EP0085674A4

EP0085674A4 - Load responsive fluid control valve.

Info

Publication number: EP0085674A4
Application number: EP19810902365
Authority: EP
Inventors: Tadeusz Budzich
Original assignee: Individual
Current assignee: Caterpillar Inc
Priority date: 1981-08-20
Filing date: 1981-08-20
Publication date: 1986-02-13
Also published as: DE3176924D1; EP0085674B1; WO1983000728A1; JPS58501283A; JPH0232483B2; EP0085674A1

Abstract

A direction flow control valve (95) for control of positive loads equipped with a pilot operated load responsive positive load control (13b, 63b, 82, 13), which automatically regulates valve inlet pressure to maintain a relatively constant pressure differential between inlet pressure and load pressure and which permits variation in the level of pressure differential in response to an external control signal (59) while this pressure differential is maintained constant at each controlled level.

Description

Load Responsive Fluid Control Valve

This is a continuation in part of application

Serial No. 949,250, filed October 6, 1978, for "Load Responsive Fluid Control Valve".

Background of the Invention This invention relates generally to load responsive fluid control valves and to fluid power systems incorporating such valves, which systems are supplied by a single fixed or variable displacement pump. Such control valves are equipped with an automatic load responsive control and can be used in a multiple load system in which a plurality of loads is individually controlled under positive load conditions by separate control valves.

In more particular aspects this invention relates to direction and flow control valves capable, in a proportional way, of controlling simultaneously a number of loads under positive load conditions.

In still more particular aspects this invention relates to pilot operated load responsive controls of direction control valves, which permit variation in the level of control differential between pump discharge pressure and the load pressure' signal, while this control differential is automatically maintained constant at each controlled level. In still more particular aspects this invention relates to pilot operated load responsive controls of direction control valves, which permit variation in the controlled pressure differential between valve ir.let pressure and the load pressure, in response to an external control signal.

O PI Closed center load responsive fluid control valves are very desirable for a number of reasons. They permit load control with reduced power loss and therefore, increase system efficiency and when controlling one load at a time provide the proportional feature of flow control, irrespective of the variation in the magnitude of the load. Normally such valves include a load responsive control, which automatically maintains pump discharge pressure at a level higher, by a constant pressure differential, than the pressure required to sustain the load. A variable orifice, introduced between pump and load, varies the flow supplied to the load, each orifice area corresponding to a different flow level, which is maintained constant, irrespective of the variation in magnitude of the load. The application of such a system is however, limited by one basic system disadvantage.

Normally in such a system the load responsive valve control can maintain a constant pressure differential and therefore constant flow characteristics when operating only one load at a time. With two or more loads, simultaneously controlled, only the highest of the loads will retain the flow control characteristics, the speed of actuation of lower loads varying with the change in magnitude of the highest load. Fluid control valve for such a system is shown in U.S. Patent No. 3,488,953 issued to Haussler.

This drawback can be overcome in part by the provision of a proportional valve, as disclosed in my U.S. Patent No. 3,470,694, dated October 7, 1969 and also in U.S. Patent No. 3,455,210 issued to Allen on July 15, 1969. However, while those valves are effective in proportionally controlling multiple positive loads at a time, they provide a constant

O PI pressure differential and therefore a constant throttling loss across each valve, thus reducing system efficiency. Also those valves use an unamplified load pressure signal in operation of their controllers requiring a control signal at a comparatively large energy level.

Summary of the Invention

It is therefore a principal object of this invention to provide an improved pilot operated load responsive direction control valve, which permits variation in the level of control differential between valve inlet pressure and load pressure, while this control differential is automatically maintained constant at each controlled level.

