapper 403 covers the open end of the intake nozzle 401 in the range of I>I 1 This shows the state in which the flapper 403 is separated from the open end of the intake nozzle 401 within the range of I=0, allowing air to flow into the control chamber 405. When the input current I=0, it is an essential requirement to block the intake side flow path connected to the high pressure source, regardless of the type of servo valve. For this reason, it is necessary to apply a preload to the flapper at the initial setting stage to block the intake side opening. Let K be the spring stiffness of the flapper, and X be the required flapper displacement. fel When this happens, the preload (restoring force) F given to the flapper is fel =KX fel is the fail-safe force. Here, the intake pressure is P S , exhaust pressure is P0 (atmospheric pressure), the intake nozzle diameter is d, the opening area of ​​the intake nozzle is A S (=d 2If the intake nozzle diameter d is increased to increase the maximum flow rate of the valve, the intake nozzle area A S and the force F multiplied by the pressure difference S [=A S (P S - P0) is applied to the flapper. Fail-safe force F fel >F S Therefore, if the intake nozzle diameter is increased, the fail-safe force must be increased. (2) Relationship between the valve flow rate and the nozzle inner diameter Figure 9 is a graph showing the relationship between the valve flow rate Q and the flapper stiffness K with respect to the single-hole nozzle inner diameter d. The maximum displacement (stroke) of the flapper 403 is X max The figure shows the characteristics of the valve flow rate Q versus the nozzle inner diameter d when the flapper stiffness ratio is changed in the range of 1≦K / K0≦10, with the flapper stiffness ratio K / K0=180 μm and the flapper stiffness K0=43.9 N / m. When the flapper stiffness ratio K / K0=1 (K=K0), as mentioned above, the valve flow rate Q increases in the range of the nozzle inner diameter d<2.5 mm, and reaches a maximum value Q at d=2.5 mm. max = 50 L / min, and decreases in the range of d > 2.8 mm, and becomes Q = 0 at d = 4.5 mm. The reason is that the valve flow rate Q is eff (=πdX eff ) and as mentioned above, increasing the intake nozzle diameter increases the fail-safe displacement X fel increases, and the effective displacement X that the flapper can move eff =X max -X fel In the figure, when the flapper stiffness ratio K / K0 is increased, the maximum flow rate Q max The reason for this is that as the flapper stiffness K increases, the fail-safe displacement X fel In order to reduce the effective displacement of the flapper, eff =X max -X fel However, if the spring stiffness K is increased, the flapper stroke X max Under the same conditions, the generated force F (=KX max) must also be increased. (3) Relationship between Maximum Flow Rate and Stiffness Increase Ratio Figure 10 is a graph showing the maximum valve flow rate versus the stiffness increase ratio K / K0 for a single-hole nozzle. The stiffness increase ratio K / K0 can be considered the increase ratio of the electromagnet generated force F=KX0. As mentioned above, increasing the electromagnet generated force requires significant modifications to the electrical circuits, magnetic circuits, and mechanical components that make up the entire valve, such as reducing responsiveness due to lower electrical and mechanical time constants and increasing the capacity of the power supply. However, as can be seen from the same figure, increasing the electromagnet generated force only increases the flow rate slowly. For example, increasing the electromagnet generated force tenfold only increases the flow rate by about three times, so the resulting effect is small. [3-1-2] Effect of the present invention on obstruction factor (1) (1) Comparison of fail-safe displacement between the present invention and a conventional valve Figure 11 shows a graph comparing the following characteristics of a "conventional single-hole nozzle" and an "annular nozzle with bypass flow path." However, the maximum displacement X max (=X0), the flapper stiffness K0 is the same condition. (i) In the case of a conventional single-hole nozzle...Fail-safe displacement X fel and the nozzle inner diameter d (ii) In the case of the annular nozzle of the present invention...Fail-safe displacement X fel and the outer nozzle inner diameter d0 In the case of the conventional single-hole nozzle 401 described above (i), when the nozzle inner diameter d=4.5 mm, X eff = 180 μm, so the effective flow area of ​​the nozzle A eff = 0. Therefore, the range of the nozzle inner diameter d that can be used as a valve is limited to a range of d < 4.5 mm. In the case of the annular nozzle 170 with bypass flow path described above in (ii), the fail-safe displacement X fel For example, when d = d0 = 3.0 mm, the application of this invention reduces the fail-safe displacement X fel =80 → 34 μm. Even when the outer nozzle inner diameter d0 = 8 mm, the fail-safe displacement X fel= 100 μm, and the stroke has a margin of 180 - 100 = 80 μm. (2) Comparison of valve flow rate characteristics with nozzle inner diameter Figure 12 is a graph showing the relationship between the valve flow rate Q and the flapper stiffness ratio K / K0 with respect to the outer nozzle inner diameter d0 of the annular nozzle of the present invention, and is compared with that of a conventional single-hole nozzle. In the case of an annular nozzle with a bypass flow path, if the disc spring stiffness K / K0 is changed from 1 to 2, the valve flow rate Q increases significantly in the range of the outer nozzle inner diameter 1 ≦ d0 ≦ 8 mm. For example, if the outer nozzle inner diameter d0 is set to 8 mm and the spring stiffness K / K0 is set to 2, the maximum flow rate Q max = 330 L / min. As mentioned above, increasing the spring stiffness K increases the force F (= KX max ) must also be increased. For example, the maximum current I max Assume that the maximum current I is 100mA. max If the power supply is modified to accommodate currents of 100 to 200 mA, the force F generated by the electromagnet is proportional to the current, making it possible to drive the large flow valve (330 L / min) mentioned above. On the other hand, in the case of a conventional single-hole nozzle, if the disc spring stiffness ratio K / K0 is changed from 1 to 5, the valve flow rate Q reaches its maximum value at the nozzle inner diameter d = 5.5 mm. Therefore, the maximum current I max Even when driven at 500mA, the maximum flow rate Q max Fig. 13 compares the maximum flow rate that can be obtained with the annular nozzle of the present invention and the conventional single-hole nozzle. Based on the graph in Fig. 12, the flow rate Q versus the stiffness ratio K / K0 is max As mentioned above, in the case of a conventional single-hole nozzle, the maximum flow rate Q max The limit of the annular nozzle is Q = 115 L / min. max=400L / min or more can be achieved. [3-2-1] Factors that hinder the achievement of high flow rates (2) The phenomenon of flapper sticking to the exhaust nozzle The second factor that hinders the achievement of high flow rates in conventional nozzle-flapper type servo valves is the phenomenon of flapper sticking to the exhaust nozzle when the exhaust nozzle diameter is increased. If the maximum intake flow rate is increased to shorten the time it takes to fill the air into the pneumatic actuator, the maximum exhaust flow rate must also be increased to shorten the exhaust time. In other words, the intake nozzle diameter and the exhaust nozzle diameter must be equally large. Usually, the intake nozzle diameter and the exhaust nozzle diameter are set to approximately the same dimensions. Figure 14 shows the relationship between the adhesion force F to the exhaust nozzle 404 and the inner diameter d of the single-hole nozzle. P 4 is a graph showing the restoring force F when the flapper 403 is displaced to the maximum and comes into close contact with the exhaust nozzle 404. K0 (=K0X0). Adsorption force F P Graph of and restoring force F K0 The intersection point of the lines is B (d=4.52mm). When d>4.52mm, the attraction force F P >Restoring force F K0 Therefore, the flapper 403 remains attached to the exhaust nozzle 404, and the servo valve becomes uncontrollable. max = 180 μm, and flapper rigidity K0 = 43.9 N / m, the exhaust flow rate can be increased only when the exhaust nozzle diameter is less than 4.52 mm. [3-2-2] Effects of the present invention on obstruction factor (2) Figure 15 shows the relationship between the suction force F to the exhaust nozzle 171 and the outer nozzle inner diameter d0 when an annular nozzle is used as the exhaust nozzle, in comparison with the graph of the conventional single-hole nozzle (Figure 14) mentioned above. P In the same graph, the restoring force F = F when the rigidity of the flapper is changed is shown. K0 , 2F K0 , 3F K0 In the case of the conventional single-hole nozzle 404, increasing the rigidity of the flapper 403 increases the limit value of the exhaust nozzle diameter d at which the exhaust flow rate can be increased. For example, the restoring force F=2F K0 In this case, the limit value of the exhaust nozzle diameter is d<6.3 mm, F=3F K0In this case, d<7.8 mm. In other words, in order to increase the exhaust flow rate, the force generated by the electromagnet must be increased, which leads to a deterioration in response due to a decrease in the electrical and mechanical time constant, and an increase in the capacity of the power supply, and so on. This is the same as in the case of the intake nozzle, and requires modifications to the electrical circuit, magnetic circuit, and mechanical parts that make up the entire valve. In the case of the exhaust nozzle 171 that uses an annular nozzle, the attraction force F Pis much smaller than that of conventional single-hole nozzles. Within the range of 1 < d < 8.0 mm, the flapper does not stick to the exhaust nozzle, allowing for ample expansion of the exhaust flow rate. [3-3-1] Factors Impeding Higher Flow Rates (3) Constraints Due to the Basic Operating Principle of Servo Valves The third factor impeding the achievement of higher flow rates in conventional nozzle-flapper servo valves relates to the basic operating principle of servo valves. First, we re-evaluate the requirements for pneumatic servo valves. (1) Dynamic Characteristics Required for the Moving Parts of a Servo Valve. A valve typically consists of a "disk" and a "valve seat." The disc (disk, flapper) is a moving part that controls the fluid and closes against the seat when the valve is closed. The valve seat (nozzle) is the part that receives the disc when the valve is in the closed position. The crucial difference between the widely used solenoid valve (ON / OFF valve) and the servo valve, the subject of this invention, is the specifications required for the moving part, the "disk." As mentioned above, the requirements for applying a pneumatic servovalve to an active vibration isolation table are as follows: (i) High-speed response—a rise time of several milliseconds or less; (ii) The primary resonance point of the moving part (valve disc) must be sufficiently high, at least several hundred Hz (e.g., 200 Hz) (details in Appendix 3); (iii) Linearity—a linear relationship between the valve current and the flow rate and generated pressure. All of the above (i) through (iii) are governed by the dynamic characteristics of the moving part, the valve disc (flapper). These dynamic characteristics are determined by the mass and shape of the disc, the spring stiffness supporting the disc, and the operating principle of the actuator that drives the disc. (2) Basic Structure of an Oscillating Pneumatic Servovalve Figure 16a shows an example of the basic structure of an oscillating pneumatic servovalve, which is currently mainstream. The basic form of the oscillating pneumatic servovalve is a technology derived from electrohydraulic servovalves, and most servovalves currently used in a wide range of industrial fields are of the oscillating type. Looking back at the history of the development of electrohydraulic servo valves, Hanley Vickers invented the pilot-operated pressure valve as a hydraulic control device in 1936. After that, during World War II, hydraulic servo technology made rapid advances in the United States to improve the speed and precision of hydraulic equipment.Oscillating pneumatic servo valves are electrohydraulic servo valves modified for pneumatic use. Compared to the wide range of applications of hydraulic servos, pneumatic servos were initially a minor player. In recent years, the advent of active vibration isolation tables has provided a new need for pneumatic servos. The adoption of the basic structure of conventional servo valves, developed through hydraulic servo technology, was historically inevitable. Figure 16b shows a model diagram of the moving part of the servo valve, the flapper 554 (disk) and the nozzles 560 and 561 (valve seats) located on the fixed side. As shown in Figure 16b, the flapper 554 oscillates around its center 558 as a fulcrum. Single-hole forward nozzle 560 and reverse nozzle 561 are located opposite the flapper tip 562. The requirements for the flapper (disk) are summarized as follows: its mass should be on the order of a few grams, and its width should be approximately 2–3 mm to minimize the moment of inertia J around the center of rotation. To achieve the aforementioned requirements (i) and (ii) for a pneumatic servovalve, the moment of inertia J of the moving parts around the center 558 must be minimized. To achieve this, the distance L from the center 558 to the tip of the flapper 554 must be shortened, and the mass must be reduced toward the tip. Therefore, given the need to prioritize the dynamic characteristics of the servovalve, as long as developers were focused on oscillating servovalves, the width of the valve disc (flapper) was limited to the same level as the width of the nozzle serving as the valve seat. In this case, the only option was a single-hole nozzle. The aforementioned structural constraints of the oscillating pneumatic servovalve, which has a long history of development and remains mainstream today, were implicitly shared among developers, likely resulting in a psychological factor (cognitive bias) that inhibited the development of new ideas. (3) Applying an annular nozzle to an oscillating pneumatic servovalve. Figure 16c illustrates the application of the annular nozzle employed in this embodiment to an oscillating servovalve. To achieve a larger flow rate, we will assume that an annular nozzle with an outer nozzle inner diameter d0 = 5.0 mm is applied to an oscillating servo valve. In this case, the shape of the tip of the flapper is shown in the figure. When comparing the responsiveness, resonant frequency, required machining and assembly precision, etc. with the embodiment of the present invention (Figure 2), it is clear that there are disadvantages.[3-3-2] Effect of the Present Invention on Impediment (3)...Servo Valve Basics (1) The Movable Part Moves Parallel to the Axis of the Nozzle Part. Figure 17 shows only the electromagnetic actuator part of the pneumatic servo valve according to the first embodiment of the present invention. Figure 17a is a front cross-sectional view, and Figure 17b is a top view of Figure 17a. The magnetic attraction actuator is composed of a first magnetic pole 174, a second magnetic pole 175, and a flapper 162 made of magnetic material. The disk-shaped flapper is composed of a valve body central portion 162c (shown as a chain circle) that faces the tip of the valve seat of the annular nozzle, and an elastically deforming portion 162ab on the outer periphery of the valve body central portion. This elastically deforming portion is formed with a spiral-shaped groove 162a and a ridge 162b. In other words, the flapper is composed of a low-rigidity portion on the periphery that is easily elastically deformed and a high-rigidity portion (valve body central portion 162c) that is resistant to elastic deformation in the center. When a current is applied to the coil (154 in FIG. 2), the flapper 162 is attracted in the direction of the first magnetic pole and the second magnetic pole, as shown by the arrows in the figure. Note that the valve body central portion 162c, where the groove portion 162a is not formed, moves parallel to the axis of the annular nozzle without deformation. In other words, the approximate circle of the valve body central portion 162c is defined as ΦD. r (radius r), and the outer diameter including the sealing portion of the outer valve seat tip 316c (FIG. 5a) in the exhaust side annular nozzle portion is Φd 0d As, ΦD r >Φd 0d ... (1) The groove of the flapper should be formed so that the above formula is satisfied. The width of the flapper of the swing-type servo valve is about 2 to 3 mm, whereas the width of the flapper of the valve of this embodiment is ΦD r= 8.8 mm, which can accommodate annular nozzles with sufficiently large outer diameters (outer nozzle inner diameters). In other words, this embodiment of the valve, which uses the magnetic attraction force of an electromagnet to move the flapper in parallel to vary the length of the flow path formed between the flapper and the nozzle, is not restricted by the outer diameter of the annular nozzle. As shown in the graph of valve flow rate Q vs. outer nozzle inner diameter d0 (Figure 7), this invention easily realizes a high-flow valve with Q exceeding 200 L / min using a simple structure comparable to the low-flow valves (35-50 L / min) used for vibration isolation control. (2) The annular nozzle valve seat tip is in close contact with the disk surface...compliance effect. The servo valve of this embodiment, which uses a thin disk as the flapper (moving part) and magnetically attracts this disk, has been found to have excellent sealing performance for shielding the annular nozzle valve seat tip. This sealing performance is effective when using an annular nozzle with a large outer diameter to increase flow rates. Here, the axis of the valve body central portion 162c is defined as the Z axis, and the coordinates of the X axis and the Y axis are defined. The reason why the valve body central portion 162c has excellent sealing performance is that it not only moves parallel to the Z axis direction, but also moves in the X axis direction. θ Direction, Y θThis is because the valve deforms with high compliance (flexibility) in both directions. Compliance is the inverse of spring stiffness and is a physical quantity that quantifies the elasticity and ductility of an object. It is also a term used to describe structures that are not rigidly fixed like cantilevers but allow a certain degree of deformation and displacement in response to external forces. Even when the tip of the valve seat of the annular nozzle is slightly tilted relative to the disk surface perpendicular to the axis due to errors in machining or assembly accuracy, the elastically deforming portion 162ab, where the groove 162a is formed, deforms three-dimensionally, resulting in the valve body center portion 162c, which faces the annular nozzle, adhering to the shape of the tip of the valve seat—an effect that can be called a compliance effect. The theoretical considerations of this compliance effect are discussed later in Appendix [1]. [4] Effect of Bypass Channel Formation [4-1] Comparison of Flow Rate with and without Bypass Channel The following (1) and (2) describe the bypass effect discovered in this study. (1) The "annular nozzle + bypass channel" significantly increases flow rate. (2) By forming a bypass passage in either the intake or exhaust annular nozzle and forming a bypass through-hole in the flapper, both the intake and exhaust flow rates increase. As mentioned above, this bypass effect was fortunately discovered in the process of identifying the factors that prevented the servo valve from achieving a high flow rate, theoretically analyzing these factors, and seeking a solution. Figure 18 is a graph comparing the characteristics of the flow rate Q versus the outer nozzle inner diameter d0 for two cases: one with a bypass passage in the annular nozzle and one without. The analysis conditions, as shown in the figure, were the maximum flapper displacement X0 = 180 μm, the disk (flapper) spring stiffness K0 = 43.9 N / mm, and the intake side pressure P S = 0.6MPa (abs). In addition, the inner diameter of the outer nozzle d0 and the outer diameter of the inner nozzle d i The difference Δr=d0-d i = 0.72 mm, so that the outer diameter of the inner nozzle d iWhen the outer nozzle inner diameter d0 is 3.0 mm, the flow rate Q when a bypass flow path is formed (point A) is 129 L / min, while the flow rate Q when no bypass flow path is formed (point B) is 72.3 L / min. The ratio of the two is 1.78. When the outer nozzle inner diameter d0 is 7.0 mm, where the flow rate is at its maximum, the flow rate Q when a bypass flow path is formed (point A') is 209 L / min, while the flow rate Q when no bypass flow path is formed (point B') is 110 L / min. The ratio of the two is 1.9. From the above results, it can be seen that the flow rate increase effect of forming a bypass flow path increases by just under two times as the outer nozzle inner diameter d0 increases. [4-2] Qualitative explanation of the flow rate increase effect of the bypass flow path (1) Why the intake flow rate doubles with the bypass flow path Figures 19a and 19b show the cases where the flapper 162 is in close contact with the tip end of the exhaust-side annular nozzle 171 (the outer valve seat tip end 316c and the inner valve seat tip end 311c) or where there is a narrow gap between them immediately after the pneumatic actuator starts intake. As a result, the exhaust-side annular flow path 317 is shielded from the control chamber 176. Figure 19a shows the case where the intake-side annular nozzle 170 having the bypass flow path 302 described above is used, and Figure 19b shows the case where the intake-side annular nozzle 180 not forming the bypass flow path is used. In both cases, the pressure P a is the supply pressure P S the inspiratory phase before reaching a <P S) is assumed. In Figure 19a, when the bypass passage 302 is provided, the fluid flowing out of the annular passage 308 branches into a centrifugal direction and a centripetal direction at the tip of the intake-side annular nozzle. The fluid flowing out in the centripetal direction flows into the bypass passage 302, merges with the fluid that has passed through the control chamber 176 and the control port 161 (see Figure 6c), and flows into the pneumatic actuator. In Figure 19b, when the bypass passage is not provided, the fluid flowing out of the annular passage 308 flows only in the centrifugal direction at the nozzle tip, and does not flow in the centripetal direction. This is because the downstream side of the centripetal flow passage is completely blocked from the exhaust side or has a narrow gap, as shown by the dashed line area. The difference between Figure 19a and Figure 19b shows the effect of the bypass passage 302 in increasing the intake flow rate. (2) Why the exhaust flow rate increases with the "bypass through-hole + bypass passage": If the intake flow rate is increased, the exhaust flow rate must also be increased. The reason for this is that the rise time and fall time of the actuator are usually required to be shortened on the same order of time. Figures 20a and 20b show the cases where the flapper 162 is in close contact with the tips (outer valve seat tip 309c and inner valve seat tip 303c) of the intake side annular nozzles 170, 180 or where there is a narrow gap between them immediately after the air filled in the pneumatic actuator starts to be exhausted. Figure 20a shows the case where the intake side annular nozzle 170 having the bypass flow path 302 described above is used, and Figure 20b shows the case where the intake side annular nozzle 180 not forming the bypass flow path is used. In both cases, the pressure P a is the supply pressure P S The exhaust stage after reaching P a <P S) is assumed. In FIG. 20A, when the bypass flow path 302 is provided, fluid flows into the annular flow path 317 from both centrifugal and centripetal directions at the tip of the exhaust-side annular nozzle 171. Fluid flowing in from the pneumatic actuator via the control port flows into the annular flow path 317 from the outer valve seat tip 316c. Fluid flowing in from the pneumatic actuator via the bypass flow path 302 passes through the bypass through-hole 177 formed in the flapper 162 and flows into the annular flow path 317 from the inner valve seat tip 311c. In FIG. 20B, when the bypass flow path is not provided, fluid flows into the annular flow path 317 only from the outer valve seat tip 316c at the tip of the exhaust-side annular nozzle 171, and no fluid flows in from the inner valve seat tip 311c. This is because the upstream side of the inner valve seat tip 311c is completely blocked from the intake side or has a narrow gap, as shown by the dashed line area. 20a and 20b show the effect of the bypass flow path 302 in increasing the exhaust flow rate, and the bypass through-hole 177 and the bypass flow path 302 play important roles in achieving this effect. To summarize this embodiment, as shown in FIG. 19a, when the flapper 162 is in close contact with the tip of the exhaust-side annular nozzle 171, the bypass through-hole is located at a position that does not interfere with the flapper 162 blocking the exhaust-side annular flow path 317. Also, as shown in FIG. 20a, when the flapper 162 is in close contact with the tip of the intake-side annular nozzle 170, the flapper 162 blocks the intake-side annular flow path 308 without any problems, and a flow path connecting the bypass flow path 302 and the exhaust-side annular flow path 317 is formed via the bypass through-hole. From the above considerations, by forming a bypass through-hole 177 in the flapper 162 to form a bypass flow path on either the intake side or the exhaust side, it is possible to significantly increase both the intake flow rate and the exhaust flow rate. (3) Diameter Conditions of the Bypass Through-hole The diameter conditions required for the bypass through-hole 177 formed in the flapper are determined below. In FIG. 20A, in which the intake-side annular nozzle 170 having the bypass flow passage 302 described above is applied, the gap between the inner valve seat tip 311c and the flapper 162 is X eff , the minimum outer diameter of the annular inner nozzle on the exhaust side d irWhen the effective flow area A connecting the bypass through-hole 177 and the exhaust-side annular flow passage 317 is ei teeth The approximate inner diameter of the bypass through-hole 177 is d B as If the bypass through-hole 177 is formed so that the above formula (3) is satisfied, the fluid flowing in from the bypass flow path 302 can flow into the exhaust-side annular flow path 317 without being restricted by the bypass through-hole 177. However, if the minimum outer diameter d of the inner nozzle in the exhaust-side annular nozzle portion ir When the exhaust-side annular flow passage 317 is blocked (FIG. 19a), B <d ir It is necessary to set the outer diameter of the inner nozzle d i and the minimum outer diameter of the inner nozzle d ir The relationship between the radial seal width at the inner valve seat tip 311c and the radial seal width is Δr, and i =d ir +2Δr. Therefore, Within the range of the above formula, the approximate inner diameter d of the bypass through-hole BBy setting the above, the exhaust flow rate can be increased, just like the intake flow rate. The fact that the exhaust flow rate can be increased without forming a bypass passage in the exhaust-side annular nozzle is extremely effective in simplifying the overall structure of the servo valve. The reason for this is that, as can be seen from the valve body diagram in Figure 2, the exhaust side magnetic actuator must be constructed around the central axis 150, consisting of the coil bobbin 153, coil 154, outer frame 152, etc., and forming a bypass passage is not easy. [5] Evaluation by Computational Fluid Analysis: The effects of (1) increasing the intake flow rate and (2) increasing the exhaust flow rate due to the formation of a bypass passage, which were qualitatively explained in the previous section, are more strictly evaluated using computational fluid analysis. [5-1] Numerical Analysis (1) Numerical Analysis Model Figures 21a and 21b are numerical analysis models that replace Figures 19 and 20, which qualitatively explain the bypass effect in Section [4-2]. Figure 21a shows the case where the intake-side annular nozzle 170 having the bypass flow path 302 described above is applied, while Figure 21b shows the case where the intake-side annular nozzle 180 without the bypass flow path is applied. (2) Analysis Conditions The analysis conditions are the same as those applied in Section [2-5] "Effects of the Invention (Summary)." The annular nozzle has the same specifications on the intake and exhaust sides, with an outer nozzle inner diameter d0 = 3.0 mm, an inner nozzle outer diameter d i = 2.28 mm, bypass diameter d B = 1.2 mm, and the maximum displacement of the flapper (movable body) X0 = 180 μm. The intake pressure is P S = 0.6MPa (abs). [5-2] Evaluation of maximum intake flow rate To obtain the maximum intake flow rate, the pressure P a is atmospheric pressure, and the exhaust-side annular flow path 317 is shielded from the control chamber 176. In other words, the flapper 162 is in close contact with the tip of the exhaust-side annular nozzle 171. [5-2-1] When a bypass flow path is formed (1) Analysis results of static pressure distribution Figure 22 shows the static pressure distribution near the intake and exhaust nozzles for the intake-side annular nozzle 170 with a bypass flow path. Figure 22a is an overall view of the object to be analyzed, and Figure 22b is an enlarged view of the dashed line part in Figure 22a. As a boundary condition, a total pressure P T (dynamic pressure PV + static pressure P Sa ) = 0.5 MPaG (gauge pressure), and the static pressure P Sa is shown in a contour diagram. Since the annular flow path boundary surface 901 has a flow velocity in the axial direction, the dynamic pressure P V The static pressure P of the boundary surface 901 S The air chamber of the air pressure actuator (not shown) is set to atmospheric pressure, and the total pressure P T = 0.0 MPaG. Similarly, the total pressure P T On the surface (chain circle 905a) facing the flapper 162 of the annular nozzle 308, the static pressure P Sa The reason why the dynamic pressure P V In addition, in the bypass flow passage 302, the static pressure P Sa The reason why the pressure decreases and becomes negative is that the dynamic pressure P VThis is because the velocity vector increases. (2) Analysis Results of Velocity Vectors Figure 23 shows velocity vectors near the intake and exhaust nozzles for the intake-side annular nozzle 170 with a bypass flow path. The fluid flowing out of the annular flow path 308 branches into a centrifugal direction (dashed circle 905c) and a centripetal direction (dashed circle 905d) at the tip of the intake-side annular nozzle. The fluid flowing in the centripetal direction flows into the bypass flow path 302. The difference in flow path area is the reason why the flow velocity of the fluid flowing in the centrifugal direction and the centripetal direction differ significantly. That is, the fluid flowing in the centrifugal direction flows out from the outer valve seat tip 309c (see Figure 6b) into the large control chamber 176, while the fluid flowing in the centripetal direction converges from the inner valve seat tip 303c into the pipe-shaped bypass flow path 302 (inner diameter = 1.2 mm) with a small flow path area. (3) Pressure and axial velocity distribution Figure 24 shows the pressure and axial velocity distribution near the flapper-side opening end of the annular passage 308 in the case of the intake-side annular nozzle 170 with bypass passage, calculated based on the analysis results of the pressure distribution and velocity vectors described above. Figure 24a is an analytical model diagram, and Figure 24b shows the static pressure versus radial position with the axis of the annular passage 308 as the origin, and Figure 24b shows the axial flow velocity. In Figure 24b, the static pressure P Sa The steep peak P SMAX This peak value P = 0.37 MPa occurs. SMAX corresponds to the maximum pressure point (chain circle 905a) in the pressure contour diagram of Fig. 22b. Also, in Fig. 24c, inside the position of the pressure peak, the flow velocity reaches a maximum value V MAX=130 m / s, but the point where the flow velocity is maximum is the point where the flow enters the bypass flow path 302 from the inner valve seat tip 303c of the annular nozzle 170, as can be seen from the velocity vector diagram (chain circle 905d) in Figure 23b. [5-2-2] Case where a bypass flow path is not formed (1) Analysis results of static pressure distribution Figure 25 shows the static pressure distribution near the intake-side annular nozzle when a bypass flow path is not formed and an intake-side annular nozzle 180 is used. Figure 25a is an overall view of the analysis target, and Figure 25b is an enlarged view of the chain line portion in Figure 25a. Except for the fact that a bypass flow path is not formed, the boundary conditions of the annular flow path boundary surface 901 and the control-side flow path boundary surface 903 are the same as when a bypass flow path is formed (Figure 22). The following discussion will be based on a comparison with the case where a bypass flow path is formed (Figure 22). The static pressure P of the annular flow path 308 Sa The reason why increases somewhat is that when no bypass flow path is formed, the intake flow rate Q decreases, and the flow velocity and dynamic pressure P VThis is because the pressure decreases. The space connecting the flapper-side opening end and the center of the annular flow passage (dashed line 905e) maintains a constant pressure equilibrium state, and no radial pressure gradient occurs. (2) Analysis Results of Velocity Vectors Figure 26 shows velocity vectors near the intake and exhaust nozzles for the intake-side annular nozzle 180 without a bypass flow passage. Figure 26a shows an overall view of the analysis target, and Figure 26b shows an enlarged view of the dashed line portion in Figure 26a. Fluid flowing out of the annular flow passage 308 flows in the centrifugal direction (dashed line circle 905f) at the tip of the intake-side annular nozzle, but does not flow in the centripetal direction (dashed line circle 905g). The reason the fluid does not flow in the centripetal direction is that, as mentioned above, the space surrounding the bypass through-hole 177 (dashed line 905e in Figure 25b) maintains a constant pressure equilibrium state, and no centripetal pressure gradient occurs. [5-2-3] Bypass effect on intake flow rate (summary) From the above analysis results, it was found that the fluid flowing out of the annular nozzle normally flows in a centrifugal direction, but not in a centripetal direction. However, by forming a bypass flow path, it was found that the fluid branches off at the tip of the annular nozzle and flows out in both a centrifugal and centripetal direction. Therefore, it was found that forming a bypass flow path at the same time as using an annular nozzle is an effective method for significantly increasing the valve flow rate. Table 1 shows the results of numerical calculations comparing the presence and absence of a bypass flow path. [5-3] Evaluation of maximum exhaust flow rate As mentioned above, when the intake flow rate is increased, the exhaust flow rate must also be increased. This is because the rise time and fall time of the actuator usually require a time reduction of the same order. In order to find the maximum exhaust flow rate, the pressure P a is the supply pressure P S0 reached (P a = P S0) and the intake-side annular flow passage 308 is shielded from the control chamber 176. In other words, the flapper 162 is in close contact with the tip of the intake-side annular nozzle 170 with a bypass flow passage or the intake-side annular nozzle 180 without a bypass flow passage. [5-3-1] When a bypass flow passage is formed (1) Analysis results of static pressure distribution Figure 27 shows the static pressure distribution near the exhaust-side annular nozzle when the intake-side annular nozzle 170 with a bypass flow passage is applied. Figure 27a is an overall view of the object to be analyzed, and Figure 27b is an enlarged view of the dashed line part in Figure 27a. The air chamber of the pneumatic actuator (not shown) is set to supply pressure, and as boundary conditions, the total pressure P T = 0.5 MPaG. The total pressure P T =0.0 MPaG. The static pressure P Sa The reason why the static pressure P decreases is because the fluid flows from the bypass flow path to the exhaust-side annular nozzle side through the bypass through-hole 177. SaThe pressure gradient at the exhaust-side annular nozzle tip (905j, 905i) is due to a sudden change in the flow velocity through a narrow gap, as shown in the velocity vector analysis results (Figure 28b) on the next page. (2) Velocity Vector Analysis Results Figure 28 shows velocity vectors near the exhaust-side annular nozzle when the bypass-equipped intake-side annular nozzle 170 is used. Figure 28a shows an overall view of the analysis target, and Figure 28b shows an enlarged view of the dashed line portion in Figure 28a. It can be seen that fluid flows into the exhaust-side annular passage 317 from two directions: (i) from the bypass passage 302 via the exhaust-side annular nozzle tip (dashed circle 905j) to the annular passage 317, and (ii) from the control chamber 176 via the exhaust-side annular nozzle tip (dashed circle 905i) to the exhaust-side annular passage. These analysis results demonstrate the important findings of this study. That is, both bidirectional flapper valves are configured with annular nozzles, with the intake-side annular nozzle 170 having a bypass passage 302 and the exhaust-side annular nozzle 171 having no bypass flow path. However, if bypass through-holes 177 are formed in the flappers 162 facing the two annular nozzles, the exhaust flow rate can be increased, just like the intake flow rate. For example, as described in the first embodiment ( FIG. 2 ), consider a case in which an actuator that magnetically attracts the flapper 162 is located on the exhaust side. This actuator is composed of a cylindrical central shaft 150, a coil bobbin 153, a coil 154, and other components. It is easy to imagine that the presence of this actuator would make it difficult to form a bypass flow path returning to the control chamber 176. However, even in this case, the combination of the bypass through-holes 177 and the intake-side annular nozzle 170 with a bypass compensates for this drawback. [5-3-2] When a bypass passage is not formed (1) Analysis results of static pressure distribution Figure 29 shows the static pressure distribution near the exhaust-side annular nozzle when an intake-side annular nozzle 180 is used without a bypass passage, with Figure 30a showing the overall view of the analysis target and Figure 29b showing an enlarged view of the dashed line portion in Figure 29a. Except for the fact that a bypass passage is not formed, the boundary conditions of the exhaust-side annular passage boundary surface 904 and the control-side passage boundary surface 903 are the same as when a bypass passage is formed (Figure 27).At the tip of the exhaust-side annular nozzle, a pressure gradient occurs at the tip of the outer valve seat (dashed circle 905i), but no pressure gradient occurs at the tip of the inner valve seat (dashed circle 905j). This is because the space (dashed line 905k) surrounding the bypass through-hole 177 from above and below is a space where pressure is in a constant equilibrium state, and no fluid flows into or out of the space (dashed line 905k) from the outside. (2) Velocity Vector Analysis Results Figure 30 shows velocity vectors near the exhaust-side annular nozzle when the intake-side annular nozzle 180 without a bypass flow path is used. Figure 30a is an overall view of the analysis target, and Figure 30b is an enlarged view of the dashed line portion in Figure 30a. As can be seen from the velocity vector diagrams of the outer valve seat tip (dashed circle 905i) and the inner valve seat tip (dashed circle 905j), only the control chamber 176 flows into the exhaust-side annular flow passage 317. Therefore, the exhaust flow rate is significantly reduced compared to when a bypass flow path 302 is formed (Figures 27 to 28). [5-3-3] Bypass effect on exhaust flow rate (summary) From the above analysis results, the important findings of this study can be summarized as follows: Both two-way flapper valves are configured with annular nozzles, a bypass path is formed in one of the annular nozzles, and a bypass through-hole is formed in the flapper facing the two annular nozzles. Even if a bypass path is formed in only one of the two annular nozzles, as long as there is a bypass through-hole in the flapper, the exhaust flow rate can be increased in the same way as the intake flow rate. Table 2 shows the results of numerical calculations comparing the presence and absence of a bypass flow path. [6] How to set the maximum stroke of an annular nozzle The flow characteristics of a nozzle flapper valve are determined by the gap where the flow is saturated, h = h S If so, the gap h is small, and h<h S In the range of h ≥ h, the flow rate increases in proportion to the gap h, so the flow rate can be controlled by varying the gap h. S In this range, the flow rate is constant and does not depend on the gap h, so flow rate control is not possible. maxThe setting of is extremely important in determining the magnetic circuit including the electromagnet, the electric circuit, and the mechanism. [6-1] Valve flow rate characteristic curve (1) Profile of valve flow rate characteristics Figure 31 shows the valve flow rate characteristics for the gap in the case of an annular nozzle with a bypass flow path, calculated by computational fluid analysis and further verified by experiment. The specifications of the annular nozzle in this embodiment are outer nozzle inner diameter d0 = 3.0 mm, inner nozzle outer diameter d i = 2.4 mm. Numerical analysis revealed that the characteristic curve of the annular nozzle valve flow rate consists of three regions. (i) The region 0 ≦ h < h0 is the region where the valve flow rate Q increases linearly with respect to the gap. When the annular nozzle has an axisymmetric structure, the fluid flows only in the radial direction between the valve seat and the valve disc (disk), which is a two-dimensional flow. (ii) h0 ≦ h < h S In this region, the gradient of the flow rate Q with respect to the clearance h gradually decreases from point A until it eventually reaches saturation point B. The air flow between the valve seat and the valve body is a three-dimensional flow in which the axial flow is added to the radial flow. This region is called the "transition region." (iii) h ≥ h S The region where the valve flow rate is saturated and becomes a constant value. (2) Maximum stroke of the valve h max How to set the maximum stroke h of the servo valve with an annular nozzle max We propose a method to set h0≦h in the following transition region: max <h S ... (5) By setting it in the above range, sufficient valve flow rate can be obtained while keeping the intake nozzle opening area the same, that is, maintaining the same fail-safe force. As mentioned above, the fail-safe force is the spring restoring force of the disk required to block the intake nozzle before the valve operates. Increasing the spring stiffness to increase the spring restoring force requires increasing the power of the electromagnet, but h max As long as the stroke is set within the above range, these modifications are not necessary. maxIf h0 is set within the range below, the characteristics of the flow rate Q with respect to the gap h can be maintained almost linearly. Therefore, in vibration control, it is advantageous to create an accurate vibration suppression force waveform to attenuate the excitation force (disturbance) applied to the vibration isolation table. h0≦h max <h C ... (6) Let envelope a be the region where the flow rate characteristics are linearly proportional to the gap h (0≦h≦h0), and envelope b be the region where the flow rate is a constant value. From the intersection C of the two envelopes, C The saturation point B of the valve flow characteristics can be found by computational fluid analysis or experiment. [6-2] Method for finding the gap h=h0 from the geometric conditions of the flow path area Below, we will find the value of the gap h=h0 where the gradient of the flow rate relative to the gap switches, for both the case where a bypass flow path is not formed and the case where a bypass flow path is formed. (1) When a bypass flow path is not formed First, we will find the condition for flow rate saturation when a bypass flow path is not formed. Flow path area A in the range of h<h0 R2 is the annular flow area A that flows centrifugal from the outer valve seat R2 It shall be determined by. Flow area A in the range of h>h0 Z2 is the cross-sectional area S(=A Z2 ) is assumed to be determined by From equations (6) and (8), A R2 =A Z2 The flow direction switching gap h=h 02 Ask for. (2) When a bypass flow path is formed, the fluid flowing out of the annular nozzle branches into two directions: toward the outer valve seat (centrifugal direction) and toward the inner valve seat (centripetal direction). The inner diameter of the outer nozzle of the annular nozzle is d0, and the outer diameter of the inner nozzle is d i Assuming that the flow area A in the range of h < h0 R teeth Flow area A in the range of h>h0 Z is the cross-sectional area S (=A Z ) is assumed to be determined by From equations (7) and (8), A R = A Z Then, the flow path direction switching gap h=h0 is calculated.

