WO1983001662A1

WO1983001662A1 - Digital drive control of compensated valves

Info

Publication number: WO1983001662A1
Application number: PCT/US1982/001415
Authority: WO
Inventors: Tadeusz Budzich
Original assignee: Tadeusz Budzich
Priority date: 1981-11-09
Filing date: 1982-10-20
Publication date: 1983-05-11
Also published as: JPS58501871A; EP0093150A1; CA1196253A; EP0093150A4

Abstract

A digital servo valve (10) with valve spool (19) driven in discrete steps by a digital spool actuator (78, 79), while the servo valve (10) is provided with positive and negative load compensation.

Description

Digital Drive Control of Compensated Valves

5 Background of the Invention

This invention generally relates to flow control valves provided with the feature of positive and negative load compensation.

In more particular aspects this invention

-^_Q relates to a digital valve spool drive, which moves the valve spool in discrete steps, in response to a digital signal, while the pressure differential acting across the spool is controlled by the positive and negative load compensators.

"L5 In still more particular aspects this invention relates to a digital servo valve, responsive to a digital control signal, with pressure differential across the control spool controlled by the positive and negative load compensator during control of positive or negative loads.

The great majority of electronic computing circuits, using micro-processors and computers, use a digital output signal. Digital valve spool drive, directly responding to such a signal, presents a difficult problem, since the valve spool must be driven in a series of discrete linear steps, for high response the number of steps per unit time being very high.

Summary of the Invention 0 It is therefore a principle object of this invention to provide a fast responding digital valve spool drive, which will move the valve spool in discrete linear steps in response to a low energy level digital input signal. 5 It is a further object of this invention to provide a digital servo valve, in which, due to positive and negative load compensation, the velocity of the load is always proportional to the input signal. It is a further object of this invention to provide a digital servo valve, the use of which provides a system which can be represented by a first order differential equation, making the system inherently stable. It is a further object of this invention to provide a servo valve, through which the position of the load can be controlled by a micro-processor without position feedback.

Briefly in the servo valve of this invention, directly responding to low energy level digital input signal, the velocity of the load is always proportional to the input signal, irrespective of the magnitude of the load or its polarity. The micro-processor, being a very powerful tool, driven by a crystal controlled clock and having the capability of performing complex mathematical and timing functions, by exerting control over the valve spool displacement, due to the proportional flow feature, will position very accurately the load without feedback. In many instances the feedback may not be necessary. When necessary, like for example in machine tools, the micro-processor will only correct for the final error.

Additional objects of this invention will become apparent when referring to the preferred embodiments of the invention as shown in the accompanying drawings and described in the following description. Description of the Drawings

Fig. 1 is a longitudinal sectional view of an embodiment of a digital servo valve including control spool, positive and negative load compensator and pilot valve stage with digital spool actuator, amplifying stage, lost motion stage, lines, system flow control, system pump, second digital servo valve and system reservoir shown diagrammatically;

Fig. 2 is a partial longitudinal sectional view of an embodiment of a spool drive provided with a hydraulic amplifying stage, the digital actuator being shown schematically;

Fig. 3 is a partial longitudinal sectional view of one embodiment of digital spool drive; Fig. 4 is a partial longitudinal sectional view of another embodiment of digital spool drive;

Fig. 5 is a partial longitudinal sectional view of spool drive of Fig. 2 including rotary to linear digital drive and lost motion mechanism with the digital motor, micro-processor, servo valve and other control system components shown diagrammatically; .

Fig. 6 is a diagrammatic representation of a digital servo valve in an analog control system.

Description of the Preferred Embodiments

Referring now to Fig. 1, an embodiment of a flow control valve composed of a control valve section, generally designated as 10, a compensator section, generally designated as 11 and a pilot valve section, generally designated as 12, is shown interposed between diagrammatically shown fluid motor 13 driving a load W and a pump 14 of a fixed or variable type driven by a prime mover not shown. If pump 14 is of a fixed displacement type, pump flow control 15 is a -differential pressure relief valve, which in a well known manner, by bypassing fluid from the pump 14 to a reservoir 16, maintains discharge pressure of pump 14 at a level, higher by a constant pressure differential, than load pressure developed in fluid motor 13. If pump 14 is of a variable displacement type pump flow control 15 is a differential pressure compensator, well known in the art, which by changing displacement of pump 14 maintains discharge pressure of pump 14 at a level, higher by a constant pressure differential, than load pressure developed in fluid motor 13.