Another object of this invention is to provide pilot operated load responsive controls of a direction control valve, through which control of system load can- be either accomplished by variation in area of the orifice, between the valve control and a fluid motor, while the pressure differential across this orifice is maintained constant at a specific level, or by control of pressure differential, acting across this orifice, while the area of the orifice remains constant. It is a further object of this invention to provide pilot operated load responsive controls of a direction control valve, which permit variation in the controlled pressure differential across a metering orifice in response to an external control signal. It is a further object of this invention to provide pilot operated load responsive controls of a direction control valve, in which an external control signal, at a minimum force level^', can adjust and control the pressure differential, acting across a metering orifice of a load responsive direction control valve, while the system load is being controlled by variation in area of the metering orifice.

It is a further object of this invention to provide load responsive controls of direction control valve, which modify control signals, supplied to the pilot operated valve controls, to control the pressure differential across an orifice of a load responsive direction control valve.

It is a further object of this invention to provide load responsive controls of direction control valve, which modify control signals supplied at minimum energy level to the amplifying stage of the valve controls, to control pressure differential across an orifice of a load responsive direction control valve. Briefly the foregoing and other additional objects and advantages of this invention are accomplished by providing novel load responsive controls of a direction control valve, to throttle fluid supplied from the pump either in response to one control input, namely variation in the area of metering orifice, to control a contant pressure differential, at a preselected level between valve inlet pressure and the load pressure, or in response to another control input, namely modification in the pressure of control signal, to vary the level of the control differential between valve inlet pressure and the load pressure, while this control differential is automatically maintained constant at each controlled level by the valve controls receiving low energy control signals to their amplifying stage. In this way a load can be controlled in response to either input providing identical control performance, or the variable pressure differential control can be superimposed on the control action controlling a load by variation in the area of the metering orifice. Therefore this control system

OMPI_ lends itself very well to an application, in which a manual control input from an operator may be modified by an electronic logic circuit, or a micro-processor. 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 detailed description.

Description of the Drawings

Fig. 1 is a diagrammatic representation of a load responsive pilot operated throttling control for adjustment in the level of control differential from a certain preselected level to zero level, with fluid motor and system pump shown schematically;

Fig. 2 is a diagrammatic representation of a load responsive pilot operated throttling control for adjustment in the level of the control differential from a certain minimum preselected value up to maximum level, with fluid motor and system pump shown schematically;

Fig. 3 is a diagrammatic representation of another embodiment of the load responsive pilot operated control of Fig. 1, with fluid motor and system pump shown schematically;

Fig. 4 is a diagrammatic representation of still another embodiment of the load responsive pilot operated control of Fig. 1, with fluid motor and system pump shown schematically; Fig. 5 is a diagrammatic representation of a four way load responsive direction control valve using the control of Fig. 1 with system pump and reservoir shown schematically; Fig. 6 is a diagrammatic representation of manual control input into the load responsive controls of Figs. 1 to 5;

Fig. 7 is a diagrammatic representation of hydraulic control input into load responsive controls of Figs. 1 to 5;

Fig. 8 is a diagrammatic representation of electrohydraulic control input into load responsive controls of Figs. 1 to 5; Fig. 9 is a diagrammatic representation of an electromechanical control input into load responsive controls of Figs. 1 to 5; and

Fig. 10 is a diagrammatic representation of an electromechanical control input into load responsive system of Fig. 4.

Description of the Preferred Embodiments

Referring now to Fig. 1, the hydraulic system shown therein comprises a fluid pump 10, equipped with an output flow control 11, connected to a reservoir

12. The output flow control 11 regulates delivery of the pump 10 into a load responsive circuit, composed of a differential throttling control, generally designated as 13, regulating the level of the pressure differential developed across schematically shown variable orifice 14, interposed between the differential throttling control 13 and a fluid motor 15, operating load W. The pump 10 may be of a fixed or variable displacement type and may respond to an external or internal control signal. With the pump 10 being of a fixed displacement type, responding to an external control signal, the output flow control 11, in a well known manner, regulates delivery from pump to the load responsive circuit, by bypassing part of the pump flow to the system reservoir 12. The pump 10 of fixed displacement type can also respond to maximum pressure limiting signal, with the use of conventional relief valve, well known in the art. With the pump 10 being of variable displacement type and capable of responding to an external control signal, the output flow control 11, in a well known manner, regulates through a displacement changing mechanism delivery from the pump to load responsive circuit, by changing pump displacement. The pump 10 of variable displacement type can also respond to the maximum pressure limiting signal, such a control, well known in the art, being in the form of a conventional pressure compensator.