[0060] Second Embodiment Figure 32 shows a pneumatic servo valve (fluid control valve) according to a second embodiment of the present invention. Figure 32a is a front cross-sectional view of the entire servo valve, with the intake nozzle in Figure 32c taken along the line BB. Figure 32b is a front cross-sectional view of the entire servo valve, with the intake nozzle in Figure 32c taken along the line CC. Figure 32c is a view taken along the line AA in Figure 32a. In the first embodiment, the bypass flow path, which is highly effective in increasing flow rate, was formed using the following structure. In Figure 2, a base member 164 was attached to the top of the intake-side housing 159. An intake-side passage A 167 and a control-side passage B 169 were formed within this base member. These passages were connected to the passages formed in the intake-side housing 159, forming a bypass flow path and annular nozzle. This embodiment achieves the same effects as above with a simpler nozzle structure. In Figures 32a and 32b, reference numeral 651 denotes an annular nozzle with a built-in bypass passage (first nozzle portion), which is an intake-side (supply-side) nozzle portion attached to an intake-side housing 652. It is composed of an inner member 653 and an outer member 654. In this embodiment, the exhaust-side (discharge-side) nozzle portion 171 (second nozzle portion) has the same structure as in the first embodiment. The annular nozzle with a built-in bypass passage 651 can be handled independently as a single nozzle unit, similar to a single-hole nozzle. The flapper sides of the inner member 653 and the outer member 654 form an outer valve seat 654a and an inner valve seat 655a. A seal portion with a width of 0.1 to 0.3 mm is formed at the tip of the outer valve seat 654a and the tip of the inner valve seat 655a, and the flapper 162 tightly contacts the seal portion to block the annular passage on the intake side (not shown). Reference numeral 666 denotes an intake-side flow passage A formed by the upper surface of the inner member 653 and the intake-side housing 652; 666a denotes an intake hole; 667a, 667b, 667c, and 667d denote intake-side flow passage B formed axially penetrating the inner member 653; and 668 denotes an annular flow passage formed by the inner circumferential surface of the outer member 654 and the outer circumferential surface of the inner member 653. That is, fluid flowing into this servo valve from a fluid supply source (not shown) passes through the intake hole 666a, intake-side flow passage A 666, intake-side flow passage B 667, and annular flow passage 668 before flowing into the intake-side gap 172. This flow passage is referred to as flow passage (1). Reference numeral 669 denotes a bypass flow passage A formed in the center of the inner member.Bypass flow paths 670a, 670b, 670c, and 670d are bypass flow paths B formed by radially penetrating the inner member 653, and bypass flow paths 671a, 671b, 671c, and 671d are bypass flow paths C formed by axially penetrating the inner member 653. That is, fluid flowing inward from the annular flow path 668 reaches the intake-side gap 172 via the bypass flow path of "bypass flow path A 669 → bypass flow path B 670 → bypass flow path C 671." This bypass flow path is referred to as flow path (2). In this embodiment, the intake hole 666a, the exhaust hole 157, and the control port 161a form a two-way, three-way valve, similar to the first embodiment. In summary, this embodiment has the bypass flow path having a second opening directly connected to the control chamber and a through-passage connecting the annular flow path to the supply side, which are formed independently of each other, inside the intake-side nozzle portion (annular nozzle with a built-in bypass flow path) 651 formed by the inner member and the outer member. This annular nozzle with a built-in bypass passage may also be provided on the discharge side. In this case, the bypass passage having a second opening directly connecting to the control chamber and a through-passage connecting the annular passage to the discharge side may be formed independently within the discharge-side nozzle section composed of the inner member and the outer member. Figure 33 shows an annular nozzle 651 with a built-in bypass passage configured by inserting an inner member 653 into an outer member 654. Figure 34a is a front cross-sectional view of the annular nozzle 651 with a built-in bypass passage, and Figure 34b is a DD cross-sectional view of Figure 34a. The annular nozzle 651 with a built-in bypass passage incorporates the two passages (1) and (2), i.e., the intake-side passage and the bypass passage, into a single nozzle unit. This simplifies the installation of the nozzle unit into the valve body, as with conventional single-hole nozzles, and simplifies the valve body.