The flow control section 10 is a four way type and has a housing 17 provided with a bore 18 guiding a valve spool 19. Valve spool 19 is equipped with lands 20, 21 and 22 which, in neutral position of valve spool 19, as shown in Fig. 1, isolate a fluid supply chamber 23, load chambers 24 and 25 and outlet chambers 26 and 27. Positive load sensing ports 28 and 29 communicate with bore 18 and are positioned between the supply chamber 23 and load chambers 24 and 25. Negative load sensing ports 30 and 31 communicate with bore 18 and are positioned between load chambers 24 and 25 and outlet chambers 26 and 27. The land 21 is provided with signal slots 32 and 33 in plane of positive load sensing ports 28 and 29 and circumferentially spaced metering slots 34 and 35. The land 20 is provided with a signal slot 36 in plane of negative load sensing port 30 and circumferentially spaced metering slot 37. The land 22 is provided with a signal slot 38 in plane of negative load sensing port 31 and circumferentially spaced metering slot 39. Load chambers 24 and 25 are connected for one way fluid flow by check valves 40 and 41 with the reservoir 16. The compensator section 11 has a housing 42 provided with a bore 43 slidably guiding a throttling spool 44. Bore 43 communicates with an inlet chamber

45, a supply chamber 46, a control chamber 47, an outlet chamber 48 and an exhaust chamber 49. The throttling spool 44 is provided with lands 50, 51 and 52 and biased, towards position as shown in Fig. 1, by a control spring 53. The land 50 of the throttling spool 44 defines space 54, which is connected to the 0 reservoir 16. The land 52 of the throttling spool 44 is provided with positive load throttling slots 55, terminating in throttling edges 56 and positioned between the inlet chamber 45 and the supply chamber

46. The land 51 of the throttling spool 44 is provided 5 with negative load throttling slots 57, terminating in throttling edges 58 and positioned between the outlet chamber 48 and the exhaust chamber 49.

The pilot valve section 12 comprises a housing 59 provided with a bore 60, slidably guiding a pilot o spool 61 and a.free floating piston 62, annular space 63 and control space 64. The pilot valve spool 61 has lands 65, 66 and 67, defining annular spaces 68 and 69. The land 65 projects into control space 64 and is biased by a pilot valve spring 70, through a spring 5 retainer 71. The land 67 is selectively engageable by the free floating piston 62, provided with a land 72, which defines spaces 73 and 74. Annular space 63 is connected with annular space 69 by a leakage orifice 75. Control space 64 is connected with annular space 69 and the reservoir 16 by a leakage flow section 76. Valve spool 19, of control valve section 10, is provided with manual input through manual input lever 77. The valve spool 19 is also provided with a digital control input from a stepper motor 78, through a hydraulic force amplifier 79 and a lost motion mechanism 80. The digital input signal is supplied to the stepper motor 78 from a solid state switch 81, while a transducer 82 supplies a feedback signal to a control circuit not shown. Positive load sensing ports 28 and 29, of the control valve section 10, are connected through lines 83, 84 and 85, check valve 86 and line 87 to control space 64. Positive load sensing ports 28 and 29 are also connected through line 88, check valve 89 and line 90 to the pump flow control 15, which also receives a control signal from control circuit 91 through a check valve 92. The output flow from the pump 14 is connected by discharge line 93 to the inlet chamber 45, while also being connected through check valve 94 and line 95 to annular space 68. Outlet chambers 26 and 27 are connected by lines 95, 96 and 97 with the outlet chambers 48, while also being connected through check valve 98 and lines 99 and 87 to control space 64. The outlet chamber 48 is connected by line 100 and check valve 101 to line 95. Negative load sensing ports 30 and 31 are connected through line 102 with space 74 in the pilot valve section 12. The supply chamber 46 is connected by line 103 with space 73. The control chamber 47 is connected by line 104 with annular space 63.

Referring now to Fig. 2, the stepper motor 78 of Fig. 1 is mounted on a cover 105 and engages, with its splined shaft 106, a coupling 107, which in turn engages the splined extension of rotary shaft 108, which is journalled by a bearing 109. The coupling 107 is provided with a gear section 110, radially spaced from a pulse pick-up 111, threaded in the cover 105 and retained by a lock nut 112. A threaded end 113 of the rotary shaft 108 engages the internal threads of an input sleeve 114, which is slidably guided in bore 115

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£*»^rΛ:T^"I provided in the cover 105 and suitably sealed by a seal 116. The input sleeve 114, with its slotted end 117 and pin 118, engages slot 119, provided in a servo link 120. The servo link 120, mounted by a pin 121 on a slotted end 122 of a pilot valve 123, engages with slot 124 a pin 125, located on extension 126 of cylindrical end 127 of an actuator 128. The actuator 128 is also provided with a piston 129, slidably engaging cylindrical surface 130 and defining spaces 131 and 132. The pilot valve 123, slidably mounted in bore

133, provided in a housing 134, has lands 135, 136, 137 and 138 defining annular spaces 139, 140 and 141. The lands 136 and 137 work in metering engagement with annular spaces 142 and 143. Annular space 142 is connected by passage 142a with space 131. Annular space 143 is connected by passage 143a with space 132. Annular space 140 is connected by passage 144, space 145, passage 146 and space 147 to schematically shown system reservoir. Space 147 is also connected by passage 148 to space 149 housing the servo link 120.