Discharge line 16 of pump 10 is connected through the differential throttling control, generally designated as 13, line 17, variable metering orifice 14 and line 18 to the fluid motor ^'15. The fluid motor 15 is also connected by lines 18 and 19 and orifice 20 with the differential throttling control 13.

The differential throttling control, generally designated as 13, composed of a throttling section, generally designated as 13a and a signal modifying section, generally designated as 13b, comprises a housing 21 having inlet chamber 22, an outlet chamber 23, a first control chamber 24 and an exhaust chamber 25, all of those chambers being connected by bore 26, slidably guiding a throttling spool 27. The throttling spool 27, equipped with lands 28 and 29 and stop 30, is provided with throttling slots 31, terminating in the cut-off edges 32, between the inlet chamber 22 and the outlet chamber 23. One end of the throttling spool 27 projects into the first control chamber 24, while the other end projects into the exhaust chamber 25 and is biased by a control spring 33. The first control chamber 24 is connected by passage 34 with annular space 35. Bore 36 connects annular space 35 with port 36 and a second control chamber 37 and axially guides a pilot valve spool 38. The pilot valve spool 38, equipped with a metering land 39 and land 40 which defines annular space 41, communicates with port 36 and projects into the second control chamber 37, where it engages a spring 42. The second control chamber 37 is connected through line 19 and orifice 20 to the fluid motor 15 and is also connected through port 43 with the supply chamber 44, connected by bore 45 with a third control chamber 46 and an exhaust chamber 47. Bore 45 slidably guides a control spool 48, equipped with land 49, provided with throttling slots 50 and positioned between the supply chamber 44 ^'and the third control chamber 46, a land 51 separating the supply chamber 44 and the exhaust chamber 47 and a flange 52. A spring 53 is interposed in the exhaust chamber 47 between the flange 52 of the control spool 48 and the housing 21. The exhaust chamber 47 connected by passage 54 with the exhaust chamber 25 and the third control chamber 46 are selectively interconnected by metering orifice created by a stem 55 guided in bore 56 and provided with metering slots 57. The stem 55 is connected to an actuator 58 responsive to external control signal 59. Exhaust chambers 47 and 25 connected by passage 54 are also connected by passage 60 with annular space 41 and leakage orifice 61.

Referring now to Fig. 2, the same components used in Fig. 1 are designated by the same numerals. The only difference between the load responsive controls of Figs. 1 and 2 is the phasing of the internal components of the differential control 13 and connections of individual ports or chambers with the fluid motor 15 and the pump 10. In both figures, in an identical way, the pump pressure is transmitted through discharge line 16, the inlet chamber 22, the outlet