[0061] Third Embodiment Figure 35 is a front cross-sectional view of a pneumatic servo valve (fluid control valve) according to a third embodiment of the present invention. This servo valve uses an electromagnetic actuator to drive a flapper, and shows a structure that increases the effective flow area of ​​the annular nozzle applied to the exhaust side to the same level as the intake side. Figures 36 and 37 show the component configuration of the magnetic poles and the exhaust-side annular nozzle, and Figure 38 shows the flow of magnetic flux between the exhaust-side components. [1] Basic Valve Structure: Since the intake side of this servo valve has a structure nearly identical to that of the first embodiment, the following description of the valve structure will focus mainly on the exhaust side. Reference numeral 750 denotes a cylindrical central shaft (support shaft), 751 denotes the bottom of this central shaft, 752 denotes the outer frame of the central shaft formed concentrically with the central axis of the central shaft, 753 denotes a coil bobbin attached to the central shaft, and 754 denotes a coil wound around the coil bobbin. An electromagnetic actuator that attracts the flapper and controls its displacement is constituted by central shaft 750, central shaft bottom 751, outer frame 752 of the central shaft, coil bobbin 753, and coil 754. Reference numeral 755 denotes a cylindrical exhaust-side housing that houses bottom 751 and outer frame 752 of the central shaft, 756 denotes the bottom of this exhaust-side housing, 757 denotes an exhaust hole formed in housing bottom 756, and 758 denotes an exhaust-side flow passage C formed in central shaft 750. Reference numeral 759 denotes an intake-side housing, 760 denotes an intake-side flow passage B formed in this intake-side housing, 761 denotes a control-side flow passage A connected to the pneumatic actuator, 762 denotes a disk-shaped flapper (movable body), 763 denotes a positioning pin, 764 denotes a base member, 765 denotes an outer sleeve that houses the outer peripheries of intake-side housing 759 and exhaust-side housing 755, and 766a and 766b denote seals. Reference numeral 767 denotes an intake-side flow path A formed inside this base member, 768 an intake hole, and 769 a control-side flow path B. Reference numeral 770 denotes an intake-side (supply-side) nozzle section (first nozzle section) which is a forward-direction nozzle attached to the center of intake-side housing 759, and an annular nozzle is formed by combining an inner member 771 and an outer member 772. Reference numeral 773 denotes an exhaust-side (discharge-side) nozzle section (second nozzle section) which is a reverse-direction nozzle attached to the flapper side of central shaft 750, and its detailed structure will be described later with reference to Figures 36 and 37.Reference numeral 774 denotes an intake-side gap, 775 denotes an exhaust-side gap, and 776 denotes a second magnetic pole formed on the flapper-side end face of outer frame 752. The intake-side gap and the exhaust-side gap are connected by a groove 777 formed in flapper 762, forming a control chamber 778. Reference numeral 779 denotes a bypass flow passage formed in the center of inner member 771, which communicates with control-side flow passage B 769. Reference numeral 779a denotes a control port. Reference numeral 780 denotes an intake-side gap formed at the flapper-side open end of this bypass flow passage. Reference numeral 781 denotes a plurality of bypass through-holes formed circumferentially near the center of flapper 762. Similar to the previously described embodiment, these bypass through-holes play an important role in connecting intake-side gap 780 and the exhaust-side gap (described below). [2] Structure of the Exhaust-Side Nozzle Section Figure 36 shows the component configuration of the exhaust-side nozzle section (second nozzle section), with Figure 37a being a top view of the exhaust-side nozzle section, Figure 37b being a front cross-sectional view, and Figure 37c being a view taken along the line AA in Figure 37b. Reference numeral 782 denotes an inner member, 783 an outer member, and both the inner member and the outer member are made of a non-magnetic material. Reference numeral 782a denotes an inner valve seat formed at the tip of the inner member, and 783b an outer valve seat formed at the tip of the outer member. Reference numeral 784 denotes an annular flow passage formed by the inner circumferential surface of the outer member and the outer circumferential surface of the inner member. The inner valve seat, the outer valve seat, and the annular flow passage constitute the exhaust-side annular nozzle. Reference numeral 785 denotes a first magnetic pole, which is the flapper-side end surface of central shaft 750. Reference numerals 786a, 786b, 786c, and 786d denote exhaust-side flow passages A formed radially penetrating the central shaft. Reference numeral 787 denotes exhaust-side flow passage B formed axially to communicate with exhaust-side flow passage A. Therefore, the exhaust flow path for air flows through control chamber 778 → annular flow passage 784 → exhaust-side flow passage A 786 → exhaust-side flow passage B 787 → exhaust-side flow passage C 758 → exhaust hole 757, before flowing to the atmosphere. Reference numeral 788 denotes an inner member mounting portion on the central shaft, and 789 denotes an outer member mounting portion on the central shaft. As shown in FIG. 36 , the inner and outer members, which are made of non-magnetic materials, are press-fitted into the inner and outer member mounting portions.It has been found that when a magnetic material is used for the valve seat, the disk rapidly adheres to the valve seat when the valve seat tip approaches the magnetic flapper (disk). This occurs because, as the gap between the valve seat tip and the flapper narrows, the magnetic flux density between the valve seat tip and the opposing surface increases locally. In the structure of this embodiment, when the flapper approaches the exhaust-side nozzle portion 773, the flapper comes to rest in close contact with the non-magnetic nozzle tip. Since an appropriate gap is maintained between the magnetic first magnetic pole and the flapper, the adhesion phenomenon does not occur. Reference numeral 791 denotes an exhaust-side gap formed by the side of the first magnetic pole and the inner periphery of the inner valve seat. A bypass through-hole 781 is formed in the flapper at a position opposite this exhaust-side gap. Figure 38 shows the flow of magnetic flux between the exhaust-side components, forming a closed-loop magnetic circuit: central shaft 750 → outer member mounting portion 789 → inner member mounting portion 788 → first magnetic pole 785 → flapper 762 → second magnetic pole 776 → outer frame portion 752 of the central shaft → central shaft 750. When applying the present invention to a magnetically attracted servo valve to increase flow rate, it is necessary to increase the outer diameters of both the intake-side and exhaust-side annular nozzles. While there is a high degree of freedom in setting the outer diameter of the intake-side annular nozzle, there are significant constraints on the exhaust-side. This is because the outer diameter of the servo valve as a finished product is heavily dependent on the outer diameter of the first magnetic pole to which the coil bobbin is attached. The first embodiment (Figure 2), in which the annular nozzle is mounted inside the first magnetic pole, is compared with this embodiment (Figure 36), in which the annular nozzle is located on the outer periphery of the first magnetic pole. This embodiment allows for an increased exhaust flow rate while maintaining the same outer diameter of the servo valve.

[0062] Fourth Embodiment In the fourth embodiment of the present invention, the nozzle structures of the intake and exhaust nozzles are simplified to simplify the configuration of the servo valve main body. The first is [1] a case in which an annular nozzle is used on both the intake and exhaust sides, and a bypass flow path is omitted for both nozzles. The second is [2] a case in which an annular nozzle without a bypass flow path is used on the intake side, and a single-hole nozzle is used on the exhaust side. [1] Annular Nozzles Applied to the Intake and Exhaust Sides (1) Specific Structure of the Embodiment Figure 39 shows a front cross-sectional view of the pneumatic servo valve (fluid control valve) according to the first embodiment. This embodiment shows a structure that compensates for the reduction in flow rate when a bypass flow path is omitted on both the intake and exhaust sides. It was found that, as mentioned above, the omission of a bypass flow path significantly reduces the intake flow rate Q when the outer nozzle inner diameter D0 of the annular nozzle is the same. In this embodiment, the reduction in intake flow rate is compensated for by adjusting the outer nozzle inner diameter D0 and the inner nozzle inner diameter D1 of the annular nozzle. iThis structure compensates for the decrease in flow rate by increasing the valve seat diameter. The exhaust-side structure of this embodiment is generally the same as that of the third embodiment, except for the intake-side annular nozzle. That is, the exhaust-side annular nozzle forms an annular flow passage around the outer periphery of the magnetic pole, with the outer valve seat and the inner valve seat as peripheral walls. In FIG. 39 , reference numeral 700 denotes an intake-side (supply-side) nozzle portion (first nozzle portion), 701 denotes an intake-side housing, 702 denotes an intake-side flow passage A formed on the upper end surface of the intake-side housing, 702a denotes an intake hole, 703 denotes an inner member constituting the annular nozzle, 704 denotes an outer member, 705 denotes an intake-side flow passage B formed in multiple circumferential directions through the inner member in the axial direction, 706 denotes a gap formed near the opening of the intake-side flow passage B on the flapper 762 side, 707 denotes an annular flow passage formed by the inner circumferential surface of the outer member and the outer circumferential surface of the inner member, and 708a and 708b denote small-diameter perforations formed in the flapper. Reference numeral 704a denotes an outer valve seat formed on the flapper side of the outer member, and 703a denotes an inner valve seat formed on the flapper side of the inner member. That is, the outer valve seat 704a and the inner valve seat 703a form an intake-side annular nozzle. Reference numeral 709 denotes a control-side flow passage, 709a denotes a control port, and 710 denotes an intake-side gap formed on the inner circumferential side of the inner valve seat 703a. Similarly, the exhaust-side annular nozzle has an inner valve seat 782a formed on the flapper side of the inner member 782, an outer valve seat 783a formed on the flapper side of the outer member 783, and an annular flow passage 784. That is, the outer valve seat 783a and the inner valve seat 782a form an exhaust-side (discharge-side) annular nozzle 785 (second nozzle portion). The small diameter perforations 708a, 708b formed in the flapper are formed to achieve the combined effects of (i) preventing the flapper from being adsorbed to the exhaust nozzle and (ii) preventing an increase in the fail-safe force due to the intake side gap 710. In this embodiment shown in Figure 39, the intake side annular nozzle has an outer nozzle inner diameter D0 and an inner nozzle inner diameter D i , the outer nozzle inner diameter d0 of the exhaust side annular nozzle, and the inner nozzle inner diameter d i Then, D0≒d0, and D i ≒d iThis configuration is designed to achieve the above. In other words, the outer diameters of the annular nozzles on the intake and exhaust sides are sufficiently large, and the shapes of the intake and exhaust sides are roughly the same, resulting in a symmetrical nozzle structure. In other words, the intake flow rate and the exhaust flow rate are nearly identical, simplifying the overall structure compared to a valve structure with a bypass flow path. (2) Effect of Perforations Formed in the Flapper As described above, in this embodiment, small-diameter perforations (anti-adsorption holes) 708a and 708b are formed in the flapper, which faces the exhaust-side gap 791. The exhaust-side gap 791 is a space formed by the side surface of the first magnetic pole 786 and the inner periphery of the inner valve seat 782a. The effect of these perforations (hereinafter also referred to as anti-adsorption holes) described above in (i) is explained using model diagram 40. However, the shape of the annular nozzle and the position of the perforations differ between model diagram 40 and front cross-sectional view 39. Figure 40a shows a case where perforations 708 are formed, while Figure 40b shows a case where perforations 708 are not formed. In FIG. 40A where the perforated portion 708 is formed, the exhaust side gap 791a at the center of the inner valve seat is connected to the perforated portion 708, so the exhaust side gap maintains the pressure of the control chamber 778. Here, the suction force F applied to the flapper due to the close contact between the flapper and the exhaust side annular nozzle is d The opening cross-sectional area of ​​the exhaust-side annular flow passage is A rd , the pressure difference between the control room and the atmospheric pressure is ΔP d Then, F d = A rd ×ΔP d 40b, where the perforated portion is not formed in the flapper, the pressure in the exhaust-side gap 791a is likely to be atmospheric pressure. The reason for this is that the tip of the outer valve seat 783a is completely sealed by the flapper, so the pressure in the exhaust-side gap 791a, which is close to the annular flow path 784, easily drops to atmospheric pressure. Therefore, when the inner diameter of the tip of the outer valve seat 783a is the exhaust-side outer nozzle inner diameter d0, and the area A rd0 =(d0 / 2) 2 If π, the pressure difference between the control room and atmospheric pressure is ΔP d The flapper suction force F d0 = A rd0 ×ΔP d A rd0 ≫Ard Therefore, the flapper attraction force F d0 ≫F d Since the flapper is in the maximum displacement state, the restoring force of the flapper is F max = K0X max Therefore, F d0 >F max In this case, even if the power is turned off while the flapper is in close contact with the exhaust-side annular nozzle, this close contact state cannot be avoided. Experimental evaluations have shown that if the perforated portion is not formed in the flapper, the servo valve may become uncontrollable. That is, when an annular nozzle is applied to the exhaust side, the presence of the perforated portion is highly effective in avoiding this suction phenomenon. The perforated portion 708 in this embodiment and the bypass through-hole 177 (Figure 2 of the first embodiment) when a bypass path is formed in the annular nozzle have significantly different purposes. The bypass through-hole 177 requires a sufficiently large hole diameter, as shown in Equation (9), to maximize the flow rate. However, the perforated portion 708 in this embodiment can be a fine hole with an inner diameter of 0.5 mm or less. Instead of forming the perforated portion 708 in the flapper, a central through-hole may be formed in the inner member 782 of the exhaust-side annular nozzle, connecting the exhaust-side gap and the control chamber 778 via the central through-hole. In this case, the flapper suction to the exhaust nozzle can still be avoided (not shown). The above has described the effect (i) of the perforated portion formed in the flapper, but the perforated portion also has the effect (ii) of preventing an increase in the fail-safe force due to the intake side gap 710. If the perforated portion is not formed in the flapper, when the flapper is in close contact with the intake side annular nozzle, the pressure in the intake side gap 710 will be less than the supply pressure P S The reason is that the tip of the outer valve seat 704a is completely shielded by the flapper, so the pressure in the intake side gap 710 adjacent to the annular flow passage 707 easily becomes the supply pressure P S When the servo valve starts to operate, the pressure in the control chamber 778 is assumed to be atmospheric pressure. In this case, the inner diameter of the tip of the inner valve seat 703a is set to the intake side inner nozzle outer diameter d i As the area A rsi =(di / 2) 2 If π, the pressure difference between the control chamber and the supply pressure ΔP s A large force F that pulls the flapper away from the intake nozzle is generated. si = A rsi ×ΔP s However, by forming the perforated portion, F si= 0, preventing an increase in the fail-safe force. Figure 41 is a graph showing the relationship between the valve flow rate Q and the flapper stiffness ratio as a function of the outer nozzle inner diameter d0 in this embodiment, in which neither the intake nor exhaust nozzles have a bypass flow path. For example, by changing the flapper stiffness ratio K / K0 from 1 to 2, the valve flow rate can be increased from Q = 108 to 174 L / min when the outer nozzle inner diameter d0 is 8.0 mm. The graph in Figure 41 is compared with the graph in Figure 12, which shows the characteristics of a single-hole nozzle and an annular nozzle with a bypass flow path. To achieve an intake flow rate Q = 200 L / min when the outer nozzle inner diameter d0 is 8 mm, a flapper stiffness ratio K / K0 = 3 is required without a bypass flow path, whereas a K / K0 = 1 is sufficient with a bypass flow path. However, in Figure 12, with a single-hole nozzle, an intake flow rate Q of 100 L / min or less can be achieved when K / K0 = 3. Therefore, the presence or absence of a bypass flow path can be selected based on the requirements of the application of this valve. [2] Annular Nozzle on the Intake Side and Single-Hole Nozzle on the Exhaust Side (1) Specific Structure of the Embodiment Figure 42 shows a front cross-sectional view of the pneumatic servo valve (fluid control valve) of the second embodiment described above. By using an annular nozzle on the intake side and a single-hole nozzle on the exhaust side, the configuration of the servo valve body is further simplified. If a high flow rate is not required to meet the specifications of the target (pneumatic servo system) of this valve, a single-hole nozzle can be used on the exhaust side. The intake side of this embodiment has a structure roughly identical to the annular nozzle of the first embodiment described above (Figure 39). 730 denotes a central axis, 731 denotes an exhaust-side nozzle attached to the flapper side of this central axis, and 732 denotes an exhaust-side flow path formed through the central axis. In this embodiment, the exhaust-side nozzle is a single-hole nozzle with an inner diameter d. In this embodiment and other embodiments, a single-hole nozzle refers to a nozzle in which an axial flow path is formed only in the center. In this embodiment, one single-hole nozzle is used, but a configuration in which multiple single-hole nozzles are arranged in parallel is also possible, and even in this case the configuration on the exhaust side of the servo valve can be sufficiently simplified. 733a and 733b are small-diameter perforations formed in the flapper.This perforated portion is formed at a position radially away from the opening 731a of the exhaust-side nozzle and facing the inner peripheral side of the intake-side gap 710. As in the previously described embodiment (FIG. 42), the perforated portion has the effect of preventing an increase in the fail-safe force when the flapper is in close contact with the intake-side annular nozzle. The flapper is at a fail-safe displacement X. fel Including the maximum displacement X=X max The exhaust nozzle is blocked and the control chamber 778 is in a state where the supply pressure P max =P s The inner diameter of the exhaust nozzle is d, the atmospheric pressure (or exhaust pressure) is P0, and the opening area of ​​the single-hole nozzle is A d =(d / 2) 2 If π, the pressure difference between the control room and atmospheric pressure is ΔP d (=P s -P0) flapper suction force F d = A d ×ΔP d The spring stiffness of the flapper is K0, and the restoring force of the flapper is F max = K0X max Therefore, within the range where the restoring force of the flapper is greater than the suction force of the flapper, the single-hole nozzle opening area A d If you set A, the adsorption phenomenon of the exhaust nozzle to the flapper can be avoided. d <K0X max / (P s -P0) within the range where the conditional expression of the exhaust nozzle opening area A d If the exhaust nozzle is composed of multiple single-hole nozzles, the opening area A dis the sum of the nozzle opening areas. The limit value of the diameter of a single-hole exhaust nozzle (see FIG. 15), which is determined by the phenomenon of adsorption to the flapper, can be larger than the limit value of the diameter of a single-hole intake nozzle (see FIG. 12), which is determined by the fail-safe force. By configuring the exhaust nozzle to satisfy the above condition, it is possible to achieve a sufficient increase in exhaust flow rate using a single-hole nozzle on the exhaust side. In this embodiment, an annular nozzle without a bypass passage is used on the intake side. However, for example, in a valve structure where the size of the outer nozzle inner diameter D0 of the annular nozzle on the intake side is restricted, a combination of "annular nozzle with a bypass passage on the intake side + single-hole nozzle on the exhaust side" may also be used.