Annular spaces 139 and 141 are connected by passage 150 with the schematically shown system pump 14. Cylindrical ends of the actuator 128 are suitably sealed by seals 151 and 152. Referring now to Fig. 3, like components of

Figs. 1 and 2 are designated by the same numerals. The digital actuator 78 is composed of a stepper motor, generally designated as 153 and lead screw mechanism, generally designated 154. The stepper motor 153 is provided with a housing 155, locating a stator winding 156 and bearings 157 and 158. Bearings 157 and 158 journal the shaft 106 with a rotor 159. The shaft 106 engages, through its splined end, the coupling 107, which in turn engages threaded end of the rotary shaft 113. The rotary shaft 108, mounted in respect to the

_CVP y, \ . i__ housing 105 by the bearing 109, engages with its threaded end 113 the input sleeve 114. The shaft 106 is provided with an extension 160 protruding outside of the housing 155 of the stepper motor 153, to which a hand wheel 161 is suitably protected by a guard 163, engaging the housing 155 of the stepper motor 153. The energy to the digital actuator 78 and specifically to the stepper motor 153 is supplied through suitable wiring from a driver or solid state switch 81, which is 0 subjected to a pulse control input 165 and a direction of rotation control input 166.

Referring now to Fig. 4, which is very similar to Fig. 3, like components are denoted by the same numerals. An enlarged shaft 167 of the rotor 159 is 5 suitably mounted in the bearings not shown, of the housing 155 of the stepper motor 153 and protrudes on both sides of the stepper motor 153. One end of the enlarged shaft 167 carries the hand wheel 161, while the other end carries the gear section 110. The o enlarged shaft 167 is internally threaded and engages a threaded shaft 168, which in turn engages the internal threads of the input sleeve 114 and lock nut 169. The input sleeve 114, although prevented from rotation by the servo link 120 may be provided with slot 170, 5 engaging antirotational pin 171.

Referring now to Fig. 5, like components of Figs. 1 and 2 are designated by the same numerals. The digital actuator 78, composed of the stepper motor 153 and the lead screw mechanism 154, together with the 0 hydraulic force amplifier 79, are identical to those shown and described in detail when referring to Fig. 2. The actuator 128 is connected to the lost motion mechanism 80. The lost motion mechanism is shown in section. The end of the valve spool 19 is provided 5 with a bore 172, mounting threaded sleeve 173, provided with stop 174, internal cylindrical surface 175 and retaining ring 176. Internal cylindrical purface 175 guides reaction members 177 and 178, which are maintained by biasing force of a spring 179 against stop 174 and the retaining ring 176. A shaft 180 of the actuator 128 is located in position in respect to the sleeve 173 by retaining rings 181 and 182 engaging reaction members 177 and 178. Reaction members 177 and 178 are provided with cylindrical extensions 183 and 184 guided on the surface of the shaft 180. An electronic computing circuit which can be a micro-processor 185 transmits a digital control signal through line 186 to the solid state switch 81 connected to the digital actuator 78. The feedback signal from the transducer 82 is transmitted from the digital actuator 78 to the micro-processor 185 through line 187. The micro-processor 185 also transmits a control signal 188 to the solid state switch 81 operating the digital actuator 78 of a servo valve 189 composed of sections 10, 11 and 12 identical to those of Fig. 1.

The servo valve 189 controls fluid motor 13 operating a load W. Load position transducer 190 transmits load position signal to the micro-processor 185 through line 191. If the load position signal is analog, analog to digital converter 192 is provided. The micro-processor 185 receives a digital command signal from an input transducer 193.

Referring now to Fig. 6 a compensated servo valve 194, identical to that shown in Fig. 1, is provided with the analog to digital converter 192. Command signal 195 from an input transducer 196 is transmitted to a differential or differential amplifier 197. The differential amplifier 197 also receives a position feedback signal 198 from position transducer 199 of a fluid motor 200. From the differential or differential amplifier 197 an error signal 204 is transmitted to the analog to digital converter 192. Referring now to Fig. 1, the digital servo valve assembly is shown composed of three separate and distinct sections and that is the control valve section 10, the compensator section 11 and a pilot valve section 12. Although those sections, for better purposes of demonstration, are shown separated, actually they are combined into a single valve assembly. In general terms the control valve section 10 controls the direction of fluid flow to and from the fluid motor 13, selectively phasing its working chambers to the pump or to the system reservoir, which chamber is being pressurized depending on the polarity of the load W. The control valve section 10 provides variable area orifices leading to and from the fluid motor 13, the area of those orifices being controlled by the displacement of the valve spool 19 from its neutral position. The variable orifices leading to the fluid motor 13 and used in control of positive loads are created by displacement of metering slots 34 and 35. The variable orifices leading from the fluid motor and used in control of negative loads are created by displacement of metering slots 37 and 39. The valve spool 19 can be manually operated by the manual input lever 77, or its position can be controlled by the digital actuator 78, through the hydraulic force amplifier 79 and the lost motion mechanism 80. The electrical energy to the digital actuator 78 is supplied from a driver, or a logic chip, or a slid state switch 81. Digital control signal 165, see Fig. 3, by a number of pulses determines the number of linear steps of the digital actuator 78, while the other steady voltage control signal 166 determines by -li¬

the voltage level the direction of the linear steps. The digital transducer 82 senses the number of linear steps and transmits a digital feedback signal. While the position of the valve spool 19 is being controlled by the digital actuator 78, through the hydraulic force amplifier 79, the control spool 19 can be fully displaced in either direction by the manual input lever 77, overriding the automatic servo action, when the operator assumes the control. This feature is made possible by the lost motion mechanism 80, operation of which will be described later in the specification when referring to Fig. 5.