OMΓI chamber 23, line 17 and variable orifice 14 to the fluid motor 15. Line 17 however, is connected in Fig. 2 through line 19 and orifice 20 to port 36 of the pilot valve spool 38. The second control chamber 37 is in direct communication through lines 62 and 19 with load pressure Pw of the fluid motor 15. The supply chamber 44 is connected through port 43 with port 36 and selectively interconnected with the third control chamber 46 by the control spool 48. Referring now to Fig. 3, the same components used in Figs. 1 and 2 are designated by the same numerals. The basic load responsive circuit of Fig. 3, with some of the circuit components including some of the internal components of differential throttling control, generally designated as 63, are the same as those of Figs. 1 and 2. The differential throttling control 63 is composed of a throttling section 63a, identical to throttling section 13a of Figs. 1 and 2, a flow control valve section 63b and a metering valve section 63c. The second control chamber 37 is connected by port 64 with a first pressure chamber 65, which in turn is connected through bore 66 with a second pressure chamber 67, which guides a stem 55, equipped with metering slots 68. The stem 55 is connected to the actuator 58, responsive to the external control signal 59. A load check 69 is interposed between the fluid motor 15 and the outlet chamber 23. The second control chamber 37 is also connected through port 70 with a third pressure chamber 71, connected by bore 72 with an exhaust^'chamber 73 which is connected through passage 74 with the exhaust chamber 25. Bore 72 axially guides a metering pin 75, provided with metering slot 76. The metering pin 75 is provided with a stop 77 and biased, towards position as shown, by a spring 78, contained in the exhaust chamber 73. Referring now to Fig. 4, the same components used in Figs. 1, 2 and 3 are designated by the same numerals. The basic load responsive circuit of Fig. 4 with some of the circuit components, including some of the internal components of differential throttling control, generally designated as 79, are the same as those of Figs. 1, 2 and 3. The second exhaust chamber 37 is connected by port 80 to the chamber 81 of the differential valve, generally designated as 82. The differential valve 82 comprises a coil 83, retained in the housing, which guides an armature 84 of a solenoid, generally designated as 85. The armature 84 is provided with conical surface 86, selectively engageable with sealing edge 87 of inlet port 88 and venting passage 89 terminating in bore 90, guiding a reaction pin 91. The coil 83 is connected by sealed connector 92 to outside of the housing 21, external control signal being applied to the sealed connector 92. The .second control chamber 37 is connected through leakage orifice 93 and passage 94 to the exhaust chamber 25 and the reservoir 12.

Referring now to Fig. 5 the same components used in Figs. 1, 2, 3 and 4 are designated by the same numerals. The differential throttling control 13 of Fig. 1 was interconnected in Fig. 5 into a four way valve assembly, generally designated as 95, which is basically equivalent to variable metering orifice 14 of Fig. 1. The four way valve assembly, generally designated as 95, comprises a housing 96 having a supply chamber 97, load chambers 98 and 99 and exhaust chambers 100 and 101, interconnected by bore 102, guiding a valve spool 103. The valve spool 103 is provided with lands 104, 105 and 106, throttling slots 107, 108, 109 and 110 and signal slots 111 and 112. The housing 96 is also provided with load sensing ports

OMPI 113 and 114, communicating through line 115 and orifice 20 to the second control chamber 37, of the differential throttling control 13.

Referring now to Fig. 6, the stem 55 of the actuator 58 of Figs. 1 to 5 is biased by a spring 116 towards position of zero orifice and is directly operated by a lever 117, which provides the external signal 59.

Referring now to Fig. 7, the stem 55 of the actuator 58 of Figs. 1 to 5, is biased by a spring 118 towards position of zero orifice and is directly operated by a piston 119. Fluid pressure is supplied to the piston 119 from a pressure generator 120, operated by a lever 121. Referring now to Fig. 8, the stem 55 of the actuator 58 of Figs. 1 to 5, is biased by a spring 122 towards position of zero orifice and is directly operated by a solenoid 123, connected by line to an input current control 124, operated by a lever 125 and supplied from an electrical supply source 126.

Referring now to Fig. 9, the stem 55 of the differential control, generally designated as 127, is biased by a spring 128 towards a position, where it isolates the third control chamber 46 from the exhaust chamber 47 and is controlled by a solenoid 129. The electrical control signal, amplified by amplifier 130, is transmitted from a logic circuit or a micro¬ processor 131, subjected to inputs 132, 133 and 134.