[0063] Fifth Embodiment Figure 43 is a front cross-sectional view of a pneumatic servo valve (fluid control valve) according to a fifth embodiment of the present invention. Figure 44a is a view taken along the line AA in Figure 44b, Figure 44b is a top view of Figure 44a, and Figure 44c is a view taken along the line BB in Figure 44b. Unlike the previously described embodiments in which the annular nozzle was constructed from multiple components, this embodiment uses a single annular nozzle. This embodiment achieves the benefits of an annular nozzle (higher flow rate) while maintaining a simple configuration comparable to that of conventional servo valves that use a single-hole nozzle. In Figure 43, reference numeral 801 denotes an intake-side (supply-side) annular nozzle (first nozzle portion) attached to the intake side, 802 denotes an intake port, and 803 denotes an exhaust-side (discharge-side) annular nozzle (second nozzle portion) attached to the exhaust side. In this embodiment, the intake-side annular nozzle 801 and the exhaust-side annular nozzle 803 have the same shape. The servo valve of this embodiment has substantially the same structure as the first embodiment, except for the intake-side annular nozzle and the exhaust-side annular nozzle. As shown in Figure 44c, the annular nozzle is composed of an inner portion (dashed line) 804 and an outer portion (dashed line) 805 formed on the outer periphery of the inner portion. In Figures 44a to 44c, 806 denotes an inner valve seat formed on the flapper side of the inner portion, 807 denotes an outer valve seat formed on the flapper side of the outer member, 808 denotes a seal formed at the tip of the inner valve seat, 809 denotes a seal formed at the tip of the outer valve seat, and 810 denotes a gap formed in the center of the inner valve seat. 811 denotes an annular flow path formed by the outer valve seat and the peripheral wall of the inner valve seat. This annular flow path consists of a complete slit penetrating in the axial direction and localized slits with axial joints. Each slit has a narrow groove width d1 on the flapper side. 812a, 812b, and 812c denote flapper-side complete slits, and 813a, 813b, and 813c denote flapper-side localized slits. Reference numeral 814 denotes a cylindrical flow path connected to the complete slit. On the cylindrical flow path side of the local slit, reference numerals 815a, 815b, and 815c denote a plurality of connecting portions formed to prevent the inner and outer portions from separating. Reference numerals 815a, 815b, and 815c denote cylindrical flow path side complete slits formed with a wide groove width d2 and connected to the flapper side complete slit.816a, 816b, and 816c are localized slits on the cylindrical flow path side, and the depth of each slit is set so that the bottom surface forms the joint. When each slit is formed, for example, using a small-diameter (e.g., Φ0.1 to Φ0.3 mm) end mill, the smaller the end mill diameter, the more limited the machinable slit depth becomes. However, in the annular nozzle of this embodiment, this limitation is overcome because the annular flow path is composed of a combination of slits with narrow groove width d1 and wide groove width d2. While wire-cut electrical discharge machining and other production methods can be used to form the annular flow path, machining alone would result in even greater cost reduction. Similar to the previously described embodiment without a bypass path, the flapper 162 has an anti-suction hole 816 formed opposite the gap 810. In this embodiment, the intake-side annular nozzle 801 and the intake-side housing 159 housing this nozzle are constructed as separate components. Additionally, the exhaust-side annular nozzle 803 and the central shaft 150 that houses this nozzle are configured as separate parts. Instead of the structure of this embodiment, the intake-side annular nozzle and the intake-side housing may be integrally formed from the same material by processing. The same applies to the exhaust-side annular nozzle. By reducing the number of parts, the overall configuration of the servo valve is significantly simplified (not shown).

[0064] Sixth Embodiment Figure 45 shows a sixth embodiment of the pneumatic servo valve (fluid control valve) of the present invention, illustrating the application of the annular nozzle described above to an oscillating servo valve. The basic form of this servo valve is a two-way (bidirectional) three-way valve whose flow path is composed of two-way nozzles (forward nozzle, reverse nozzle) and three openings (intake port, exhaust port, and control port). As mentioned above, when an annular nozzle is applied to an oscillating servo valve, the width of the valve disc (the flapper, which is the movable body) is limited to the same level as the width of the nozzle, which is the valve seat. This puts the oscillating type at a disadvantage compared to a translation type, which has fewer restrictions on valve disc width. However, as shown in this embodiment, an annular nozzle can be applied as long as the maximum flow rate is within the allowable range. Figure 45a is a front cross-sectional view of the servo valve, and Figure 45b is a diagram showing the shape of the flapper tip, which is the opposing surface of the annular nozzle. The outer diameter of the outer valve seat tip of the annular nozzle is Φd 0dThe width of the flapper tip, which is the surface facing the outer valve seat, is D r As, D r >d 0d The magnet assembly 851 constitutes a magnetic circuit including a permanent magnet, the coil 852, the flapper (movable body) 853, the armature 854, the pair of yokes 855a, 855b, and 856a, 856b attached with their tips facing each other, the cylindrical spring 857 called a flexible tube, and the support portion of the flexible tube 858. The magnet assembly 851, the coil 852, the flapper 853, the armature 854, and the flexible tube 857 constitute a torque motor, which is an electromechanical conversion portion. The flexible tube 857 supports the armature 854, which is the moving portion of the torque motor, and also serves as a seal for the gas in the control chamber. The seal portion 859 and the tip of the flapper on the nozzle side 860 are shown. Reference numeral 861 denotes a housing, and 862 denotes a suction-side (supply-side) nozzle section (first nozzle section), a forward-direction nozzle installed through the housing. As described in the first embodiment, this annular nozzle is formed by combining two components (an inner member 863a and an outer member 863b). Reference numeral 864 denotes a reverse-direction nozzle installed in the housing. Similar to the suction-side nozzle section, this annular nozzle is formed by combining two components (an inner member 865a and an outer member 865b). Reference numeral 866 denotes a through-hole formed at the center of the flapper tip, facing the inner member 863a of the suction-side nozzle section and the inner member 865a of the exhaust-side nozzle section. Figure 44b shows a side view of the flapper tip, showing the outer nozzle inner diameter d0 = 3 mm of the annular nozzle used in this embodiment. Reference numeral 867 denotes an intake port, 868 denotes an exhaust port, and 869 denotes a control port, all of which are connected to a pneumatic actuator (not shown). 870 is the control chamber of this servo valve. S The gas is supplied to the control chamber 870 through the forward nozzle 862. At the same time, the gas in the control chamber 870 is discharged to the atmosphere through the reverse nozzle 864. The difference between the amount of gas flowing in from the forward nozzle and the amount of gas flowing out from the reverse nozzle determines the control pressure P aand the outflow rate from the control port 869 is determined. Figure 46a shows the state in which the flapper tip 860 of the flapper 853 is displaced to the intake nozzle portion 862 at its maximum. Figure 46b is a side view of the intake nozzle portion 862. In Figure 46a, the distance L between the rotation center 853a of the flapper and the center 871 of the intake nozzle portion is 10 mm. The total stroke of the flapper tip 860 is set to X0 = 150 μm, and the stroke on one side is set to X0 / 2 = 75 μm. In this case, the flapper 853 tilts toward the intake nozzle portion at a maximum angle of θ = 0.43 deg, with the center 853a as the fulcrum. In Figure 46b, the intake nozzle portion has an outer nozzle inner diameter d0 = 3.0 mm, an inner nozzle outer diameter d i = 2.28 mm, where 872 is the outer valve seat tip of the intake nozzle, 873 is the inner valve seat tip, and 874 is the annular flow path. The above configuration and dimensions will hereafter be referred to as the basic specifications of the oscillating motion type servo valve in this embodiment. As shown in Figure 46a, when an annular nozzle is applied under the above conditions, a gap of δ = 22.5 μm will be generated between the outer valve seat tip 872 and the flapper tip 860. Therefore, when the input current I = 0, this servo valve cannot completely block the intake side flow path, causing air leakage. The same is true on the exhaust side, where the input current I = I maxIn this state, the exhaust flow path cannot be completely blocked. Figure 47 shows solution (1) to the aforementioned problem, which is essentially inevitable when applying a large-diameter annular nozzle to an oscillating motion servo valve. The servo valve has the same basic specifications as described above. The axes of the intake nozzle section 862 and the exhaust nozzle section 864 are installed with an inclination of θ = 0.43°. As a result, the outer valve seat tip 872 of each annular nozzle and the flapper tip 860 can be in complete contact on both the intake and exhaust sides. Figure 48a shows solution (2) to the aforementioned problem. Figure 48b shows the shape of an improved flapper tip 875. The servo valve is the same as the basic specifications described above, except for the shape of the flapper tip 875. In solution (2), the left and right surfaces of the flapper tip 875 are inclined at θ = 0.43°. As a result, the outer valve seat tip 872 and the flapper tip 875 of each annular nozzle can be in perfect contact with each other on both the intake and exhaust sides. Figure 49a illustrates solution (3) for the aforementioned problem. The servo valve has the same basic specifications as described above, except for the shapes of the intake nozzle section 876 and the exhaust nozzle section 877. In solution (3), the outer valve seat tip 876a and the inner valve seat tip 876b of the intake nozzle section 876, and the outer valve seat tip 877a and the inner valve seat tip 877b of the exhaust nozzle section 877 are inclined at θ = 0.43°. As a result, the tip of each annular nozzle and the flapper tip 875 can be in perfect contact with each other on both the intake and exhaust sides. As described above, in the specifications of the oscillating motion servo valve of this embodiment, the flapper is inclined toward the annular nozzle at a maximum angle of θ = 0.43°, with the center as the fulcrum. As a result, a gap of δ = 22.5 μm is generated between the tip of the outer valve seat and the tip of the flapper. Solutions (1) to (3) were implemented to prevent air leakage due to this gap. As an alternative, as shown in Figure 49b, taking into account that the gap is very small (several tens of microns), a structure may be used in which the tip of the flapper elastically deforms in the same direction as the swinging motion. For example, the thickness of the area indicated by the chain line C in Figure 49b may be made thinner than the other areas.The flapper tip may be shaped to be easily elastically deformed so that it can tightly contact the inclined valve seat of the annular nozzle when pressed against the annular nozzle. This measure may be combined with solutions (1) to (3) above. Furthermore, if the width of the flapper tip is increased so that it can tightly contact the outer valve seat tip of the annular nozzle, the moment of inertia J of the moving part around the center of rotation (853a in Figure 46a) increases. The torsional spring stiffness is K. θ If so, K θThe primary resonance frequency (f0) of the servo valve decreases, proportional to the square root of the vibration energy (J). This decrease in resonance frequency (f0) degrades the responsiveness of the servo valve. Furthermore, as explained in Supplementary Note [3], this not only significantly impacts the performance of the active vibration isolation system, but also causes noise generation in higher-order resonance modes. To address this issue, constructing the flapper 853 (moving body) using a material with high vibration damping capacity is an effective solution. Generally, metallic materials have lower damping capacity than rubber or synthetic resins, but they also possess many excellent properties, such as toughness, heat resistance, and wear resistance. Alloys have been developed that combine the excellent properties of metallic materials with high damping capacity. Examples include Co-Ni alloys, Mg-Zr alloys, and Mn-Cu alloys. Vibration-damping alloys are classified according to the factors or mechanisms that contribute to their damping capacity, such as composite, ferromagnetic, transition, and twin crystal types. Using these high-damping materials to construct the flapper 853 can reduce the impact on the equipment in which the servo valve is installed. The annular nozzle applied to the oscillating servo valve may not be composed of multiple components as in this embodiment, but may be constructed as a single component using a manufacturing method such as wire-cut electrical discharge machining, as shown in the fifth embodiment. In this embodiment using an oscillating servo valve, an annular nozzle is applied to both the intake and exhaust sides. This also applies to the direct-acting servo valve described above. Depending on the specifications required by the pneumatic servo system, for example, an annular nozzle may be provided only on the intake side, and a single-hole nozzle may be used on the exhaust side. In this case, the exhaust nozzle inner diameter d may be set within a range that does not cause the "flapper adsorption phenomenon to the exhaust nozzle" described in Section [3-2-1]. The above describes a method for applying the present invention to a conventional oscillating servo valve composed of a permanent magnet and an electromagnetic coil. The above method can also be applied to the magnetic attraction type valve structure described in the first embodiment, which does not use a permanent magnet and in which the movable body (flapper) is fixed at one end and oscillates. That is, the outer valve seat tip and the inner valve seat tip of the annular nozzle may be configured to be in close contact with the movable body when the movable body is tilted at the maximum angle θ (not shown).