The pressure differential across the variable control orifices of the control valve section 10, interposed between the pump 14, the fluid motor 13 and the reservoir 16, during control of both positive and negative loads is controlled by throttling by the throttling spool 44 of the compensator section 11. While the positive load is being controlled the throttling edges 56, of positive load throttling slots 55, assume a position to sufficiently throttle the fluid flow from the system pump to maintain a constant pressure differential across metering slots 34 or 35. With constant pressure differential automatically maintained across the metering slots 34 or 35 the fluid flow into the fluid motor, during control of positive load, becomes proportional to the displacement of the valve spool 19 from its neutral position and independent of the magnitude of the positive load W. With negative load being controlled the throttling edges 58, of the negative load throttling slots 57 assume a position to sufficiently throttle the outlet fluid flow from the fluid motor 13, to maintain a constant pressure differential across metering slots 37 or 39. With constant pressure differential, automatically maintained across metering slots 37 or 39, the flow out of the fluid motor 13, during control of negative load, becomes proportional to the displacement of the valve spool 19 from its neutral position and independent of the magnitude of the negative load W. During control of positive load, positive load throttling slots 55 are always positioned upstream of the metering slots 34 and 35, while during control of negative load, negative load throttling slots 57 are always positioned downstream of metering slots 37 and 39. The position of the throttling spool 44 is determined by the control pressure in the control chamber 47, against the biasing force of the control spring 53. The pressure in the control chamber 47 of the throttling section 11 and therefore the amount of throttling of the pump pressure or the negative load pressure is controlled by the pilot valve assembly 12. During control of positive load the pilot spool 61 is subjected on one end to the positive load pressure in control space 64, transmitted from positive load sensing port 28 or 29 through line 83 or 84, line 85, check valve 86 and line 87, together with the biasing force of the pilot valve spring 70, while at the other end through line 103 it is subjected to pressure in the supply chamber 46, which is positioned downstream of positive load throttling slots 55. Subjected to those forces the pilot valve spool 61 assumes a modulating position, in which it controls the pressure in the control chamber 47, to sufficiently throttle the fluid flow from the inlet chamber 45, to maintain a constant pressure differential across metering slot 34 or 35. While controlling a positive load the free floating piston 62 is maintained by the pressure differential maintained across it all the way to the left, out of contact with the pilot valve spool 61. During control of negative load the pilot valve spool 61 is subjected on one end to the pressure in control space 64, which is connected by lines 87 and 99, check valve 98 and lines 95 and 96 to the outlet chambers 26 and 27, downstream of metering orifice 37 or 39, together with the biasing force of the pilot valve spring 70, while the other end of the pilot valve spool 61, through the free floating piston 62, is subjected to pressure in negative load sensing port 30 or 31 connected to space 74 by line 102. Subjected to those forces the pilot valve spool 61 assumes a modulating position, in which it controls the pressure in the control chamber 47, to sufficiently throttle fluid flow from the outlet chambers 26 and 27 to maintain a constant pressure differential across metering slots 37 and 39. While controlling a negative load the free floating piston 62 is maintained in contact with the pilot valve spool 61 by the pressure differential developed across it. The control space 64 is connected through the logic system of check valves 86 and 98 either with positive load sensing port 28 or 29, during control of positive load, or with outlet chamber 26 or 27 during control of negative load. This specific feature, together with the action of the free floating piston 62, permits the use of the same pilot valve section 12 in control of both positive and negative loads.

During control of positive loads the positive load pressure signals from the valve section 10 and the control circuit 91 are transmitted through the logic system of check valves 89 and 92 to the pump flow control 15.

The logic system of check valves 94 and 101, in a well known manner, transmits the fluid energy to the pilot valve section 12 either from the pump 14 or

OMPI from the negative load through the outlet chamber 48. This feature permits control of negative load W with the system pump 14 inactive.

During control of negative load, with the fluid being throttled by negative load throttling slots 57, throttling edges 56 cut off communication between the inlet chamber 45 and the supply chamber 46. Under those conditions the make-up fluid to the load chamber 24 or 25 is supplied through check valve 40 or 41 from the system reservoir 16 increasing the capacity of the pump 14 to perform useful work.

Referring now to Fig. 2, the digital actuator 78 may be in the form of a stepper motor 78, which will translate electrical pulses into discrete mechanical rotational movements of the shaft 106. With such a device, for each electrical impulse, the shaft 106 will rotate through a specific arc of rotation say, for example 15°. The direction of rotation of the stepper motor is determined by the signal supplied to the stepper motor driver, not shown. Each angular step of the shaft 106 will be transmitted through the coupling 107 to the rotary shaft 108, provided with threaded extension 113. Since the threaded extension 113 engages the internal thread of input sleeve 114, each angular step of the rotary shaft 108 will correspond to a certain specific linear displacement of the input sleeve 114, the magnitude of the linear step being established by the characteristics of the thread. Therefore the number of angular steps of the digital actuator 78 will be translated by the action of the rotary threaded shaft 108 into an equal number of linear steps, transmitted to input sleeve 114. The input sleeve 114 is part of the hydraulic force amplifier 79, which transmits those linear steps at higher force level to the valve spool 19, see Fig. 1. A very small stepper motor 78, in a manner as previously described, controls the position of the input sleeve 114, each angular step of the stepper motor 78 resulting in a proportional linear step of the input sleeve 114. The input sleeve 114 is provided with a slotted end 117, locating a pin 118, which engages through slot 119 the servo link 120. The servo link 120 is pivoted by slot 124 on pin 125, located on the extension 126 of the cylindrical end 127, which is part of the actuator 128. The servo link 120 is also pivoted for angular rotation by pin 121 secured to slotted end 122 of the pilot valve 123.