Referring now to Fig. 10, a logic circuit or a micro-processor 135, supplied with control signals 136, 137 and 138, transmits an external control signal to the differential valve 82 through an amplifier 139. Referring now to Fig. 1, the differential throttling control 13 is interposed between the pump 10 and the fluid motor 15 and controls the fluid flow and

OMPI WIPO pressure therebetween. The differential throttling control 13 is composed of the throttling section 13a and signal modifying section 13b. The throttling section 13a with its throttling spool 27 throttles with throttling slots 31 fluid flow from the inlet chamber 22, connected by discharge line 16 to the pump 10, to the outlet chamber 23, connected by line 17 and variable orifice 14 with the fluid motor 15, to automatically maintain a constant pressure differential across variable orifice 14. This control action is accomplished in the following way. Fluid from the outlet chamber 23 at P.. pressure, which is the pressure acting upstream of variable orifice 14, is transmitted through line 17 to port 36 where, reacting on the cross-sectional area of the pilot valve spool 38, generates a force tending to move the pilot valve spool 38 upward to connect P, pressure through annular space 35 and passage 34 to the first control chamber 24 and therefore increase the pressure level in the first control chamber 24. Fluid at load pressure

Pw, which is the pressure acting downstream of variable orifice 14, is transmitted through line 19 and orifice 20 to the second control chamber 37 where, reacting on the cross-sectional area of the pilot valve spool 38, it generates a force tending to move the pilot valve spool downward, to connect the reservoir pressure from annular space 41 to annular space 35, passage 34 and to the first control chamber 24 and therefore decrease the pressure level in the first control chamber 24. This force due to pressure in the second control chamber 37 is supplemented by the biasing force of the spring 42. Increase in pressure level in the first control chamber 24, above the level equivalent to preload of control spring 33, reacting on cross-sectional area of the throttling spool 27, will

y^S^

O PI generate a force tending to move the throttling spool 27 from right to left, in the direction of closing of the flow area through the throttling slots 31 and therefore in direction of increasing the throttling action of the throttling spool 27. Conversely, a decrease in pressure level in the first control chamber 24, below level equivalent to preload of control spring 33, will result in the control spring 33 moving the throttling spool 27 from left to right, in the direction of increasing the flow area through the throttling slots 31 and therefore in direction of decreasing the throttling action of the throttling spool 27. Therefore by regulating pressure level in the first control chamber 24 the pilot valve spool 38 will control the throttling action of the throttling spool 27 and consequently the pressure drop between the inlet chamber 22 subjected to Pp pressure and the outlet chamber 23 subjected to P, pressure. Assume that the stem 55 is in the position as shown in Fig. 1, isolating the third control chamber 46 from the exhaust chamber 47 and therefore making the signal modifying section 13b inactive. The pilot valve spool 38, subjected to P, and P pressures and the biasing force of spring 42 will reach a modulating position, in which by throttling action of metering land 39 will regulate the pressure in the first control chamber 24 and therefore the throttling action of the throttling spool 27 to throttle the pump pressure Pp to a level of P, pressure which is higher, by a constant pressure differential ΔP, than P„ pressure and equal to the quotient of the biasing force of spring 42 and the cross-sectional area of the pilot valve spool 38. In this way the pilot valve spool 38, subjected to low energy pressure signals, will act as an amplifying stage using the energy derived from the pump 10 to control the position and therefore the throttling action of the throttling spool 27. Leakage orifice 61, connecting the first control chamber 24 through passage 60 and the exhaust chamber 24 to the reservoir 12, is used, in a well known manner, to increase the stability of the pilot valve spool 38. If P- pressure is equal to Pw pressure which is the case when the stem 55 is in the position, as shown in Fig. 1, the throttling section 13a, by throttling fluid flow from the inlet chamber 22 to the outlet chamber 23, will automatically maintain a constant pressure differential ΔP between the outlet chamber 23 and the second control chamber 37 and with Δ Py becoming ΔP, will also maintain a constant pressure differential across variable orifice 14. With constant pressure differential, acting across an orifice, the flow through an orifice will be proportional to the area of the orifice and independent of pressure in the fluid motor. Therefore by varying the area of variable orifice 14, the fluid flow to the fluid motor 15 and velocity of the load W can be controlled, each specific area of variable orifice 14 corresponding to a specific velocity of load W, which will remain constant, irrespective of the variation in the magnitude of the load W. In the arrangement of Fig. 1 the relationship between load pressure Pw and signal pressure P- is controlled by the signal modifying section, generally designated as 13b, and orifice 20. Assume that the stem 55, positioned by the actuator 58 in response to external control signal 59, as shown in Fig. 1, blocks completely metering orifice through metering slots 57, isolating the third control chamber 46 from the exhaust chamber 47. The control spool 48 with its land 49, protruding into the third control chamber 46, will generate pressure in the third control chamber 46, equivalent to the preload of the spring 53. Displacement of the stem 55 to the left will move metering slots 57 out of bore 56, creating an orifice area, through which fluid flow will take place from the third control chamber 46 to the exhaust chamber 47. The control spool 48, biased by the spring 53, will move from right to left, connecting by throttling slots 50 the supply chamber 44 with the third control chamber 46. Rising pressure in the third control chamber 46, reacting on cross-sectional area of control spool 48, will move it back into a modulating position, in which sufficient flow of pressure fluid will be throttled from the supply chamber 44 to the third control chamber 46, to maintain the third control chamber 46 at a constant pressure, equivalent to preload in the spring 53. When displacing metering slots 57, in respect to bore 56, area of metering orifice between the third control chamber 46 and the exhaust chamber 47 will be varied. Since constant pressure differential is automatically maintained between the exhaust chamber 47 and the third control chamber 46 and therefore across the metering slots 57, by the control spool 48, each specific area of metering slots 57 will correspond to a specific constant flow level from the third control chamber 46 to the exhaust chamber 47 and from the supply chamber 44 to the third control chamber 46, irrespective of the magnitude of the pressure in the supply chamber 44. Therefore, each specific position of stem 55, within the zone of metering slots 57, will correspond to a specific flow level and therefore a specific pressure drop ΔPx through the fixed orifice 20, irrespective of the magnitude of the load pressure Pw. When referring to Fig. 1 it can be seen that P, - Pw =Δ Py, Pi - P₂ = P, maintained constant by the throttling section 13a and

