CA1196253A

CA1196253A - Digital drive control of compensated valves

Info

Publication number: CA1196253A
Application number: CA000414250A
Authority: CA
Inventors: Tadeusz Budzich
Original assignee: Individual
Current assignee: Individual
Priority date: 1981-11-09
Filing date: 1982-10-27
Publication date: 1985-11-05
Also published as: EP0093150A4; WO1983001662A1; EP0093150A1; JPS58501871A

Abstract

Digital Drive Control of Compensated Valves Abstract A digital servo valve with valve spool driven in discrete steps by a digital spool actuator, while the servo valve is provided with positive and negative load compensation.

Description

Description Digital Drive Control of Compensated Valves Background of the Invention ~ his invention generally relates to flow control valves provided with the feature of positive and negative load compensation.
In more particular aspects this invention 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.
In still more particular aspects this invention relates to a digital servo valve, responsive to a digital control signal, with pressure di~ferential across the control spool controlled by the positive and negative load compensator during control o~ 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 driving in a series o~ discrete linear steps, for high response the number of steps per unit time being very high.

_mmary of the Invention According to one aspect of the present invention, a valve assembly is provided having valve means operable to control by throttling fluid flow to ~362~3 and from a fluid motor subjected to an opposing or aiding load, first fluid throttling means operable to control pressure differential across said valve means during control of an aiding load, second fluid throttling means operable to con-trol pressure differential across said valve means during control of an aiding load, pressure fluid operated force amplifying means operably connec-ted t~) said valve means1 linear step output means including a s-tepper rnotor means interconnected -to a rotary to linear motion -translating means operable -through said force amplifying means to actuate said valve means in discrete linear steps in response to an intermittent pulse type control signal, and adjus-ting means having means operable through said linear step output means to selectively adjust the position of said valve means -to any inter-mediate position in respect to said stepper motor means in the absence of said intermittent pulse type control signal.
Briefly, in the servo valve of this invention, directly responding to a low energy level digital input signal, the velocity of the load is always proportional t~, 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 mathema-tical and timing functions, by exerting control over the valve spool displacement, due to the proportional flow feature, will posi-tion 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 Einal error.
Additional objects of this invention will become apparent .~ ~ - 2 -6~3 when referring to the preferred embodiments of the invention as shown in the accompanying drawings and described in the following descrip-tion.

IS!~

i3 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;
~ig. 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 sectionr generally designated as 10, a compensator section, generally designated as 11 and a pilot valve sectiont generally designated as 12, is shown interposed between diagrammatically shown fluid motor 13 driving a load W
and a pump 1~ 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 relieE
valve, which in a well known manner, by bypassing fluid from the pump 14 to a reservoir 16, maintalns discharge pressure of pump 14 at a level, higher by a constant pressure differential, than load pressure developed in ~luid 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 differentiall than load pressure developed in fluid motor 13.
The flow control section 10 is a :Eour 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 2~ 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 circumferentiall~
spaced metering slot 39. Load chambers 24 and 25 are connected for one way fluid flow by check valves ~0 and 41 with the reservoir l~o 5~