Assume that the input sleeve 114, with its pin 118, will be moved a number of linear steps from right to left. Since the pin 125 remains stationary, the servo link 120 will rotate in a counterclockwise direction moving through the pin 121, the pilot valve 123 from right to left. This motion, through displacement of land 136, will connect annular space 143 with annular space 141, thus automatically connecting, through passage 143, the oil under system pressure with space 132. At the same time through equal displacement of land 137, the annular space 142 will be connected to annular space 140, which is connected through passages 144 and 146 with the system reservoir, thus effectively connecting through passage 142a the space 131 with the system reservoir. The pressure differential, developed between spaces 132 and 131, will move the piston 129 and the actuator 128 from right to left, subjecting the servo link 120, through pin 125, to clockwise rotation and therefore moving the pilot valve 123 through pin 121 from left to right, to the position as shown, with the lands 136 and 137 effectively isolating spaces 131 and 132. Therefore each linear step of the input sleeve 114 from right to

OMPI left, through the above described action of the servo link 120 and pilot valve 123, will result in a proportional linear step of the actuator 128, the linear step of the actuator 128 being longer than the linear step of input sleeve 114 by the ratio of distances between pin 125 and pin 118 and pin 118 and pin 121. Therefore, small linear steps of the input sleeve 114 can be amplified into proportional larger linear steps of the actuator 128, as dictated by the geometry of the servo link 120.

Movement of the input sleeve 114 from left to right will rotate the servo link 120 around the pin 125 in a clockwise direction, moving the pilot valve 123 from left to right. The displacement of pilot valve 123 will connect space 131 with oil at system pressure and space 132 with system reservoir. The pressure differential between spaces 131 and 132 will move the piston 129 and the actuator 128 from left to right, rotating the servo link 120 around pin 118 in a counterlockwise direction and bringing the pilot valve 123 to the position as shown in Fig. 2. Therefore, each linear step of the input sleeve 114 from left to right will result in a proportional larger linear step from left to right of the actuator 128, due to the control action of the servo link 120 and pilot valve 123, the motion of the actuator 128 and pin 125 providing mechanical feedback.

Since a very small force is required to displace the pilot valve 123, a very small digital actuator 78, with a very high response, can be used. The rotary to linear motion converting mechanism of a screw is characterized by very high mechanical advantage and very large reduction in the length of the linear steps. Through the action of the servo link 120 of Fig. 2 those small digital linear input steps can be

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^Z°4θ amplified by the geometry of the servo link 120 of Fig. 2 into much larger digital steps of the actuator 128. Therefore the arrangement of Fig. 2 acts not only as a force amplifier, but also amplifies the digital linear input into a proportional larger digital output of the actuator 128. Therefore position of the actuator 128 can be effectively controlled in response to the digital input signal through the arrangement of Fig. 2. In a well known manner a pilot valve, similar to the pilot valve 123, can be located in the centrally located bore of the cylindrical end 127, providing a follow-up servo arrangement. With this type of servo the displacement of the input sleeve 114, directly connected to the pilot valve, will be exactly duplicated by the displacement of the actuator 128.

Since with this type of arrangement no amplification of the input signal takes place, the pilot valve must be displaced through the full control stroke of the actuator 128, thus resulting in a much slower acting mechanism with a much slower response.

The coupling 107 is provided with gear section 110, which preferably has the same number of teeth as the number of angular steps of the digital actuator 78, required for one complete revolution. The pulse pick-up 111, well known in the art, is positioned in respect to the periphery of the gear section 110, to obtain a proper working gap. The digital actuator 78, in the form of a stepper motor, is capable of high angular accelerations and decelerations, permitting a traverse of the individual teeth of the gear section 110 at comparatively high velocity past the pulse pick-up 111. This rapid traverse of each gear tooth, equivalent to each angular step of the stepper motor, will generate, in a well known manner, an electrical pulse in the pulse pick-up 111, which can be used to establish if any specific angular step of the digital actuator 78, in the form of a stepper motor, did take place.