[0065] Seventh Embodiment Figure 50 shows a seventh embodiment of the pneumatic servo valve (fluid control valve) of the present invention, in which the aforementioned annular nozzle is applied to an oscillating servo valve and a bypass flow path is formed only on the intake side. Figure 50a is a front cross-sectional view of the servo valve, Figure 50b is a view taken along the arrow AA in Figure 50a, and Figure 50c is a diagram showing the shape of the flapper tip, which is the opposing surface of the annular nozzle. The structure of the magnet assembly 851, which forms a magnetic circuit including a permanent magnet, is the same as that of the sixth embodiment. (1) When a bypass flow path is formed only on the intake side. In Figure 50a, 901 denotes a housing, and 902 denotes an intake side (supply side) nozzle section (first nozzle section), which is a forward-direction nozzle installed through the housing. As described in the first embodiment, the annular nozzle is formed by combining two parts (an inner member 903a and an outer member 903b). 904 denotes a flow path conversion member directly connected to the inner member 903a. Figure 50b is a cross-sectional view of the flow path conversion member 904 (cross-sectional view taken along the line AA in Figure 50a), showing intake flow path A 905a, 905a, 905a, and 905a formed axially. Reference numeral 906a denotes an axial bypass flow path, and 906b denotes a radial bypass flow path. In the inner member 903a, an intake flow path B (shown by a chain line 907) is also formed to connect with intake flow path A formed in the flow path conversion member. Reference numeral 908 denotes an axial bypass flow path formed in the inner member 903a, and 909 denotes an intake-side annular flow path. Reference numeral 910 denotes an exhaust-side (discharge-side) nozzle section (second nozzle section) that is a reverse-direction nozzle attached to the housing. Similar to the intake-side nozzle section, this annular nozzle is formed by combining two parts (an inner member 911a and an outer member 911b). Reference numeral 912 denotes an exhaust flow path (shown by a chain line) formed in the inner member 911a. Reference numeral 913 denotes an inner member 903a of the intake side nozzle portion and a bypass through-hole formed at the center of the tip of the flapper facing the inner member 911a of the exhaust side nozzle portion, and 914 denotes an exhaust side annular flow passage. 915 denotes an intake hole, 916 denotes an exhaust port, and 917 denotes a control port which is connected to a pneumatic actuator (not shown). 918 denotes a control chamber of this servo valve. The supply pressure P SThe gas flows into the control chamber 918 via the intake flow path A formed in the flow path conversion member 904, the intake flow path B formed in the inner member 903a, and the intake-side annular flow path 909. At the same time, the gas in the control chamber 918 is discharged to the atmosphere via the exhaust-side nozzle portion 910. 919 is a bypass flow path formed inside the housing 901, and 920 is a first opening of the bypass flow path, which is the flapper-side opening end of the inner member 903a. 921 is a second opening formed on the housing side of this bypass flow path. The fluid flowing through the bypass flow path takes a detour path in the following order: first opening 920, axial bypass flow path 908 formed in the inner member 903a, axial bypass flow path 906a formed in the flow path conversion member 904, radial bypass flow path 906b, bypass flow path 919 formed inside the housing 901, second opening 921, and control chamber 918. In this embodiment, the bypass flow path is formed only on the intake side. However, by forming a bypass through-hole 913 in the flapper tip 860, not only the intake flow rate but also the exhaust flow rate can be increased simultaneously, similar to the two bypass effects in the previously described embodiment. (2) When Bypass Flow paths are formed on both the intake side and the exhaust side. Figure 51 is a front cross-sectional view of the servo valve of this embodiment when bypass flow paths are formed on both the intake side and the exhaust side. 922 is an exhaust-side flow path conversion member attached to the exhaust side and directly connected to the inner member 911a. 923 is an axial bypass flow path formed in the inner member 911a, and 924 is a bypass flow path formed axially and radially inside the exhaust-side flow path conversion member. 925 is an exhaust-side bypass flow path formed inside the housing 901. In this embodiment, where bypass flow paths are formed on both the intake side and the exhaust side, both the intake flow rate and the exhaust flow rate can be increased without forming a bypass through-hole in the flapper tip 926. To address the issue of applying an annular nozzle to an oscillating motion servo valve, namely, to prevent leakage due to the occurrence of a gap between the flapper and the nozzle, the solutions (1) to (3) shown in the sixth embodiment can be applied.The difference between the servo valve of this embodiment and the magnetic attraction type embodiment in which the flapper moves in parallel can be summarized as follows: this is a fluid servo valve in which the flapper drive means (magnet assembly 851) is located away from the axis of the intake-side flow path that houses the intake-side nozzle portion and the exhaust-side nozzle portion, and the exhaust-side flow path. The intake-side flow path and the exhaust-side flow path are formed through a housing 901, and a flow path connecting the first opening and the second opening of the bypass flow path is formed in the housing. The features of this valve structure can be summarized as follows: a cylindrical flow path is formed inside the housing, and the intake-side nozzle portion and the exhaust-side nozzle portion are formed within this cylindrical flow path. Furthermore, the flapper drive means (actuator) is located away from the axis of each nozzle portion. Bypass flow paths can be freely formed on both the intake side and the exhaust side using this housing. As shown in the second embodiment, an annular nozzle with a bypass passage applied to an oscillating servo valve may have a structure in which two passages, namely, (1) an intake-side passage and (2) a bypass passage, are incorporated into one nozzle unit. In this case, as with a conventional single-hole nozzle, the work of attaching the nozzle unit to the valve body can be simplified, and the valve body can be simplified.

[0066] Eighth Embodiment In the previous embodiments, the flapper was moved by the Maxwell attraction stress generated between the electromagnet and the flapper. In this embodiment, a linear motor (voice coil motor) operating on the Lorentz force is used as the flapper moving mechanism, and annular nozzles are used for the intake and exhaust nozzles. Figure 52 shows a linear motor-type pneumatic servo valve (fluid control valve) according to the eighth embodiment of the present invention. Figure 52a is a front cross-sectional view, and Figure 52b is a view taken along the arrow AA in Figure 52a. Reference numeral 350 denotes a cylindrical central shaft (support shaft), 351 denotes an outer frame formed concentrically with the central shaft, 352 denotes a permanent magnet attached to the outer frame, 353 denotes a flapper support member, 354 denotes a cylindrical exhaust-side housing that accommodates the bottom of the central shaft and the outer frame 351, 355 denotes the bottom of the exhaust-side housing, and 356 denotes an exhaust hole formed in the bottom of the exhaust-side housing. Reference numeral 357 denotes a flapper (movable body), 358 denotes a bolt fastening the flapper to the flapper support member 353, 359 denotes a coil bobbin attached to the flapper, and 360 denotes a coil wound around the coil bobbin. Reference numeral 361a denotes a gap between the coil and the permanent magnet, and 361b denotes a gap between the inner circumferential surface of the coil bobbin and the outer circumferential surface of the central shaft 350. Reference numeral 362 denotes a flow hole formed in the flapper, 363 denotes an exhaust-side flow passage C formed through the central shaft 350, and 364 denotes a reverse-direction nozzle attachment portion formed on the flapper side of the exhaust-side flow passage. In this embodiment, the reverse-direction nozzle uses the exhaust-side annular nozzle 171 ( FIG. 5 ) composed of the annular flow passage 317 used in the first embodiment and a flow passage conversion portion (cylindrical portion 313). This exhaust-side annular nozzle becomes the exhaust-side flow passage B. Reference numeral 365 denotes a flapper mounting portion formed on the flapper side of coil bobbin 359, and 367 denotes a bolt that fastens this flapper mounting portion to flapper 357. Reference numeral 368 denotes exhaust-side flow passage A formed by penetrating the flapper mounting portion and the center of the flapper. An exhaust flow passage that connects the control chamber (described below) and exhaust hole 356 is formed by exhaust-side flow passage A 368 → exhaust-side flow passage B 171 (exhaust-side annular nozzle) → exhaust-side flow passage C 363.Reference numeral 369 denotes a coil bobbin internal space formed by the inner peripheral surface of the coil bobbin 359 and the outer peripheral surface of the central shaft 350. Reference numeral 370 denotes a through-hole formed in the flapper mounting portion and the flapper, which connects the coil bobbin internal space to a control chamber (described below). The coil bobbin 359, coil 360, permanent magnet 352, outer frame 351, flapper 357, and flapper support member 353 constitute a linear motor (voice coil motor) that controls the displacement of the flapper 357 by applying current to the coil 360. The principle of this linear motor utilizes the Lorentz force acting on the current-carrying coil 360 placed in a magnetic field. Reference numeral 372 denotes an intake-side housing, 373 denotes a control-side flow path connected to a pneumatic actuator (not shown), and 373a denotes a control port. Reference numeral 374 denotes an intake hole, 375 denotes an intake-side gap formed between the intake-side housing 372 and the flapper 357, and 376 denotes an exhaust-side gap formed between the flapper 357 and the flapper support member, the permanent magnet, the coil bobbin, etc. Reference numeral 377 denotes a recess formed on the flapper side of the intake-side housing for mounting a forward-direction nozzle. In this embodiment, the forward-direction nozzle is the bypass-flow-path-integrated annular nozzle 651 (first nozzle), which is the intake-side (supply-side) nozzle portion used in the second embodiment ( Figures 32 to 34 ). The exhaust-side (discharge-side) nozzle portion 171 (second nozzle portion), which is the reverse-direction nozzle, has the same structure as in the first embodiment. Figure 51b shows a cross-sectional view of the bypass-flow-path-integrated annular nozzle 651 as viewed along the arrow AA in Figure 51a. Reference numeral 378 denotes a bolt fastening the intake-side housing and the exhaust-side housing. The intake side gap and the exhaust side gap form a control chamber 379 of this valve. This servo valve is a two-way, three-way valve whose flow path is composed of two-way nozzles (forward nozzle 651, reverse nozzle 171) and three openings (intake hole 374, exhaust hole 356, control port 373a). The linear motor servo valve described above, as well as the embodiments described below, has an issue with the above-mentioned necessary condition required for an active control vibration isolation table, namely, that "the primary resonance point of the moving part (valve element) must be sufficiently high, at least several hundred Hz (at least 200 Hz)." The reason for this is the mass of the moving part of the linear motor.This is because the moving part mass m of the linear motor, consisting of the coil and the coil bobbin, is inevitably larger than that of the magnetic attraction type servo valve of the above-described embodiment of the present invention and the conventional oscillating type servo valve. The present invention is applicable to applications where the primary resonant frequency condition does not need to be considered and where (i) the linear motor type requires high power drive and (ii) costs are higher than those of the nozzle flapper type. Furthermore, the advantages of the linear motor type can be utilized in applications requiring a larger flapper stroke. While this embodiment uses a moving coil type in which a permanent magnet is located on the fixed side and a coil on the moving side, a moving magnet type in which a permanent magnet is located on the moving side and a coil on the fixed side may also be used. While the above-described embodiments of the present invention have been described using a magnetic attraction type actuator as the driving means acting on the flapper, similar effects can be achieved using the linear motor of this embodiment (not shown).

[0067] Ninth Embodiment In this embodiment, similar to the above-described embodiments, a linear motor (voice coil motor) operating on the Lorentz force is used as the flapper moving means, and a throttled intake valve and an annular nozzle are combined to form a two-way four-way valve and a one-way three-way valve. First, an example of application to a two-way four-way valve will be described. [1] Application Example to a Two-Way Four-Way Valve (1) Specific Structure of a Two-Way Four-Way Valve Unlike the configuration of the above-described two-way three-way valve, the servo valve of this embodiment is a two-way (bidirectional) four-way valve whose flow path is composed of two-way nozzles (forward nozzle and reverse nozzle) and four openings (intake port, exhaust port, and two control ports). The two-way four-way valve is useful, for example, in controlling the pressure and flow rate of the left and right air chambers sandwiching the piston in opposite phases in a pneumatic actuator composed of a piston, a rod directly connected to the piston, and a cylindrical cylinder that houses them. Figure 53 shows a linear motor-type pneumatic servo valve (fluid control valve) according to a ninth embodiment of the present invention. Figure 53a is a front cross-sectional view, and Figure 53b is a view taken along the arrow AA in Figure 53a. Reference numeral 450 denotes a cylindrical central moving shaft, 451 denotes an outer frame portion formed on the fixed side and concentric with the axis of the central moving shaft, 452 denotes a permanent magnet attached to the outer frame portion, 453L and 453R denote flapper support members, 454 denotes a cylindrical intermediate housing, 455L denotes a left flapper, 455R denotes a right flapper, and 456 denotes a bolt fastening these flappers to the flapper support member. 457 denotes a coil bobbin attached to the left flapper 455L, and 458 denotes a coil wound around the coil bobbin. 459 denotes an R-side support member fixed to the center of the right flapper 455R, and 460 denotes an L-side support portion formed in the center of the coil bobbin. The R-side support member and the L-side support portion fix the left and right ends of the central moving shaft 450. In the two-way four-way valve of this embodiment, the left and right fluid circuits from intake to exhaust can be considered to be "mirror symmetrical," and the components formed and attached to the left and right intake housings have the same shape, so only the structure of the L side will be described.Reference numeral 461L denotes the L-side intake housing, 462 denotes the bolts fastening this L-side intake housing to the intermediate housing 454, and 463L denotes the L-side intake plate, which is fastened to the L-side intake housing with bolts 464. Reference numeral 651L denotes the annular intake nozzle attached to the L side, and uses the same structure as the bypass flow path-integrated annular nozzle 651 (Figures 32 to 34) used in the second embodiment. Figure 53a shows the cross-sectional view taken along the line B-B in Figure 53b. Reference numeral 464L denotes the L-side spacer, which is attached between the L-side intake plate and the L-side intake housing. An L-side intake hole (supply hole) 465L with an L-side restriction is formed in the center of the spacer. Reference numeral 466L denotes an L-side control chamber (L-side intermediate control chamber) formed between the L-side spacer and the annular intake nozzle 651L, 467L an L-side control flow path connecting this L-side control chamber with the L-side control port 468L, 469L an L-side intake opening, 470L an L-side exhaust hole (discharge hole), 471L an L-side exhaust chamber (discharge chamber) formed between the L-side intake housing and the right-side flapper, and 472L a bypass flow path opening. Figure 54a shows the air flow from the air supply source to the exhaust hole 470L when the annular intake nozzle 651L (forward direction nozzle) is shown in the B-B cross section of Figure 53b. Specifically, the flow path is "air supply source → L-side intake opening 469L → L-side throttled intake hole 465L → L-side intermediate control chamber 466L → intake-side flow path 667 (four through-holes) → annular flow path 668 → L-side exhaust chamber 471L → L-side exhaust hole 470L." Figure 54b shows the air flow in the bypass flow path when the annular intake nozzle 651L is shown in the CC cross section of Figure 53b. That is, the fluid flowing inward from the annular flow path 668 reaches the L-side exhaust chamber 471L via the detour flow path of "bypass flow path opening 472L (first opening) → bypass flow path A 669 → bypass flow path B 670 → bypass flow path C 671 (second opening)." The flow path of the annular intake nozzle 651R (reverse direction nozzle) is similar to the above. In summary, the first opening of the bypass flow path is formed on the movable body side of the center of the inner valve seat, and the second opening is formed in the discharge chamber or in the flow path connecting the discharge chambers.The servo valve of this embodiment focuses on the fact that by configuring the two nozzles and two flappers symmetrically, it is possible to arrange the "throttle-equipped intake port + annular nozzle with bypass flow path" that requires a large installation space symmetrically and with ample space. (2) Operational Principle of the Two-Way Four-Way Valve In the two-way four-way valve of this embodiment, unlike the two-way three-way valve embodiment described above, the gap 471L downstream of the intake nozzle is maintained at atmospheric pressure. The pressure in the L-side control chamber (L-side intermediate control chamber) 466L connected to the L-side control port 468L is determined by the fixed air resistance R0 of the L-side throttle-equipped intake port 465L and the variable resistance R of the L-side annular intake nozzle 651L. LX The gap h between the tip of the L-side annular intake nozzle 651L and the left flapper 455L is determined by L When the gap h is large, the pressure in the L-side control chamber decreases. LWhen the pressure in the L-side control chamber is small, the pressure in the L-side control chamber increases. In this embodiment, the L-side and R-side servo valves are mirror-symmetric. Therefore, the outputs of the L-side control port 468L and the R-side control port 468R change in opposite phases. By connecting these two control ports to the left and right air chambers on either side of the piston, the displacement, speed, and generated force of the pneumatic actuator can be controlled in two directions. Alternatively, two independent actuators may be arranged in series and controlled to achieve mechanical equilibrium when stationary. In this embodiment, flappers are arranged in two locations, left and right, and move in the same direction. This configuration facilitates the installation of an annular intake nozzle with a bypass passage, which requires a large installation space. In this embodiment, an annular nozzle with a bypass passage is used. However, as shown in the fourth embodiment, the bypass passage can be omitted and the outer diameter of the annular nozzle can be further increased to compensate for insufficient flow rate. Furthermore, if a large flow rate is not required, as shown in the eighth embodiment, for example, the exhaust side bypass passage can be omitted, the outer diameter of the annular nozzle can be reduced, and the annular nozzle can be housed inside the coil bobbin 359 (Figure 52a). (3) Application to Magnetic Attraction Type: The actuator driving the two-way four-way valve is not limited to a linear motor and may be a magnetic attraction type. For example, to convert the second embodiment (FIG. 32) into a two-way four-way valve, a first intake throttle is incorporated into the upper part of the annular nozzle 651 with a bypass flow path, a spacer with a first intake hole is attached, and the intake-side passage A 666 becomes the first control chamber. The control-side passage 161 is shielded, and a second intake throttle is incorporated into the lower end of the exhaust hole 157, and a spacer with a second intake hole is attached. The interior of the exhaust-side flow passage 158 becomes the second control chamber. Openings connecting the first and second control chambers are formed to form two control ports (not shown), as in the ninth embodiment. Constructing a two-way four-way valve using the magnetic attraction type eliminates the primary resonance frequency issue of the linear motor servo valve described above. [2] Application to one-way three-way valves (1) Specific structure of one-way three-way valves In the two-way four-way valve embodiment described above (Figure 53), by taking advantage of the fact that the intake and exhaust nozzle structures are symmetrical, it can also be easily applied to one-way three-way valves (fluid control valves).That is, the one-way three-way valve shown in Figure 55a uses the structure of the two-way four-way valve described above, which has annular nozzles attached to the left and right, but attaches an annular nozzle to only one side and makes the other side a sealed structure. Figure 55b shows the electrical equivalent circuit of the one-way three-way valve, P. S is the supply pressure, R0 is the fixed air resistance of the supply hole with a throttle, R X is a variable resistor, which is an annular nozzle. In Figure 55a, the linear motors driving the flappers 455L and 455R and the coil bobbin 457 are configured in the same manner as in the two-way four-way valve. The configurations of components such as the control port 468, the throttled intake port (supply port) 465, the exhaust port (discharge port) 470, and the bypass flow path-integrated annular nozzle 651, as well as the fact that the left and right exhaust chambers (discharge chambers) 471L and 471R are all at atmospheric pressure, are also similar to the two-way four-way valve. However, components used only in one-way three-way valves are described as, for example, "L-side throttled intake port 465L → throttled intake port 465." 473 is a shielding plate for sealing the R-side exhaust chamber 471R, and 474 is a fastening bolt. The air flow direction shown in Figure 55a is the same as the L side of a two-way four-way valve: air supply source → intake opening 469 → throttled intake hole 465 → control chamber (intermediate control chamber) 466 → intake-side flow path 667 (four through-holes) → annular flow path 668 → exhaust chamber 471L → exhaust hole 470. The air flow direction may be reversed by connecting the exhaust hole 470 to an air supply source (high-pressure source) and opening the intake opening 469 to the atmosphere. Even in this case, the fluid pressure in the control chamber 466 or the flow rate into and out of the control chamber can be controlled by changing the distance between the tip of the annular nozzle and the flapper 455L. (2) Effect of Applying an Annular Nozzle The effect of applying an annular nozzle to a one-way three-way valve will be explained using the equivalent electrical circuit in Figure 55b. Supply source pressure P S , fixed air resistance R0 of the throttling supply hole, variable air resistance R of the intake nozzle X As a result, the valve flow rate Q=P S / (R0+R X ) The variable air resistance is R Xmin ≦R X The gap between the tip of the intake nozzle and the flapper is maximum h = h max In the case of R X =R XminWhen the tip of the intake nozzle is blocked by the flapper, that is, when h=0, R X The magnitude of the fixed air resistance R0 of the intake port 465 with a throttle is usually determined when the gap is at the intermediate position h=h max The value of the variable resistor R when it is at / 2 X =R Xmid In this case, the valve flow rate Q mid =P S / (R0+R Xmid )= P S / 2R Xmid That is, the air resistance R X =R Xmid The smaller the value, the larger the valve flow rate. Here, the annular nozzle has a larger valve flow rate than the single-hole nozzle at the same source pressure P S It is noteworthy that the air resistance can be significantly reduced under the same gap h. In other words, the effect of the annular nozzle, which can increase the flow rate, also significantly reduces the nozzle's fluid resistance. By using an annular nozzle, if the flow rate increase ratio relative to a single-hole nozzle is η, the air resistance passing through the nozzle can be set to be small at the same ratio η. For example, in the first embodiment, using graphs showing the relationship between flow rate and nozzle inner diameter (FIGS. 12 and 18), we compare the flow rates when the nozzle inner diameter d (or outer nozzle inner diameter d0) is 3 mm and the stiffness ratio K / K0 = 1. Under these conditions, the flow rate of the single-hole nozzle is Q = 50 L / min, the annular nozzle without a bypass flow path is Q = 72.3 L / min, and the annular nozzle with a bypass flow path is Q = 129 L / min. In other words, compared to when a single-hole nozzle is used as a variable resistor, the flow rate increase ratio when an annular nozzle without a bypass flow path is used as a variable resistor is η = 1.45, and when an annular nozzle with a bypass flow path is η = 2.58. In other words, compared to when a single-hole nozzle is used for the variable resistance, when an annular nozzle without a bypass passage is used for the variable resistance, the variable resistance and fixed air resistance R0 of the nozzle can be reduced by 1 / 1.45 = 0.69 times, and when an annular nozzle with a bypass passage is used, the variable resistance and fixed air resistance R0 of the nozzle can be reduced by 1 / 2.58 = 0.39 times. In summary, the maximum flow rate Q of a one-way three-way valve (or a two-way four-way valve) maxBy applying an annular nozzle, the flow rate can be increased by 1.45 times without a bypass flow path, or 2.58 times with a bypass flow path. The weakness of the conventional one-way three-way valve (two-way four-way valve) that was configured with a single-hole nozzle was that the flow rate control range was narrow, but with this invention, a servo valve with a wide flow rate range can be realized while maintaining the basic configuration of the conventional one-way three-way valve. Rather than applying an annular nozzle, the air resistance R can be increased by increasing the inner diameter of the single-hole nozzle. XmidThe difference can be reduced. In this case, the effective displacement of the movable body due to the fail-safe mechanism decreases. This results in issues such as increased electromagnet force, increased power supply capacity, and a larger control valve body. To summarize the benefits of using an annular nozzle, a one-way, three-way valve or a two-way, four-way valve with a wide flow control range can be realized while maintaining a simple valve structure. (3) Other Forms of One-Way, Three-Way Valves: As with two-way, four-way valves, the actuator driving the one-way, three-way valve is not limited to a linear motor, but can also be a magnetic attraction type. For example, to modify the second embodiment (FIG. 32) into a one-way, three-way valve, an intake throttle can be incorporated into the upper part of the annular nozzle 651 with a bypass flow path, a spacer with an intake hole can be attached, and the intake-side passage A 666 can be used as a control chamber. Two sets of two independent one-way, three-way valves can also be combined to form a servo valve with the function of a two-way, four-way valve. Combining even more one-way, three-way valves can realize a fluid control system with a more complex flow path configuration. The effect of being able to control a large flow rate over a wide flow rate range is similar to that of the previously described two-way four-way valve. Using the structure of the one-way three-way valve shown in Figure 55a, a one-way two-way valve designed solely for flow rate control can be easily constructed. To achieve this, simply remove the spacer 464 with the throttle-equipped intake hole 465 and block the control flow path 467. In this case, the intake opening 469L may serve as an exhaust hole, and the exhaust hole 470 may serve as an intake hole. (3) One-way Three-way Valve with Rectifier: Combining an Annular Nozzle and a Rectifier. Figure 56 shows the previously described one-way three-way valve in which the single throttle-equipped intake hole (supply hole) 465L is replaced with a rectifier. This rectifier is provided to suppress turbulence, vortices, and shock waves generated by restricting the gas flow through an orifice, i.e., to smooth out turbulent gas flow. In this embodiment, the rectifier has a large outer diameter D and a narrow gap δ. sHereinafter, as in the above-described embodiment, parts used only in one-way three-way valves will be described as, for example, "L-side control port 468L → control port 468." Reference numeral 481 denotes an outer member fastened to the intake housing 461 with bolts 482, and 483 denotes a through passage formed in the center of this outer member. Reference numeral 484 denotes an intake side member fastened to the end face of this outer member with bolts 485, 485a denotes a cylindrical portion constituting this intake side member, and 485b denotes a cylindrical narrow-diameter portion. The cylindrical portion 485a is housed within the through passage, and a narrow gap δ is formed between the outer peripheral surface of the cylindrical portion and the inner peripheral surface of the through passage. s The annular flow passage 486 for rectifying the flow has a gap 487 formed between the outer peripheral surface of the cylindrical small diameter portion 485b and the inner peripheral surface of the through passage. The intake spacer 488 is screwed to the end face of the intake side member, the intake hole 488a is formed in the center of the intake spacer, and the flow holes 489a, 489b, 489c, and 489d are formed penetrating the intake side member in the axial direction (489b and 489d are not shown). That is, the gas flows into the control chamber (intermediate control chamber) 466 via the intake hole 488a → flow hole 489 → gap 487 → annular flow passage 486. The gap δ of the annular flow passage 486 for rectifying the flow s is the inner diameter d of the intake hole 465L with a restriction in the embodiment described above (FIG. 54). f Furthermore, the annular flow passage gap δ of the annular intake nozzle 651 r That is, δ s ≪d f , and δ s ≪δ r Incidentally, the annular flow passage gap δ of the annular intake nozzle 651 r is the inner diameter of the outer nozzle d0, and the outer diameter of the inner nozzle d i As δ r =(d0-d i) / 2. As is well known, whether a gas flow is in the laminar flow region or the turbulent flow region is evaluated by the Reynolds number. In fluid mechanics, the Reynolds number is a dimensionless quantity defined as the ratio of inertial force to viscous force. When the Reynolds number is low, the flow becomes laminar, where viscous force is dominant and the flow is smooth and stable. When the Reynolds number is high, the flow becomes turbulent, where inertial force is dominant and the flow becomes disorderly vortexes and unstable. If the kinematic viscosity (dynamic viscosity coefficient) is v, the characteristic length (gap) of the flow path is δ, and the flow velocity of the fluid is V, the Reynolds number is Re=Vδ / ν. In this embodiment, the outer diameter of the flow-straightening annular flow path 486 is D, the gap δ s Then, the opening area is A s =2πDδ s On the other hand, the inner diameter d f The opening area of ​​the intake hole with throttle 465L is A f =(d f / 2) 2 π. Therefore, A s =A f For example, in the above-described embodiment, if the inner diameter d of the throttle-equipped intake hole 465 is set to f = 2 mm, and the outer diameter of the annular flow path 486 is set to D = 6.2 mm, the same opening area (A s =A f =3.14mm 2 ) gap δ s =0.08 mm. Under the same supply pressure conditions, the Reynolds number of gas passing through the flow-straightening annular flow passage 486 of this embodiment is approximately 1 / 40 of that of gas passing through the throttle-equipped intake port 465. As a result, the control chamber 466 connected to the control port 468 can maintain a stable, noise-free control pressure. Therefore, the instability of the gas flow caused by the increased flow rate due to the use of an annular intake nozzle can be avoided by providing a flow straitener that also functions as a throttle upstream of the annular intake nozzle. The flow straitener structure is not limited to this embodiment; for example, a structure in which multiple thin tubes are combined in a honeycomb pattern or a structure in which gas flows through narrow radial gaps can be applied. As mentioned above, the one-way, three-way valve of this embodiment can easily be adapted to a two-way, four-way valve with symmetrical intake and exhaust nozzle structures.