The compensator section 11 has a housing 42 provided with a bore ~3 slidably guiding a throttling spool 4~. Bore 43 communicates with an inlet chamber 45, a supply chamber 46, a control chamber 47, an outlet chamber 48 and an e~haust chamber 49. The throttling spool 44 is provided with lands 50, 51 and 52 and biased, towards position as shown in FigO 1, by a control spring 53~ The land 50 of the throttling spool 44 defines space 54, which is connected to the 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 ~5 and the supply chamber 46. The land 51 of ~he throttling spool 44 is provided 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 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 retainer 71. The land 67 is selectively engageable by the free floating piston 62, provided with a land 7~, which defines spaces 73 and 74. Annular space 63 is connected with annular space 69 by a leakage orifice 75. Control space 6~ 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 p.rovided with manual input through manual input lever 77. The valve spool 19 is also provided with a digital control input ~rom 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 ~4. 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. Ou~let chambers 26 and 27 are connected by lines 95 t 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 95O Negative load sensing ports 30 and 31 are connected through line 102 with space 74 in the pilo~ valve section 12. Th 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 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 5 slotted end 1~2 oE 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 10 132. 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 15 connected by passage 142a with space 131. Annular space 143 is connected by passage 1D~3a with space 132.
Annular space 140 is connected by passage 144, space 145, passage 14Ç and space 147 to schematically shown ~ystem reservoir. Space 147 is also connected by 20 passage 148 to space 149 housing the servo link 1200 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 ~ 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 30 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 oE the rotary shaft 35 113. The rotary shaft 108, mounted in respect to the i;25i3 housing 105 by the bearing 109, engages with its threaded end 113 the input sleeve 114. The shaft 10~
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 ~he stepper motor 153 is supplied through suitable wiring from a driver or solid state switch 81, which is subjected to a pulse control input 165 and a direction of rotation control input 16~.
Referring now to Fig. 4, which is very similar to Fig. 3, like components are denoted by the same numerals. An enlarged shaft 167 Gf the rotor 159 is suitably mounted in the bearings not shownl 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 enlarged shaft 167 is internally threaded and en~ages a threaded shaft 168, which in turn engages the internal threads of the input sleeve 114 and lock nu~ 169. The input sleeve 114, although prevented from rotation by the servo link 120 may be provided with slot 170, 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 hydraulic force amplifier 79, are identical to those shown and described in detail when Leferring 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 with a bore 172, mounting threaded sleeve 173, provided 2~;;3 with stop 17~, internal cylindrical surface 175 and retaining ring 176. Internal cylindrical surface 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 en~aging reaction members 177 and 1780 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 i5 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 7~ of a servo valve 189 composed of sections 10, 11 and 12 identical to those of FigO 1~
The servo valve 18g 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 conver~er 192.
Command signal 195 from an input transducer 196 is transmi~ted to a differential or differential amplifier 197. The differential amplifier 197 also receives a position feedback signal lg8 from position transducer 2~3 lg9 of a fluid motor 200. From the differential or differential arnplifier 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 lS chamber is being pressurized depending on the polarity of the lsad 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 3~ 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 mot1on mechanism 80. The electrical energy to the digital actuator 78 is supplied Erom a driver, or a logic chip, or a slid state switch 81. Digital control signal 165, see Fig.

3 r 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 ~9E;~

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 actionr 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. 5v 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 3~ 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 ~otor 13, to maintain a constant pressure difEerential across metering slots 37 or 39. With constant pressure differentialJ

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 neutr~l 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 5ubjected on one end to the positive load pressure in control space 64, transmitted from positive load sensing port 28 or 2~ 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 floatlng 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 Eorce of the pilot valve spring 70, while the other end of the pilot valve spool 61~ through the free ~loating 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 sufficientl~ throttle 1uid 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 ito 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 c~ntrol 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 ~9 and 92 to the pump flow control 15.
The logic system of check valves 94 and 101, in a well kno~n manner, transmits the fluid energy to the pilot valve section 12 either from the pump 14 or 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 negati~e 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 ~apacity of the pump 1~ 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 shaEt 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 wiil 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 11~ 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.

.

2~

A very small stepper motor 78, in a manner as previously described, controls the position of ~he input sleeve 114, each angular step of the stepper motor 78 resulting in a proportional lin~ar 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 12~ on pin 125, located on the extension 126 oE 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 leftO Since the pin 125 remains stationary, the servo link 120 will rotate in a counterclockwise direction moving through the pin 121, the pilot ~alve 123 from right to left. This motion, through displacement of land 136, will connect annular space 143 with annular space 141, thus automatically connecting, ~hrough passage 143, the oil under system pressure with space 132. At the same time through equal displacement of land 137, the annular space 14 will be connected to annular space 1~0, 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 131f 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 13~ and 137 effectively isolating spaces 131 and 132. Therefore each linear step of the input sleeve 114 from right to , 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 lnput sleeve 114 can be amplified into proportional larger linear steps of the actuator 128, as dictated by the geometry of the servo link 120.
~ ovement of the input sleeve 114 from lef-t 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 12Q
of ~ig. 2 those small digital linear input steps can be ~ 5~

amplified by the geometry of the servo link 120 of Fig.
2 into much larger digital steps of the actuator 1280 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 1~3, 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 ~alve~ 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 i3 establish if any specific angular step of the di~ital 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 ampli~ying device. The rotary motion, or rotary digital steps, of the shaft 106 are translated into linear steps by the rotary to linear motion translating 5i3 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 ~he 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 %0. The force and linear displacement of the actuator 128, of the hydraulic force amp:Lifier 79, is transmitted to the 5 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, 10 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 inpu-t from the operator, through manual input lever 77. The 15 linear control input from ~he actuator 128 can be ful~y 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 20 177 against stop 174 and the reaction member 178 against the retaining ring 17~, 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, 25 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 f orce 30 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 35 and therefore to the valve spool 19. Therefore angular Ç;2~