Referring now to Fig. 3, the digital actuator 78, in the form of a stepper motor, is shown in greater detail. The stator 156 is usually composed of two coils. Two stator caps formed around each of those coils, with pole pairs mechanically displaced by half a pole pitch become alternately energized north and south magnetic poles. Between the two stator coil pairs the displacement is a quarter of a pole pitch. The permanent magnet rotor 159 is magnetized with the same number of pole pairs as contained by one stator coil section. Interaction between the rotor 159 and the stator 156 causes the rotor 159 to move one quarter of a pole pitch per winding polarity change. Depending on construction, a typical stepper motor will move either 48 steps per revolution or 7.5 per step, or will move 24 steps per revolution or 15 per step. The rotor 159 with its shaft 106 is journalled in the bearings 157 and 158. The electrical power to the stator 156 is supplied from the driver 81, which usually takes the form of a logic chip. The driver 81 receives a low power pulse signal 165, which determines the number of angular steps of the shaft 106 and also receives a steady voltage signal 166, the level of this voltage determining the direction of rotation of the shaft 106. The logic chip is essentially a solid state switching device, which responds to a low energy switching signal and connects, at an instant, comparatively high input current to the stepper motor 153. Therefore the logic chip acts as a form of amplifying device. The rotary motion, or rotary digital steps, of the shaft 106 are translated into linear steps by the rotary to linear motion translating

OMPI ™ mechanism 154, which was described in detail, when referring to Fig. 2. The linear digital steps of the drive are transmitted directly to the input sleeve 114 by the threaded end 113. One end 160 of the shaft 106 protrudes outside of the digital actuator 78 and is provided with the hand wheel 161, fastened to the shaft end 160 by the lock screw 162. With the stepper motor inactive, by manually turning the hand wheel 161, while utilizing the existing rotary to linear translating mechanism and the existing servo link 120, the position of the actuator 128 can be adjusted. This feature is very important in case of control failure, or when adjustment in the position of the load has to be made with the electrical system inactive. The end of the shaft 160 and the hand wheel 161 are protected by the removable guard 163, which can be either removed or installed on the stepper motor.

Referring now to Fig. 4, the digital actuator 78, in the form of a stepper motor, is provided with an enlarged shaft 167, secured to the rotor 159, the shaft and rotor being journalled in bearings, not shown. The enlarged shaft 167 is internally threaded to receive threaded shaft 168, which is threaded into input sleeve 114, of hydraulic force amplifier 79 and locked in position by the lock nut 169. The cylindrical end of the input sleeve 114 is provided with slot 170, which is engaged by the anti-rotation pin 171. Rotation of the rotor 159 and the enlarged shaft 167, in a well known manner, will transmit an axial movement to the threaded shaft 168. The arrangement of Fig. 4 performs in an identical way as the arrangement of Fig. 3, but it is simpler, since it requires one less bearing.

Referring now to Fig. 5, the digital control servo system using compensated servo valve of Fig. 1 is shown. The digital drive of Figs. 1, 2, 3 and 4 of the valve spool 19 of Fig. 1 is shown in detail together with the lost motion mechanism 80. The force and linear displacement of the actuator 128, of the hydraulic force amplifier 79, is transmitted to the valve spool 19 through the lost motion mechanism 80, which is provided to permit the manual displacement of the valve spool 19, using the manual input lever 77, see Fig. 1, through its entire control stroke, irrespective of the position of the actuator 128, position of which is controlled by the digital input drive 78. In this arrangement the automatic servo control function, say in position of a load, can be completely overridden at any instant by manual input from the operator, through manual input lever 77. The linear control input from the actuator 128 can be fully transmitted to valve spool 19 as long as the total effort to move the valve spool 19 does not exceed the preload in the spring 179. In the position as shown in Fig. 5, the spring 179 maintains the reaction member 177 against stop 174 and the reaction member 178 against the retaining ring 176, while also maintaining the reaction member 177 against the retaining ring 182 and reaction member 178 against the retainer ring 181. Therefore any force transmitted by the actuator 128, lower than the preload of spring 179, will be automatically transmitted from right to left through retainer ring 181, the reaction member 178, the spring 179, reaction member 177 to the stop 174 and therefore to the valve spool 19. Conversely any force transmitted to the actuator 128, lower than the preload of spring 179, will be automatically transmitted from left to right through the retaining ring 182, the reaction member 177, the spring 179, the reaction member 178 and the retaining ring 176 to the sleeve 173 and therefore to the valve spool 19. Therefore angular digital steps of the shaft of the stepper motor 153, in a clockwise or counterclockwise direction, will be transmitted as linear digital steps through the hydraulic force amplifier 79, moving the valve spool 19 from right to left or left to right, as long as the actuating force, transmitted through the lost motion mechanism 80, does not exceed the preload in the spring 179.

Assume that with the digital actuator 78 inactive the valve spool 19 must be moved manually to perform a function. Since as is well known in the art, the conventional thread of threaded extension 113, engaging the input sleeve 114, is mechanically irreversible, the position of the input sleeve 114 will remain unchanged. Movement of the valve spool 19 from left to right will then, through the reaction member 177, compress the spring 179, with the retaining ring 182 leaving the reaction member 177, while the reaction member 178 is maintained stationary by the retainer ring 181, the reaction force of the compressed spring 179 being transmitted to the hydraulic force amplifier 79 or to the input sleeve 114. Distance between the reaction members 177 and 178 is so selected, that it is greater than the maximum stroke of the valve spool 19. In this way, irrespective of the position of the actuator 128, the valve spool 19 can be manually displaced from left to right through its entire control stroke.