O PI WIFO Pw - P« = ΔPx. From the above equations, when substituting and eliminating P-, and P_ a basic relationship of Δ Py = Δ P - Δ Px is obtained. Since Δ Px can be varied and maintained constant at any level by the signal modifying section 13b, so can ΔPy, acting across variable orifice 14, be varied and maintained constant at any level. Therefore with any specific constant area of variable orifice 14, in response to control signal 59, pressure differential Δ Py can be varied from maximum to zero, each specific level of Δ Py being automatically controlled constant, irrespective of variation in the load pressure Pw. Therefore, for each specific area of variable orifice 14 the pressure differential, acting across orifice 14 and the flow through orifice 14 can be controlled from maximum to minimum by the signal modifying section 13b, each flow level automatically being controlled constant by the differential throttling control 13, irrespective of the variation in the load pressure Pw. From inspection of the basic equation Δ Py = Δ P - Δ Px it becomes apparent that with ΔPx = 0, ΔPy = Δ P and that the system will revert to the mode of operation of conventional load responsive system, with maximum constant Δ P of the differential throttling control 13. When Δ Px = ΔP, Δ Py becomes zero, outlet pressure from the differential throttling control 13 P.. will be equal to load pressure Pw and the flow through variable orifice 14 will become zero. With Δ Px larger than Δ P, pressure P-, will become smaller than load pressure Pw and the load check 18a will seat.