digita]. steps of the shaft o:E 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 5 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 ~ith the digital actuator 78 10 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 15 remain unchangedO 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 20 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.
25 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 30 valve spool manually displac~d 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 35 of the actuator 128, the reaction force of the spring ;3 compression being transmitted to the hydraulic force amplifier 79 or to the input sleeve 11~. Since as previously described the ~istance between the reaction members 177 and 178 is greater than the maximum control stroke o~ the valve spool 19, the valve spool 19 can be actuated fro~ right to left through its entire control stroke, irrespective of the position oE the actuator 128. Therefore with the digital actuator 78 inactiv~, the valve spool 19 can be manually displaced through its entire control s~roke in either direction through the lost motion mechanism 80, permitting manual control of the flow control valve of Fig~ 1 or flow control valve 1~9 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 ~01 and to servo valve 189 by line 202 or 2~3.
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 o~ load i3 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 signalsO 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 ~eedback 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 pressureO 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 ~low rate characteristics imposed at each interval between steps attenuate the mass-compressibility induced load velocity variatio~s -2~-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 inp~t 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 si~nal to the driver, or solid state switch 81, connected to the digital actuator 78 through line 186 and may, in a manner as described ~hen referring to Fig. 3, be provided with pulse signal feedback, deining 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 or 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 135 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 strating 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 converter 198 is positioned inline 191, between the position transducer 190 and the micro-processor 185. If the position transducer 190 is of a digital type, analog to digital converter 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. ~ 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 20~ of the diEferential 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 transmit~ed 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 valveO The fully compensated 10w control valve of Fig. 1 will provide proportional flow during control oE
both positive and negative load, while supplied from such constant pressure systems. When used in such systems this flow control valve is rot 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 Erom the scope of the invention as defined in the claims.

Claims

THE EMBODIMENTS OF THE INVENTION IN WHICH AN EXCLUSIVE
PROPERTY OR PRIVILEGE IS CLAIMED ARE DEFINED AS FOLLOWS:

1. A valve assembly having valve means operable to control by throttling fluid flow to and from a fluid motor subjected to an opposing or aiding load, first fluid throttling means operable to control pressure differential across said valve means during control of an opposing load, second fluid throttling means operable to control pressure differential across said valve means during control of an aiding load, pressure fluid operated force amplifying means operably connected to said valve means, linear step output means including a stepper motor means interconnected to a rotary to linear motion translating means operable through said force amplifying means to actuate said valve means in discrete linear steps in response to an intermittent pulse type control signal, and adjusting means having means operable through said linear step output means to selectively adjust the position of said valve means to any intermediate position in respect to said stepper motor means in the absence of said intermittent pulse type control signal.

2. A valve assembly as set forth in claim 1 wherein limiting force translating lost motion means is interposed between said force amplifying means and said valve means.

3. A valve assembly as set forth in claim 1 wherein manual control means is interconnected to said valve means.

4. A valve assembly as set forth in claim 1 wherein said intermittent pulse type signal is generated by signal generating means including digital computing circuit means.

5. A valve assembly as set forth in claim 1 wherein said adjusting means includes manual input means and connecting means operable to transmit a rotary motion between said manual input means and said stepper motor means.

6. A valve assembly as set forth in claim 1 wherein said adjusting means includes manual input means and connecting means operable to transmit a rotary motion between said manual input means and said rotary to linear motion translating means.

7. A valve assembly as set forth in claim 6 wherein a selectively disengageable closure means is operable to isolate said manual input means.

8. A valve assembly as set forth in claim 1 wherein inter-mittent pulse type signal generating means is operable to transmit an intermittent pulse type control signal to said linear step output means.

9. A valve assembly as set forth in claim 8 wherein a switching logic means is interposed between said intermittent pulse type signal generating means and said linear step output means.

10. The valve assembly as set forth in claim 1 wherein said adjusting means is manually actuated.

11. The valve assembly as set forth in claim 1 wherein said adjusting means has means operable to actuate a fluid power ampli-fying means in said linear output means.