[0068] Tenth Embodiment Figure 57 shows a fluid control device according to a tenth embodiment of the present invention, illustrating the overall configuration of an active vibration isolation table when a vibration suppression valve using an annular nozzle according to the present invention is mounted on a separate pneumatic servo system. Figure 58a is a model diagram of Unit (A), one of four actuator units. Figure 58b is a model diagram of the vibration suppression servo valve attached to Unit (A). Figure 58c is a graph of the actuator's generated force versus time. As mentioned above, the inventors have proposed and are currently applying for a pneumatic servo system in which vibration suppression actuators and vibration suppression actuators are arranged separately and independently. In this case, the vibration suppression valve that draws and exhausts gas inside the vibration suppression actuator is also used separately and independently from the vibration suppression valve. In a manufacturing process in which vibration suppression control and vibration isolation control are alternately repeated, separately driving the vibration suppression and vibration suppression actuators and the vibration suppression and vibration suppression servo valves eliminates the trade-off between vibration suppression and vibration suppression performance and allows the selection of a configuration that achieves the best performance for each. Figure 57 shows a fluid control system (active vibration isolation table) equipped with a fluid control valve using an annular nozzle according to the present invention. It is a model diagram of a vertical actuator unit equipped with feedforward vibration control (FF control). In the actuator unit, vibration isolation and vibration control actuators are arranged coaxially, and vibration isolation and vibration control servo valves that draw and exhaust gas from these actuators are arranged separately and independently. 1. Overall Configuration of the Active Vibration Isolation Table. In the figure, four sets of actuator units [Unit (A) to Unit (D)] are arranged on floor 253 to support surface plate 254 (vibration isolation table), on which precision equipment is mounted. These actuator units have vibration control and vibration isolation actuators arranged coaxially [Unit (B) and Unit (C) are not shown]. Reference numeral 255 denotes a horizontal actuator unit (details not shown). Below, the configuration of Unit (A) is described, but Units (B) to (D) have a similar configuration. Reference numeral 256a denotes a surface plate-side intermediate support member, and 257a denotes a floor-side support member having a U-shaped cross section. Reference numeral 258A denotes the floor side of the floor-side support member 257a, and 258B denotes the surface plate side of the floor-side support member 257a.Reference numerals 259aA and 259aB denote vibration-damping actuators, sandwiched between the base-side intermediate support member 256a and the floor-side support member 257a and arranged facing each other. Reference numerals 260aA and 260aB denote vibration-damping valves installed on the floor side 258A and base side 258B of the floor-side support member. Reference numerals 261aA and 261aB denote flow paths connecting the vibration-damping valves and the vibration-damping actuators. These flow paths are roughly L-shaped, which is the shortest flow path shape in the structure of this embodiment, in which the vibration-damping and vibration-isolating actuators are arranged coaxially. Sensor 264a is representative of the acceleration detector and displacement detector that detect signals in the Z-axis direction of Unit (A). Reference numeral 262a denotes a vibration-isolating actuator (vibration-isolating means), sandwiched between the base-side support member 263a and the floor-side support member 257a. The vibration isolation actuator in this embodiment suppresses vibration transmission between the floor (foundation) and the base plate (vibration isolation table) and also serves as a support means for the base plate relative to the floor. Reference numeral 265a denotes a vibration isolation valve installed on the base plate support member 263a, and reference numeral 266a denotes a flow path connecting the vibration isolation valve and the vibration isolation actuator. 2. Vibration Isolation Servo Valve and Vibration Isolation Actuator. Figure 58a is a model diagram of Unit (A), one of four actuator units, and Figure 58b is a model diagram of the vibration isolation servo valve 260aB attached to this Unit (A). The vibration isolation servo valve 260aB shown is the one used in the first embodiment (Figure 2), but an appropriate structure can be selected to match the specifications required for the active vibration isolation table. For example, the oscillating servo valve used in the sixth embodiment or the linear motor servo valve used in the eighth embodiment may also be used. Figure 58c is a graph of the actuator force versus time. In this figure, F. Z1 is the force generated when the lower actuator 259aA is driven, F Z2 is the force generated when the upper actuator 259aB is driven. The sum of the two forces F Z =F Z1 -F Z2This acts as a Z-axis force on the surface plate 254, which is directly connected to the surface plate-side intermediate support member 256a. The magnetic attraction valves of the first to fifth embodiments, which control the nozzle opening using the elastic deformation of a thin disk (e.g., 162 in Figure 58b), can reduce the effective mass of the moving part (thin disk) compared to conventional valves. Because of their high resonant frequency f0 and high responsiveness, pneumatic servo systems with superior vibration suppression performance compared to conventional valves can be realized. The resonant frequency f0 of a servo valve is one of the important responsiveness evaluation indexes for vibration suppression valves in high-acceleration and vibration suppression control. As mentioned above, in this embodiment, the flow path connecting the vibration suppression valve 260aB and the vibration suppression actuator 259aB is formed with the shortest, approximately L-shaped flow path shape 261aB. The same applies to the straight flow path 178d between the servo valve 207d and the actuator 206d in the first embodiment. That is, as a pneumatic servo system, by simultaneously satisfying (i) high responsiveness of the servo valve itself and (ii) minimal flow resistance connecting the servo valve and the pneumatic actuator, the characteristics of the valve of the present invention, which combines high flow rate and high responsiveness, can be utilized. 3. Benefits of Applying an Annular Nozzle Valve (1) A highly responsive pneumatic servo system can be realized. In conventional nozzle flapper valves, increasing the intake nozzle diameter to increase the intake flow rate limits the increase in intake flow rate because the effective displacement of the movable body due to the fail-safe mechanism decreases. Increasing the support rigidity of the movable body reduces the fail-safe displacement and increases the effective displacement, thereby increasing the intake flow rate. However, increasing the rigidity of the movable body requires increasing the electromagnetic force, which leads to a problem of degraded responsiveness due to reduced electrical and mechanical time constants. In the case of a valve using an annular nozzle of the present invention, the fail-safe displacement can be reduced compared to a conventional single-hole nozzle under the same nozzle outer diameter, allowing the effective stroke of the movable body to be increased, significantly increasing the intake flow rate. Furthermore, the electrical and mechanical time constants can be maintained at low values ​​equivalent to those of a low flow rate valve, making it possible to realize a pneumatic servo system with high responsiveness.(2) Achieves optimal vibration suppression and vibration isolation performance. Conventional servo valves that combine vibration suppression and vibration isolation have a problem: increasing the flow rate of the servo valve to improve vibration suppression performance results in a decrease in vibration suppression performance. This is because the gradient of the control flow rate relative to the valve input current increases as the flow rate increases, reducing the resolution of the servo valve. In other words, vibration suppression and vibration isolation performance are inherently in a trade-off relationship. Separating the actuators and servo valves for vibration suppression and vibration isolation allows the actuators and servo valves to select the drive principle and structure that best achieves their respective performance. In other words, increasing the flow rate of the servo valve for vibration suppression control using an annular nozzle does not affect vibration suppression performance. (3) Resolves the issue of air consumption during steady state operation. As mentioned above, applying pneumatic servos to vibration suppression control has the problem of high air consumption and poor responsiveness. A high flow rate servo valve is essential for high acceleration and vibration suppression control. In conventional pneumatic servos, increasing the servo valve's flow rate to improve responsiveness inevitably leads to increased steady-state air consumption. In this servo system, which separates the vibration isolation actuator and the pneumatic vibration control actuator, the high flow rate through the vibration control servo valve is only momentary (several millimeters to several tens of milliseconds) during vibration control. Therefore, the average air consumption is not an issue. In other words, the combination of a separate pneumatic servo and the annular nozzle valve of the present invention fully utilizes the advantages of a servo valve with an annular nozzle, which can achieve a high flow rate compared to conventional servo valves with a single nozzle. (4) Excellent vibration isolation performance and large vibration control force are achieved. The air chamber of the vibration control actuator is set to open to the atmosphere when vibration control is completed. This setting eliminates the problem of increased resonance frequency due to the parallel connection of the air chamber springs of the vibration control and vibration control actuators when switching from vibration control to vibration isolation control. That is, by maintaining the spring stiffness of the vibration suppression actuator at zero or close to zero during the vibration suppression control stage, the presence of the vibration suppression actuator does not affect the vibration suppression performance.Furthermore, during vibration isolation control, external disturbance vibrations are prevented from propagating to the vibration isolation actuator and the surface plate via the air spring of the vibration isolation actuator. Furthermore, if an independently configured vibration isolation actuator is configured as an annular nozzle valve using a two-way, three-way valve, it can be driven over a wide pressure range from atmospheric pressure to near the maximum pressure (supply pressure), thereby generating a large force. To generate maximum force, the flapper and intake nozzle of the vibration isolation servo valve can be driven at full stroke from zero to maximum. Furthermore, by connecting the control port of the vibration isolation valve and the intake port of the vibration isolation actuator via the shortest path, i.e., by arranging the vibration isolation valve and the vibration isolation actuator so that they are connected by a roughly straight flow path (178a in Figure 1a) or a roughly L-shaped flow path (261aA in Figure 58a), the characteristics of the servo valve with an annular nozzle can be more effectively utilized.