With the digital actuator 78 inactive and the valve spool manually displaced from right to left, the manual actuating force is transmitted through the sleeve 173, retaining ring 176 and reaction member 178, compressing the spring 179, while the reaction member 177 is maintained stationary by the retaining ring 182 of the actuator 128, the reaction force of the spring

OMPI compression being transmitted to the hydraulic force amplifier 79 or to the input sleeve 114. Since as previously described the distance between the reaction members 177 and 178 is greater than the maximum control stroke of the valve spool 19, the valve spool 19 can be actuated from right to left through its entire control stroke, irrespective of the position of the actuator 128.. Therefore with the digital actuator 78 inactive, the valve spool 19 can be manually displaced through its entire control stroke in either direction through the lost motion mechanism 80, permitting manual control of the flow control valve of Fig. 1 or flow control valve 189 of Fig. 5, irrespective of the position of the input sleeve 114 and therefore irrespective of the actuating position of the digital actuator 78 and the hydraulic force amplifier 79.

The fluid under pressure is supplied from schematically shown pump 14 to the servo valve assembly of Fig. 1 including control valve section 10, compensator section 11 and pilot valve section 12 by line 201 and to servo valve 189 by line 202 or 203. The micro-processor or micro-computer 185 may be of conventional electronic type, well known in the art. Such a micro-computer, being a very powerful tool, driven by a crystal controlled clock and having the capability to be programmed to perform complex mathematical, logic and timing functions, can provide both valve displacement control and output response modeling. The positive and negative load, compensation of the servo valves of this invention make real time modeling without pressure or position feedback possible. The positive and negative load compensation automatically provides a flow rate to and from the servo valve, proportional to the displacement of its spool, independently of a very wide range of load reactions. Therefore, the micro-computer can very accurately compute servo valve flow rate during a time interval and perform the time integration over the interval in accordance with a fluid motor-load dynamics algorithm, to compute the output position at discrete instants of time. The computed position can be utilized in a control algorithm to provide a position control system through stepper motor displacement of the servo valve, without the use of a physical position or pressure feedback signals. This positioning control can be so accurate, that in many instances, especially when dealing with a very stiff system, with low compressibility of oil, or oil with high bulk modulus, the position feedback may not be necessary. Even when dealing with a comparatively soft system, with large volume of oil under pressure and low bulk modulus of the oil subjected to random loads, the compressibility effect can be accurately computed by the micro-processor and the position of the load corrected, once the micro-processor is provided with the input signal related to load pressure. When the accuracy of positioning of the load or tool is extremely important like in the case for example in machine tools, the micro-processor will not only control the velocity of the approach of the tool and therefore its rate of feed, but it will arrive at the required position in minimum time the position feedback being only used to correct for the final small error. The load response to the incremental displacement of the servo valve under stepper motor actuation is significantly effected by the positive and negative load compensation of this invention. The constant flow rate characteristics imposed at each interval between steps attenuate the mass-compressibility induced load velocity variations and thereby increases system stability. The increased stability permits increased loop gain and thereby increased speed of response.

The micro-processor 185 receives a digital input signal either from an input transducer or from another computing device 193. In response to this signal or to its program the micro-processor 185 sends a command signal to the driver, or solid state switch 81, connected to the digital actuator 78 through line 186 and may, in a manner as described when referring to Fig. 3, be provided with pulse signal feedback, defining the actual rotation, or the number of angular steps, actually performed by stepper motor 153. Such a feedback may be of importance when the number of steps °^r pulses, transmitted from the micro-processor per unit time, exceed the capability of the stepper motor 153. This might be the case when the micro-processor 185 would be substituted by a simple computing circuit, incapable of control of the maximum rate of the digital pulses. In such a case a signal storage device, well known in the art, can be incorporated in line 186, to store the number of pulses transmitted from the computing circuit and only release those pulses at a certain maximum rate, without the range of the response capability of the stepper motor 153. Then with the sophisticated micro-processor, capable of tuning its output signal to the capability of the stepper motor, the pulse feedback may be very useful, since it may detect the malfunction of the stepper motor or an increased load due to stiction, high oil viscosity during cold starts etc.

The micro-processor 185 also controls through the digital control signal 188 the servo valve 189, provided with the lost motion mechanism 80, the hydraulic force amplifier 79, the digital actuator 78 and a position switch 82a. The position switch 82a is used upon starting of the electronic system, but before starting the system pump 14, to establish exactly the position of the control spool 19 and to bring it into its neutral position, from which the micro-processor can establish the control base line. The fluid motor 13, controlling load W, is provided with a position transducer 190, which through line 191 provides the micro-processor 185 with the position feedback of the load. If the position transducer 190 is of an analog type, most common and well known in the art, the analog to digital convertor 198 is positioned in line 191, between the position transducer 190 and the micro-processor 185. If the position transducer 190 is of a digital type, analog to digital convertor 198 is not necessary. Instead of positioning switch 82a, a spool position transducer, say well known LVDT, with an analog to digital converter, can be used. The manual input lever 77 is provided so that through the lost motion mechanism 80, in a manner as previously described, the control cycle of the micro-processor 185 can be interrupted and the manual control assumed by the operator.