[0069] [Correction pursuant to Rule 91 03.03.2025] <Eleventh Embodiment> Figure 59 shows the eleventh embodiment of the present invention, illustrating the configuration of an actuator unit when the present invention using an annular nozzle is applied as a vibration isolation valve to a separate-type pneumatic servo system. The overall configuration of the active vibration isolation table is similar to that of the tenth embodiment. Figure 59a is a model diagram of Unit (A), one of four actuator units, and Figure 59b is a model diagram of the vibration isolation servo valve 265a and vibration isolation actuator 262a attached to this Unit (A). The structure of the vibration isolation servo valve 265a is the same as the one-way, three-way valve shown in the ninth embodiment, which is configured with a throttled intake port and an annular nozzle. Combining two sets of this one-way, three-way valve forms a two-way, four-way valve. As mentioned in the ninth embodiment, a weakness of the conventional two-way, four-way valve configured with a single-hole nozzle is its narrow flow control range, and the same is true of the conventional one-way, three-way valve. In this embodiment, the use of the annular nozzle 651 reduces the fixed air resistance R0 of the throttle-equipped intake port 486, thereby enabling a wider flow control range. In an integrated pneumatic servo, the servo valve alternates between vibration isolation control and vibration suppression control, centered around an operating point where pressure is in equilibrium. During vibration isolation control, a small pressure control range is sufficient, while during vibration suppression control, a full-stroke pressure control range from atmospheric pressure to supply pressure is often required. However, in a separate pneumatic servo system, full-stroke pressure control from atmospheric pressure to supply pressure is performed by the vibration suppression servo valves (260aA, 260aB), so the pressure control range required for the vibration isolation servo valve 265a is narrow. Therefore, the narrow pressure control range of a one-way three-way valve is not a problem with a separate pneumatic servo. Furthermore, in this embodiment, the throttle-equipped intake port 486 is configured as a rectifying annular flow passage with a large outer diameter and a narrow gap. This annular flow path for straightening is provided to suppress turbulence, vortexes, and shock waves that occur when the gas flow is narrowed by the orifice, i.e., to smooth out turbulent gas flow, thereby reducing noise from the servo valve.In summary, the active vibration isolation table supported by four actuator units [Unit (A) to Unit (D)] can combine high acceleration / vibration suppression control with high resolution / vibration isolation control by applying the annular nozzle of the present invention to both the servo valves for vibration suppression control and vibration isolation control. Supplementary Note [1-1] Compliance action by combination of annular nozzle and thin diskDiagram 100 is a model diagram explaining the compliance action of the annular nozzle and thin disk, with Fig. 100a being a top view of the disk which is a flapper, Fig. 100b being a case where the tip of the valve seat of the annular nozzle is not inclined relative to the disk which is the opposing surface, and Fig. 100c being a side view when it is inclined at an angle θ. In section [3-3-2] of the first embodiment, it was explained that even if the valve seat tips 311c, 316c of the annular nozzle are slightly inclined relative to the disk surface 162c due to errors in machining accuracy and assembly accuracy, the elastic deformation portion 162ab in which the groove 162a is formed deforms three-dimensionally, thereby achieving an effect that the valve body central portion 162c, which is the surface opposing the annular nozzle, closely adheres to the shape of the valve seat tip...an effect that can be called a compliance effect. The radius of the approximate circle of the portion where no groove is formed (valve body central portion 162c) is defined as r (=D. r / 2). Figure 100b shows the case where the entire valve body central portion 162c moves in parallel by Δz in the Z-axis direction due to the axial force F applied to the disk. The axial force F is a distributed load applied to the disk, and represents a fail-safe force in the case of the intake nozzle, and a magnetic attraction force in the case of the exhaust nozzle. Stiffness K z The elastic energy (potential energy) required for the axial spring to deform (translate) by Δz in the Z-axis direction is E z Figure 100c shows a state in which the torsional torque T applied to the disk causes the center point 162d of the valve body central portion 162c to deform Δz in the Z-axis direction and tilt by an angle Δθ. The torsional torque T occurs when the valve seat tips 311c, 316c are slightly tilted relative to the disk surface 162c. The stiffness K θ The elastic energy required to tilt the angle Δθ by the torsion spring is E θ The elastic energy E θThe smaller E, the more easily the valve body central portion 162c, which is the surface facing the annular nozzle, is deformed in a tilted manner along the shape of the tip of the valve seat. z =E θ Let's say. Equation (11) shows the condition under which the center point 162d of the central part of the valve body is deformed in the axial direction by Δz and is deformed by an inclination of Δθ along the shape of the tip of the valve seat. The spring stiffness in the Z-axis direction, K, is set to satisfy the condition of equation (12). z , the spring stiffness in the torsional direction K θ If the torsional spring stiffness is set to a value less than the axial value, the central portion of the valve element is more likely to deform in the torsional direction than in the axial direction, and the sealing action of tightly sealing the tip of the valve seat can be reliably obtained. θX >K θY In the case of , a torsion spring stiffness with a large value is used, and K θ = K θX As a concrete example, K z =43.9N / mm=43.9×10 3 N / m, r=D r / 2=4.4mm=4.4×10 3 In the case of K, from equation (14) θX =K θX =2.27Nmm / rad=2.27×10 -3 If the through groove of the disk is curved, for example, by applying a spiral curve or an Archimedes curve, X θ Direction and Y θ Torsional spring stiffness in the direction K θX , K. θX is small, that is, its inverse, the compliance λ θX , λ θYThe seal size can be increased. Servo valves consisting of a thin disk (flapper) and a small, single-hole nozzle with a small inner diameter (d0 = 1.0 to 1.5 mm) are well known. However, the structure of this servo valve, which utilizes compliance to improve sealing performance for annular nozzles requiring large outer diameters (e.g., outer nozzle inner diameter d0 = 3 to 8 mm), is a new discovery. Flappers used in conventional oscillating servo valves are rigid and deform only in the same angular direction of the oscillating motion, but they do not achieve the compliance effect that requires three-dimensional deformation. Supplement [1-2] Other Annular Nozzle Shapes In the previously described embodiments, the inner and outer valve seats of the annular nozzle were all perfectly circular. However, the reason for using a perfect circular shape is solely from the perspective of productivity (ease of manufacturing). The function and effectiveness of the annular nozzle are not limited to a perfect circular shape; various shapes can be realized by selecting the appropriate production method. Figure 101 shows an annular nozzle with each valve seat configured as a square. Reference numeral 951 denotes an outer valve seat formed on the flapper side of the outer member, 952 an inner valve seat formed on the flapper side of the inner member, and 953 an annular flow passage formed by the wall surfaces of the outer valve seat and the inner valve seat. 951a denotes a seal portion at the tip of the outer valve seat, 952a denotes a seal portion at the tip of the inner valve seat, and 954 a bypass flow passage with a square cross section. As a method for roughly evaluating the performance of this annular nozzle, 955 is the inscribed circle of the inner valve seat, and 956 is the circumscribed circle of the inner valve seat. d i1 is the diameter of the circumscribing circle 956, d i2 is the diameter of the inscribed circle 955, and the approximate outer diameter of the inner nozzle d i ≒ (d i1 +d i2 ) / 2. In the same way, the approximate outer nozzle inner diameter d o≒. The annular nozzle shown in Figure 102 shows a case in which the annular flow passage is not formed as a continuous concentric circle but is formed by discontinuously dividing it in the circumferential direction. 961a, 961b, 961c, and 961d are annular flow passages divided into four in the circumferential direction, with 962a, 962b, 962c, and 962d being seal portions, 963 being a bypass flow passage, and 964 being a valve seat. The seal portion is formed to protrude toward the opposing flapper. As for other annular nozzle shapes, the shape of the circumferentially divided nozzles may be formed by a combination of straight lines such as a rhomboid, diamond, or kite shape rather than a curved shape. Furthermore, the circumferential division angle of each nozzle may be sufficiently small to form a nozzle group with many openings in the circumferential direction. The nozzle flow passage shape and the inner and outer valve seats can be any shape. For example, elliptical, rectangular, triangular, and other shapes can be used. Alternatively, the nozzle may have a structure in which multiple independent orifices, each having a rectangular long side in the radial direction and a rectangular short side in the circumferential direction, are arranged circumferentially. If an inscribed circle closest to the center and a circumscribed circle farthest from the center are drawn for these orifices, the combination of the inscribed and circumscribed circles forms an annular shape. Therefore, it is acceptable to call it an annular nozzle. The key point is that the nozzle, having a narrow slit (flow path), can adjust the flow rate by being relatively close to the flapper on the opposing surface. It is preferable to form a valve seat at the outer periphery of each orifice, which can be tightly sealed with the flapper. Figure 103 shows an annular nozzle with an annular flow path and bypass flow path that are formed with a multi-layer structure compared to the previously described embodiment. Reference numeral 971 denotes the outermost valve seat, 972 denotes a first intermediate valve seat, 973 denotes a second intermediate valve seat, and 974 denotes the innermost valve seat. Reference numeral 975 denotes a first annular passage formed by the wall surfaces of the outermost outer valve seat and the first intermediate valve seat, and 976 denotes a second annular passage formed by the second intermediate valve seat and the innermost inner valve seat. Reference numeral 977 denotes a first bypass passage formed by the wall surfaces of the first intermediate valve seat and the second intermediate valve seat, and 978 denotes a second bypass passage formed in the center of the innermost inner valve seat. If this annular nozzle with a multilayer structure is applied to an intake nozzle or an exhaust nozzle, the intake flow rate and exhaust flow rate can be further increased.The above describes non-circular annular nozzle shapes, circumferentially discontinuous annular flow paths, and multilayer structures consisting of annular flow paths and bypass flow paths. However, nozzle structures combining these structures are also possible. Supplementary Note [1-3]: Why a high resonant frequency is required for servo valves. The overall system response of an active vibration isolation system (fluid control device) driven by a pneumatic actuator is at most a few Hz to 10-odd Hz. Nevertheless, we will explain why a high resonant frequency of several hundred Hz is required for servo valves. Figure 104 shows an example of a control block diagram for an active vibration isolation system. Part A, indicated by the dotted line, is the controlled object, including the surface plate (581 in Figure 68). Figure 105 shows an example of open-loop transfer characteristics (Bode diagram) obtained from the control block diagram in Figure 104, showing the open-loop transfer function (G). L =X out / X in) versus frequency. (Figure 105a) and (Figure 105b) show the gain characteristics (Figure 105a) and phase characteristics (Figure 105b). As shown in the table in the graph, (1) curve P represents the case where acceleration feedback (hereinafter referred to as acceleration feedback) is not applied and the resonant frequency of the pneumatic servo valve is low, f0 = 100 Hz (point A in the graph). (2) curve Q represents the case where acceleration feedback is applied and the resonant frequency of the pneumatic servo valve is low, f0 = 100 Hz (point B in the graph). (3) curve R represents the case where acceleration feedback is applied and the pneumatic servo valve of the present invention is used, with a high resonant frequency, f0 = 1000 Hz (point C in the graph). The above (1), (2), and (3) are evaluated from the perspective of control stability. Incidentally, point D (5.5 Hz) in the graph is an eigenvalue determined by the spring stiffness of the pneumatic actuator and the controlled object, including the surface plate. As is well known, a system is stable if the following two points are satisfied on the Bode diagram of the open-loop transfer function: (i) There is a positive gain margin at the phase crossover point. (ii) There is a positive phase margin at the gain crossover point. In the case of (1) above, even when the resonant 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, applying acceleration feedback increases the gain and delays the phase by 180 degrees. Furthermore, at the resonant point of the servo valve, f0 = 100 Hz (point B), the gain margin is negative (gain > 0), making the system unstable. In the case of (3) above, applying acceleration feedback increases the gain and delays the phase by 180 degrees, just like in (2). However, at the resonant point of the servo valve of the present invention, f0 = 1000 Hz, the system gain is sufficiently reduced, and there is a sufficiently large gain margin (gain < 0), making the system stable. After extensive experiments, we found that if the resonant frequency of the moving part (flapper) is set to 200 Hz or higher, the acceleration feedback gain can be set to the minimum required level. However, a frequency of 300 Hz or higher is desirable. The high resonance frequency f0 of a servo valve is not only a measure of the vibration isolation characteristics described above, but also serves as an index for evaluating the responsiveness of a vibration control valve in high-acceleration vibration control. The magnetic attraction type valves of the first to fifth embodiments, which control the nozzle opening by utilizing the elastic deformation of a thin disk, can reduce the effective mass of the moving part compared to conventional valves, making it possible to realize an active vibration isolation table or pneumatic servo device with excellent vibration isolation and vibration control performance.Supplementary Note [1-4] When the Inner Member Through-passage is Used as the Main Control Passage In the first to third embodiments in which a bypass passage is formed in the annular nozzle, a control-side passage is provided in addition to the bypass passage as a passage connected to the air chamber of the pneumatic actuator. The following describes the benefits of omitting this control-side passage and using only the "inner member through-passage" formed in the center of the annular nozzle as the main control passage. [4-1] Structure with Independent Bypass and Control-Side Passages. For example, in the first embodiment (Figure 2), the fluid flowing from the intake-side annular nozzle 170 in the centrifugal direction from the outer valve seat 309 flows into the control chamber 176 and then into the control-side passage A 161. The fluid flowing centripetally from the inner valve seat 303 flows into the control-side passage B 169 via the bypass passage 302. In other words, the two fluids branching out of the annular nozzle merge again in the control-side passage B 169. The combined fluid flows through control-side flow path C 178d (shown in phantom lines) and connects to the air chamber of the vibration-damping pneumatic actuator 206d. As previously mentioned, this structure doubles the flow rate. However, depending on the environmental conditions of the valve, we discovered that the dynamic behavior of the flapper can become unstable during the initial stage of valve operation. This instability is accompanied by vibrations and abnormal noise, and significantly degrades the dynamic and static characteristics of the servo valve. [4-2] Structure with only the inner member through-passage as the main control flow path Figure 106 proposes a structure discovered as a result of investigating the cause of the above-mentioned instability, namely, a structure in which only the inner member through-passage as the main control flow path. In this figure, the intake side is similar in structure to the third embodiment (Figure 35) in which control-side flow path A 761 and control-side flow path B 769 are omitted, and therefore the same reference numerals are used. The exhaust side is similar in structure to the fourth embodiment (Figure 42) in which a single-hole nozzle is used, and therefore the same reference numerals are used. Reference numeral 770a denotes an intake side (supply side) nozzle section (first nozzle section) that is a forward direction nozzle attached to the center of intake side housing 759, and an annular nozzle is configured by combining an inner member 771 and an outer member 772. Reference numeral 771a denotes an inner valve seat, and 772a denotes an outer valve seat. Reference numeral 779A denotes an inner member through-passage formed in the center of inner member 771.This inner member through-passage is connected to the air chamber of an externally installed pneumatic actuator. 730 denotes a central axis, 731 denotes an exhaust nozzle attached to the flapper side of this central axis, and 732 denotes an exhaust flow passage formed through the central axis. In this embodiment, the exhaust nozzle is configured as a single-hole nozzle. 733A, 733B, 733C, and 733D denote through-passages formed in the flapper 762 (733B and 733D are not shown). [4-3] Effect of Using the Inner Member Through-passage as the Main Control Flow Passage: In this valve structure, the control flow passage provided in addition to the bypass flow passage in the previous embodiment is omitted, reducing the total intake flow rate of the valve by more than half. In exchange for this reduction in total flow rate, the following effect is achieved. In the configuration shown in Figure 2, the fluid ejected from the intake nozzle 170 flows centrifugal toward the control flow passage A 161, which is located away from the intake nozzle. As the fluid flows along the disk wall, the static pressure on the wall decreases by the amount of increase in dynamic pressure. This results in a pressure difference between the front and back of the elastic disk, causing coupled vibration between the fluid and the structure (disk). It was found that this coupled vibration is more likely to occur when the supply pressure is high, the flow rate is large, and the gap between the flapper and the nozzle is small. In the proposed valve, the fluid flows out of the outer valve seat 772a in the centrifugal direction via the annular flow path and instantly fills the control chamber 778. However, the centrifugal flow rate then decreases significantly compared to the valve shown in Figure 2. This is because the fluid flowing out to the control port connected to the pneumatic actuator is ...

Claims

1. A fluid control valve comprising: a movable body; a first nozzle portion and a second nozzle portion; a movable body support portion for fixing the movable body; and a drive means for the movable body, the drive means being configured so that the movable body approaches a tip portion of one of the first nozzle portion and the second nozzle portion and moves away from a tip portion of the other, and the fluid control valve controls a flow rate difference passing through the first nozzle portion and the second nozzle portion, or a flow rate passing through the first nozzle portion and a flow rate passing through the second nozzle portion, individually, by movement of the movable body, the first nozzle portion and the second nozzle portion being composed of an inner member and an outer member that houses the inner member, and comprising: an inner valve seat formed on the movable body side of the inner member; an outer valve seat formed on the movable body side of the outer member; an annular flow passage formed by peripheral walls of the outer valve seat and the inner valve seat; and a flow passage connecting the annular flow passage to a supply hole for a working fluid, or connecting the annular flow passage to a discharge hole for the working fluid, A fluid control valve in which an annular nozzle is configured to adjust the opening state between the annular flow path and the movable body by moving the movable body closer to or farther away from the outer valve seat tip and the inner valve seat tip.

2. A fluid control valve as claimed in claim 1, comprising a control chamber which is a space for accommodating the movable body, and a control port communicating with the control chamber, wherein the working fluid flows from the supply hole through the first nozzle portion which is a supply side nozzle into the control chamber, and flows out from the control chamber through the second nozzle portion which is a discharge side nozzle into the discharge hole, and the fluid pressure in the control chamber or the flow rate flowing in and out of the control port is controlled by the difference in flow rate passing through the first nozzle portion and the second nozzle portion.

3. A fluid control valve according to claim 1, wherein said driving means is configured to move said movable body surface, which is the opposing surface of said annular nozzle, approximately parallel to the axial direction of said annular nozzle.

4. The movable body is composed of an elastically deformable low-rigidity portion formed on the outer periphery and a high-rigidity portion formed in the center and less elastically deformable than the low-rigidity portion, and the high-rigidity portion is configured to move approximately parallel to the axial direction of the exhaust-side annular nozzle portion by the driving means, and the area occupied by the high-rigidity portion is approximately defined by a circle ΦD r The outer diameter of the outer valve seat tip of the exhaust side annular nozzle portion is approximated by Φd 0d As, ΦD r >Φd 0d 4. The fluid control valve according to claim 3, wherein the movable body and the annular nozzle portion are configured so that the following holds true:

5. A fluid control valve in which a closed loop magnetic circuit is formed by an electromagnet, the movable body, and a yoke material, and the driving means for the movable body is an attractive force due to Maxwell stress generated between the magnetic poles of the electromagnet and the movable body, wherein 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 part made of a magnetic material arranged to accommodate the outer periphery of the coil and the bottom side of the coil opposite to the movable body side, and a through flow passage is provided that passes through the support shaft, one side of the through flow passage is connected to an opening in the control valve body, and the annular nozzle is provided on the movable body side of the through flow passage.

6. A fluid control valve as described in claim 5, wherein the annular nozzle equipped on the movable body side of the through passage is composed of an annular nozzle section having the annular passage formed by the inner member and the outer member, and a passage conversion section connecting the through passage on the opening side, and the annular nozzle section and the passage conversion section are arranged in series within the through passage.

7. A fluid control valve as described in claim 4, wherein the magnetic pole is formed in the center of the movable body side end face of the support shaft, and an annular flow passage is formed on the outer periphery of the magnetic pole with the outer valve seat and the inner valve seat as two peripheral walls.

8. A fluid control valve as defined in claim 6, wherein said outer valve seat and said inner valve seat are constructed from a non-magnetic material.

9. A fluid control valve as claimed in claim 2, further comprising a bypass passage formed through said inner member, a first opening of this bypass passage being formed on said movable body side of a center of said inner valve seat provided at approximately the center in the radial direction of said inner valve seat, and a second opening being formed in said control chamber or in a passage connecting said control chambers.

10. A fluid control valve as described in claim 9, wherein the bypass passage is formed in either the first nozzle portion or the second nozzle portion, and a bypass through hole communicating with the bypass passage is formed in a movable body surface facing the center of an inner valve seat constituting the annular nozzle on the side where the bypass passage is not formed.

11. The minimum outer diameter of the inner nozzle in the annular nozzle portion where the bypass flow passage is not formed is d ir The maximum value of the gap between the inner valve seat tip and the movable body is X eff The approximate inner diameter of the bypass through hole is d B As, 11. The fluid control valve according to claim 10, wherein the bypass through hole is formed so that the above formula is satisfied.

12. A fluid control valve according to claim 10, wherein said drive means for said movable body is disposed on the annular nozzle side where no bypass flow passage is formed.

13. A fluid control valve as claimed in claim 9, wherein the bypass flow passage having a second opening directly connected to the control chamber and a through passage connecting the annular flow passage with the supply side or the discharge side are formed independently of each other inside the annular nozzle constituted by the inner member and the outer member.

14. A fluid control valve as described in claim 9, wherein the annular nozzle is composed of a through passage formed perpendicular or inclined to the flow direction of the annular passage, an opening of the through passage, and a flow passage conversion section connecting the annular passage and the through passage.

15. A fluid control valve comprising: a movable body; a first nozzle portion and a second nozzle portion; a movable body support portion for fixing the movable body; and a drive means for the movable body, the drive means being configured so that the movable body approaches one of the tips of the first nozzle portion and the second nozzle portion and moves away from the other tip portion, and the fluid control valve controls the flow rate difference passing through the first nozzle portion and the second nozzle portion, or the flow rate passing through the first nozzle portion and the flow rate passing through the second nozzle portion, individually, by the movement of the movable body, the first nozzle portion being composed of an inner member and an outer member that houses the inner member, and comprising: an inner valve seat formed on the movable body side of the inner member; an outer valve seat formed on the movable body side of the outer member; an annular flow passage formed by the peripheral walls of the outer valve seat and the inner valve seat; and a flow passage connecting the annular flow passage with a supply hole for a working fluid, or a flow passage connecting the annular flow passage with a discharge hole for a working fluid, a first nozzle portion configured to be connected to the first valve seat and a second nozzle portion configured to be connected to the second valve seat and a second ...

16. A fluid control valve as described in claim 15, comprising a control chamber which is a space for accommodating the movable body, and a control port communicating with the control chamber, wherein the working fluid flows from the supply hole through the first nozzle portion which is a supply side nozzle into the control chamber, and flows out from the control chamber through the second nozzle portion which is a discharge side nozzle into the discharge hole, and the fluid pressure in the control chamber or the flow rate flowing in and out of the control port is controlled by the difference in flow rate passing through the first nozzle portion and the second nozzle portion.

17. The opening area of ​​the single-hole nozzle is A d , the support stiffness of the movable body is K0, and the maximum displacement of the movable body is X max , the supply pressure is P s , the discharge pressure of the fluid control valve is P0, A d <K0X max / (P s -P0) within the range where the conditional expression is satisfied, d 17. The fluid control valve of claim 16, wherein:

18. A fluid control valve as described in claim 16, wherein a space having an inner valve seat of the first nozzle portion as its outer peripheral wall is defined as a supply side gap, and a perforated portion connecting the supply side gap and the control chamber is formed on the movable body surface at a position radially away from the opening of the single-hole nozzle.

19. A fluid control valve as claimed in claim 16, further comprising a bypass passage formed through said inner member, a first opening of this bypass passage being formed on the movable body side of a center of the inner valve seat provided at approximately the center in the radial direction of said inner valve seat, and a second opening being formed in said control chamber or in a passage connecting said control chambers.

20. A fluid control valve as described in claim 1, which is composed of an annular nozzle in which only a flow path is formed, a space whose outer wall is the inner valve seat of the annular nozzle on the supply side is a supply side gap, and a space whose outer wall is the inner valve seat of the annular nozzle on the discharge side is a discharge side gap, and a perforated portion is formed in the movable body surface to connect the supply side gap and the discharge side gap to the control chamber.

21. A fluid control valve as claimed in claim 1 or 15, wherein the annular nozzle is composed of an inner portion and an outer portion formed on the outer periphery of the inner portion, and the annular nozzle is composed of an inner valve seat formed on the movable body side of the inner portion, an outer valve seat formed on the movable body side of the outer member, an annular flow passage formed by the peripheral walls of the outer valve seat and the inner valve seat, and a plurality of connecting portions formed to prevent separation of the inner portion and the outer portion.

22. A fluid control valve comprising: a movable body; a nozzle section disposed opposite the movable body; an inlet of the nozzle section connected to a supply source of working fluid and a movable body support member for fixing the movable body; and a drive means for the movable body, the drive means changing a separation distance between the tip of the nozzle section and the movable body to control a flow rate of working fluid flowing out of the nozzle section, the nozzle section being composed of an inner member and an outer member for housing the inner member, and comprising: an inner valve seat formed on the movable body side of the inner member; an outer valve seat formed on the movable body side of the outer member; an annular flow passage formed by peripheral walls of the outer valve seat and the inner valve seat; and a flow passage connecting the annular flow passage with a supply hole for working fluid or a discharge hole for working fluid, the movable body and the tip of the outer valve seat and the tip of the inner valve seat approaching or separating from each other to form an annular nozzle for adjusting an opening state between the annular flow passage and the movable body, the movable body being configured as follows: The annular nozzle is configured with a valve body central portion, which is the opposing surface of the valve seat tip portion of the annular nozzle, and an elastic deformation portion having a through groove formed on the outer circumferential side of the valve body central portion. The axis of the movable body is defined as the Z axis, and coordinates of the X axis and the Y axis are defined with respect to the Z axis. The axial spring stiffness in the Z axis direction is defined as K Z , and X θ Direction, Y θ The torsional spring stiffness in the direction is K θX , K θY a valve seat tip portion of the annular nozzle by the driving means, the valve body central portion, which is an opposing surface of the valve seat tip portion, is capable of deforming three-dimensionally.

23. The approximate radius of the center of the valve body is r, and the K θX and the aforementioned K θY Compare the numbers and choose the larger one as K. θ Represented as 23. The fluid control valve according to claim 22, wherein the elastic deformation portion is configured to satisfy the above condition.

24. A fluid control valve comprising: a supply hole with a restriction; an intermediate control chamber into which the working fluid that has passed through the supply hole with a restriction flows; a control port communicating the intermediate control chamber with the outside; a nozzle portion into which the working fluid that has passed through the intermediate control chamber flows; a movable body; a discharge chamber which is a space into which the working fluid that has passed through the nozzle portion flows and which accommodates the movable body; a fluid discharge hole which communicates with the exhaust chamber; a movable body support portion which fixes the movable body; and a drive means for the movable body, wherein the drive means is configured to move the movable body between a position close to the tip of the nozzle portion and a position away from it, and wherein the drive means changes the distance between the tip of the nozzle portion and the movable body to control the fluid pressure in the intermediate control chamber, or the flow rate flowing into the intermediate control chamber from the supply hole or the flow rate flowing out from the intermediate control chamber to the discharge hole, The nozzle portion is composed of an inner member and an outer member that houses the inner member, and is provided with an inner valve seat formed on the movable body side of the inner member, an outer valve seat formed on the movable body side of the outer member, an annular flow passage formed by peripheral walls of the outer valve seat and the inner valve seat, and a flow passage connecting this annular flow passage to the supply hole or the annular flow passage to the discharge hole, and an annular nozzle is configured to adjust the opening state between the annular flow passage and the movable body by moving the movable body, and a tip end of the outer valve seat and a tip end of the inner valve seat closer to or farther away from each other.

25. The fluid control valve of claim 24, wherein said restrictor supply hole comprises a flow straightener for smoothing turbulent fluid flow.

26. A fluid control valve as claimed in claim 24, further comprising a bypass passage formed through said inner member, a first opening of this bypass passage being formed on the movable body side of the center of said inner valve seat, and a second opening being formed in said discharge chamber or in a passage connecting said discharge chambers.

27. When the input signal is applied to the driving means for the movable body, the gap between the movable body and the tip of the annular nozzle is h=0. When the gap h is increased, the flow area of ​​the centrifugal flow path between the tip of the annular nozzle and the movable body is A. R , the axial cross-sectional area of ​​the annular nozzle flow passage is A Z , A R = A Z The gap where the valve flow rate saturates at a constant value is defined as the flow direction switching gap h = h0. S Then, h0≦h max <h S Maximum stroke h within the range max 25. The fluid control valve according to claim 1, 15, or 24, wherein:

28. What is the envelope a of the flow characteristic curve in the range of 0≦h<h0, the envelope b of the region where the flow rate is constant, and the gap h=h at the intersection of the envelope a and the envelope b? C Then, h0≦h max <h C Maximum stroke h within the range max 28. The fluid control valve of claim 27, wherein:

29. A bypass passage is provided through the inner member, and the approximate inner diameter of the outer valve seat tip is the outer nozzle inner diameter d0, and the approximate outer diameter of the inner valve seat tip is the inner nozzle outer diameter d i Then, the flow path direction switching gap is h0 = (d0 - d i 25. The fluid control valve of claim 24, wherein the annular nozzle is configured such that:

30. A fluid control valve as claimed in claim 24, comprising as a first configuration the configuration of the fluid control valve of claim 22, and further comprising as a second configuration, similar to said fluid control valve, a second supply hole with a second restriction, a second control chamber, a second control port, a second nozzle portion, a second movable body, a second exhaust chamber, a second discharge hole, and a second movable body support portion, wherein, when actuated by a driving means of the first configuration, the nozzle portion and the movable body of one of the first and second configurations are brought close to each other and the nozzle portion and the movable body of the other configuration are separated, and a fluid control valve controls the fluid pressure in the control chamber, or the flow rate flowing into the control chamber from the supply hole with a restriction, or the flow rate flowing out from the control chamber to the discharge hole, by changing the distance between the tip of the nozzle portion and the movable body by the driving means.

31. A flow control valve as described in claim 1, 15, or 24, which is composed of a drive means for oscillating the movable body around a support part to which the movable body is fixed, wherein the movable body and the annular nozzle are configured so that the outer valve seat tip and the inner valve seat tip of the annular nozzle are in close contact with the movable body when the movable body is inclined at the maximum angle.

32. A fluid control valve as described in claim 31, wherein, when the movable body is at its maximum inclination, the axis of the annular nozzle is installed in a state inclined at an angle Φ so that the outer valve seat tip and the inner valve seat tip of the annular nozzle are in close contact with the movable body, with the inclination angle being Φ.

33. A fluid control valve as described in claim 31, wherein the annular nozzle side opposing surface of the movable body is formed in a state inclined at an angle Φ so that, when the movable body is at its maximum inclination, the inclination angle is Φ, and the outer valve seat tip and the inner valve seat tip of the annular nozzle are in close contact with the movable body.

34. A fluid control valve as described in claim 31, wherein the movable body-facing surface of the annular nozzle is formed in a state inclined at an angle Φ so that, when the movable body is at its maximum inclination, the outer valve seat tip and the inner valve seat tip of the annular nozzle are in close contact with the movable body, with the inclination angle being Φ.

35. A fluid control valve as claimed in claim 31, wherein a portion of the movable body is formed to be easily elastically deformed so that the outer valve seat tip and the inner valve seat tip of the annular nozzle are in close contact with the movable body when the movable body is at its maximum inclination.

36. The outer diameter of the outer valve seat tip of the annular nozzle is Φd 0d The width of the tip of the movable body, which is the opposing surface of the outer valve seat, is D r As, D r >d 0d 32. The fluid control valve of claim 31, wherein the movable body and the annular nozzle are configured so that:

37. A fluid control valve as set forth in claim 31, wherein said movable body is made of an alloy material having high vibration damping properties, such as a Co-Ni alloy, an Mg-Zr alloy, or an Mn-Cu alloy.

38. A fluid control valve as described in claim 1, 15, or 24, in which the driving means for the movable body is disposed at a position away from the axis of a supply side flow passage and a discharge side flow passage which house the forward direction nozzle section and the reverse direction nozzle section, the supply side flow passage and the discharge side flow passage being formed through a housing, and a flow passage connecting the first opening and the second opening of the bypass flow passage being formed in the housing.

39. A fluid control valve as claimed in claim 1, 15 or 24, comprising a coil bobbin fixed to the movable body, an electromagnetic coil wound around the coil bobbin, and a permanent magnet arranged around the outer periphery of the electromagnetic coil with a narrow gap therebetween, the coil bobbin, the electromagnetic coil and the permanent magnet constituting a linear actuator utilizing the Lorentz force, and the linear actuator being used as the driving means for the movable body.

40. The fluid control valve of claim 24, wherein said throttled supply hole comprises a flow straightener for smoothing turbulent flow of the working fluid.

41. A fluid control system comprising: a vibration isolation table; support means for the vibration isolation table relative to the base, which suppresses vibration transmission between the base and the vibration isolation table; vibration isolation means having a function of reducing micro-vibrations of the vibration isolation table; the vibration isolation table is installed in an environment having a vibration generating source that excites the vibration isolation table, and a vibration control actuator that applies a vibration control force to the vibration isolation table that reduces disturbances generated in the vibration isolation table based on information of the vibration excitation force applied to the vibration isolation table by the vibration generating source; the vibration control actuator being of a pneumatic type driven by a vibration control valve that draws in and exhausts gas therein, the vibration isolation means and the vibration control actuator being independently arranged so as to generate a force with a component in the same direction, and the vibration control valve being a fluid control valve citing claim 1, 15, or 24.

42. A fluid control system as described in claim 41, wherein the vibration damping valve and the vibration damping actuator are arranged so that the control port of the vibration damping valve and the intake hole of the vibration damping actuator are connected by a substantially straight flow path or a substantially L-shaped flow path.

43. A fluid control system comprising: a vibration isolation table; a support means for the vibration isolation table relative to the base, which suppresses vibration transmission between the base and the vibration isolation table; a vibration isolation actuator having a function of reducing micro-vibrations of the vibration isolation table, and a vibration isolation valve for driving the vibration isolation actuator; the vibration isolation table is installed in an environment having a vibration generating source that excites the vibration isolation table, and a vibration control actuator that applies a vibration control force to the vibration isolation table that attenuates disturbances generated in the vibration isolation table based on information of the vibration excitation force applied to the vibration isolation table by the vibration generating source, the vibration control actuator being of a pneumatic type driven by a vibration control valve that draws in and exhausts gas therein, the vibration isolation actuator and the vibration control actuator being independently arranged so as to generate forces with components in the same direction, and the vibration isolation valve being a fluid control valve recited in claim 24.

44. A fluid control valve and a fluid control device according to claim 1, 15, or 24, wherein the primary resonance frequency of the movable body is 200 Hz or higher.

45. A fluid control valve as claimed in claim 2 or claim 16, wherein a main control passage connected to a controlled object installed externally is formed penetrating the inner member, and an opening of the main control passage is formed on the movable body side of the center of the inner valve seat provided at approximately the radial center of the inner valve seat.

46. ​​A fluid control valve comprising: a supply-side nozzle section; a movable body; a movable body support section for fixing the movable body; and drive means for the movable body, wherein the drive means controls the flow rate passing through the supply-side nozzle section by varying the distance between the movable body and a tip of the supply-side nozzle section by movement of the movable body, wherein the supply-side nozzle section is composed of a supply hole, a throttling section for reducing the pressure of the working fluid that has passed through the supply hole, and a supply nozzle into which the working fluid that has passed through the throttling section flows, wherein the supply nozzle is composed of an inner member and an outer member for housing the inner member, and comprises: an inner valve seat formed on the movable body side of the inner member; an outer valve seat formed on the movable body side of the outer member; and an annular flow passage formed by the outer valve seat and peripheral walls of the inner valve seat, wherein the movable body and the tip of the outer valve seat and the tip of the inner valve seat are brought close to or apart from each other to constitute an annular nozzle that adjusts the opening state between the annular flow passage and the movable body.

47. A fluid control valve as described in claim 46, comprising: an exhaust side nozzle arranged opposite the supply side nozzle with the movable body in between; and the driving means configured so that the movable body approaches the tip of one of the supply side nozzle and the discharge side nozzle and moves away from the tip of the other, and the flow rate difference passing through the supply side nozzle and the discharge side nozzle is controlled by the movement of the movable body.

48. A fluid control valve as claimed in claim 46, wherein the space between the peripheral wall of said outer valve seat and the peripheral wall of said inner valve seat is constituted by an opening portion penetrating in the axial direction and a shielding portion having a sealing portion in the axial direction.

49. Distance X=X max Maximum flow rate Q max The angle ratio η of the opening is set to satisfy the following equation: The angle of the arc portion where the opening is formed is set to θ a The angle of the arc portion at which the shielding portion is formed is θ b Then, the hole angle ratio η = θ a / (θ a +θ b ) is defined as the maximum value of the distance X between the movable body and the tip of the supply side nozzle portion, and the distance X max So, X=X max Near the valve maximum flow rate Q max 49. The fluid control valve of claim 48, wherein the aperture angle ratio η is set so that:

50. A two-way, three-way valve with an arc-shaped annular nozzle on the intake side and a single-hole nozzle on the exhaust side. The annular nozzle is arranged on the supply side and the single-hole nozzle is arranged on the discharge side, with the movable body sandwiched between them, and a through hole is formed in the movable body between the outer periphery of the single-hole nozzle opening and the inner periphery of the inner valve seat of the annular nozzle.

51. A fluid control valve as claimed in claim 46 or 47, in which the intake side throttle portion is a single-hole nozzle, and the throttle portion in the supply side nozzle portion is constituted by a single-hole nozzle.

52. A fluid control valve as set forth in claim 48, wherein the throttle portion in said supply nozzle portion is constituted by a single-hole nozzle, and said supply nozzle is constituted by an annular nozzle constituted by an arc-shaped flow passage.

53. Let γ be the critical pressure ratio of the working fluid. c , the supply pressure P s The critical pressure for c =γ c P s , the downstream pressure P of the throttle portion a Flow rate Q characteristic and downstream pressure P a C against d =dQ / dP a In terms of characteristics, P a =P c Under the conditions of d 47. The fluid control valve according to claim 46, wherein the restricting portion is configured to satisfy .gtoreq.

0.

54. A fluid control valve as claimed in claim 46, wherein when a pressure difference is applied between the nozzle opening and outlet, the relationship between this pressure difference and flow rate is governed only by the viscosity of the working fluid, e.g., characteristic A of a nozzle whose flow path is constituted by a narrow gap, and characteristic B of a single-hole nozzle, e.g., characteristic B of a nozzle that is not affected by the viscosity of the working fluid, a nozzle having characteristics in the intermediate range between characteristic A and characteristic B is defined as a pseudo-viscosity nozzle, and the throttling portion in the first nozzle portion is constituted by the pseudo-viscosity nozzle.

55. The fluid control valve according to claim 46, wherein the restricting portion formed by the pseudo-viscosity nozzle is housed inside the housing of the servo valve.

56. The fluid control valve according to claim 46, wherein said annular nozzle and said restriction are coaxially formed.

57. A fluid control valve as set forth in claim 46, in which the throttle portion is disposed externally to a housing which houses major components such as the movable body, the supply nozzle, and the drive means for the movable body, and the opening portion of this throttle portion on the fluid supply source side serves as the supply hole.

58. A fluid control valve as set forth in claim 54, wherein said pseudo-viscosity nozzle is constructed by arranging a plurality of narrow gap flow paths in parallel.

59. A fluid control valve as described in claim 46, wherein the supply hole, the throttling portion, the annular nozzle, a flow passage connecting the throttling portion and the annular nozzle, a control port, etc. are contained in a supply side housing, the outer member constituting the annular nozzle is integrally formed with the supply side housing, and the inner member is contained and disposed in the center of the outer member.

60. A fluid control valve as described in claim 46, comprising a drive means for causing the movable body to oscillate around a support part to which the movable body is fixed, wherein the movable body and the annular nozzle are configured so that when the oscillation angle θ of the movable body is at 0, the outer valve seat tip and the inner valve seat tip of the annular nozzle are in close contact with the movable body, and the movable body performs a reciprocating motion.

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