Even using the servo valve to control the position of the load W in a conventional way, with the feedback from the position transducer, the servo system of Fig. 5 carries an enormous advantage over conventional analog and digital servo systems. The micro-processor can compute the position of the load, compare this position with the position as indicated by the position feedback and sound the warning, stop the system or automatically override the feedback signal during malfunction, or shift in transducer calibration. Referring now to Fig. 6 the servo valve system of this Invention is shown adapted to replace an analog servo valve of the torque motor-flapper type. The system of Fig. 6 accepts an analog input and utilizes an analog feedback transducer. Thus the analog input signal 195 from input generator 196 and the analog position feedback signal 198a from position transducer 199 are fed into differential 197. The output 204 of the differential 197 is fed to analog to pulse converter device 192. The analog to pulse converter

192 may be provided with a micro-computer capability or might be just a simple device converting an analog error signal 204 into a specific number of pulses proportional to the magnitude of the analog error signal 204. Those pulses are transmitted to the driver or a solid state switch 81 which, in a manner as previously described, will trigger per each pulse a step of the stepper motor and lead screw combination 78. Each linear step, delivered to the fluid power amplifier 79, will be transmitted at higher force level through the lost motion mechanism 80 to the valve spool of the digital servo valve 194. This embodiment permits the digital servo-motor system of this invention to be a direct substitution in a system designed for an analog servo motor.

The flow control valve of Fig. 1 is shown using a variable or fixed displacement type pump, flow of which is varied in response to the load pressure signal 90. The flow control valve of Fig. 1, due to its feature of positive and negative load compensation can be supplied from a variable displacement type pump, provided with a constant pressure control, well known in the art, or from a fixed displacement pump, provided with a conventional maximum system pressure relief valve. The fully compensated flow control valve of Fig. 1 will provide proportional flow during control of both positive and negative load, while supplied from such constant pressure systems. When used in such systems this flow control valve is not load responsive. Although the preferred embodiments of this invention have been shown and described in detail it is recognized that the invention is not limited to the precise form and structure shown and various modifications and rearrangements as will occur to those skilled in the art upon full comprehension of this invention may be resorted to without departing from the scope of the invention as defined in the claims.

Claims

1. A valve assembly having valve means (19) operable to control by throttling fluid flow to and from a fluid motor (13) subjected to an opposing or aiding load, first fluid throttling means (55,56) positioned upstream of said valve means (19) responsive to pilot valve means (61) and operable to control pressure differential across said valve means (19) during control of an opposing load, second fluid throttling means (57,58) positioned downstream of said valve means (19) responsive to pilot valve means (61) and operable to control pressure differential across said valve means (19) during control of an aiding load, actuating means (79,78) operable to actuate said valve means having control means (81) responsive to an electrical control signal (186) , and electrical signal generating means (185,196) operably connected to said control means (81) whereby fluid flow to and from said fluid motor (13) can be proportionally controlled in respect to the magnitude of said electrical control signal (186) irrespective of the change in magnitude of said opposing or said aiding load.

2. A valve assembly as set forth in claim 1 wherein said actuating means (79,78) includes fluid power force amplifying means (79,123).

3. A valve assembly as set forth in claim 1 wherein limiting force transmitting lost motion means

(80) is interposed between said actuating means (79,78) and said valve means (19) .

4. A valve assembly as set forth in claim 1 wherein manual control means (77) is interconnected to said valve means (19) .

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5. A valve assembly as set forth in claim 1 wherein said control means (81) has means responsive to intermittant pulse type signal (186) .

6. A valve assembly as set forth in claim 1 wherein said electrical signal generating means (185) includes intermittant pulse type signal generating means.

7. A valve assembly as set forth in claim 1 wherein said actuating means (79,78) includes linear step output means (128) operable to actuate said valve means (19) in discrete steps.

8. A valve assembly as set forth in claim 1 wherein said actuating means (79) includes stepper motor means (153) .

9. A valve assembly as set forth in claim 1 wherein said electrical signal generating means

(185,196) includes computing means (185) operable to control and compute position of said opposing or said aiding load by computing fluid flow through said valve means (19) while controlling linear displacement of said valve means (19) .

10. A valve assembly as set forth in claim 9 wherein said computing means (185) includes micro-processor means.

11. A valve assembly as set forth in claim 9 wherein command signal means (193) is interconnected to said computing means (185) .

12. A valve assembly as set forth in claim 9 wherein position transducer feedback means (198,190) providing a proportional signal to position of said opposing or aiding load is interconnected to said computing means (185) .

13. A valve assembly as set forth in claim 9 wherein pulse set feedback means (82,111,110) provides a signal from step output means (110) to said computing means (185) .

14. A valve assembly as set forth in claim 1 wherein said electrical signal generating means

(185,196) includes analog signal generating means (196) and analog to digital signal converting means (192) .

15. A valve assembly as set forth in claim 14 wherein said analog signal generating means (196) includes differential means (197) and analog feedback means (198,199) operable to generate an analog signal proportional to position of said load W and supply it to said differential means (197) .