In the load responsive system of Fig. 1 for each specific value of Δ Py, maintained constant by the modifying section 13b through the throttling section 13a of the differential control 13, the area of

O PI variable orifice 14 can be varied, each area corresponding to a specific constant flow into the fluid motor 15, irrespective of the variation in the magnitude in the load pressure Pw. Conversely, for each specific area of the variable orifice 14 pressure differential Δ Py, acting across orifice 14, can be varied by the signal modifying section 13b, through the throttling section 13a of the differential throttling control 13, each specific pressure differential Δ Py corresponding to a specific constant flow into the fluid motor 15, irrespective of the variation in the magnitude of the load pressure Pw. Therefore fluid flow into fluid motor 15 can be controlled either by variation in area of variable orifice 14, or by variation in pressure differential Δ y each of those control methods displaying identical control characteristics and controlling flow, which is independent of the magnitude of the load pressure. . Action of one control can be superimposed on the action of the other, providing a unique system, in which, for example, a command signal from the operator, through the use of variable orifice 14 can be corrected by signal 59 from a computing device, acting through the signal modifying section 13b. Referring now to Fig. 2, the signal modifying section 13b, is identical to the signal modifying section 13b of Fig. 1 and performs in an identical way, by modifying a control signal transmitted to the throttling section 13a. The throttling section 13a of Fig. 2 is identical to the throttling section 13a of Fig. 1. However, the signal modifying section 13b of Fig. 2 modifies the control signal from outlet chamber 23 and therefore from upstream of the variable orifice 14, instead of modifying the control signal of load pressure Pw, as shown in the system of Fig. 1. In Fig. 2 the load pressure signal Pw is transmitted directly from fluid motor 15 to the second control chamber 37, of the throttling section 13a. Then, as can be seen in Fig. 2, P- - Pw = Δ Py, P, - P₂ = Δ x and P„ - Pw = P, which, in a manner as previously described, is the basic system differential and is maintained constant by the throttling section 13a of the differential throttling control 13. From the above equations, when substituting and eliminating P, and ₂ the basic relationship of Δ Py = ΔP + Δ Px can be obtained. Since Δ x can be varied and maintained constant at any level, so can Δ Py, acting across variable orifice 14 be varied and maintained constant at any level. From inspection of the basic equation Δ Py = ΔP + Δ Px it becomes apparent that with ΔPx = 0, ΔPy = ΔP and that the system will revert to the mode of operation of conventional load responsive system, with minimum constant Δ P equal to pressure differential of the throttling section 13a. Any value of ΔPx, other than zero, will increase the pressure differential Δ Py, acting across variable orifice 14 above the level of constant pressure ■ differential Δ P of the throttling section 13a. Therefore, the load responsive control arrangement of Fig. 1 will control Δ Py in a range between ΔP and zero, while the load responsive arrangement of Fig. 2 will control Δ Py in a range above the level of constant pressure differential Δ P of the throttling section 13a. Referring now to Fig. 3, the load responsive system is identical to the load responsive system of Fig. 1 with the exception of the flow control valve section 63b and the metering valve section 63c, which, when combined together are equivalent to the signal modifying section 13b of Fig. 1 and perform in a very

_O ?I_ similar way. The throttling section 63a of Fig. 3 is identical to the throttling section 13a of Fig. 1. It is apparent that the differential throttling control 63 of Fig. 3 performs in an identical way as the differential throttling control 13 of Fig. 1. The flow control valve section 63b of the differential throttling control 63 in the housing 21, is provided with bore 72, guiding the metering pin 75, which is subjected to pressure in the third pressure chamber 71, which is connected by port 64 with the second control chamber 37, to the reservoir pressure in the exhaust chamber 73 and to the biasing force of the spring 78. Subjected to pressure in the third pressure chamber 71 the metering pin 75 will move from left to right, each specific pressure level corresponding to a specific position of metering pin 75, in respect to the housing 21 and also corresponding to the specific biasing force of spring 78. Each specific position of metering pin 75, in respect to the housing 21, will correspond to a specific flow area of metering slot 76, interconnecting the third pressure chamber 71 with the exhaust chamber 73. The shape of metering slot 76 and the characteristics of the biasing spring 78 are so selected, that variation in effective orifice area of metering slot 76, in respect to pressure in the third pressure chamber 71, will provide a relatively constant flow from the third pressure chamber 71 to the exhaust chamber 73. To obtain special control characteristics of the load responsive control the shape of the metering slot 76 may be so selected, that any desired relationship between the flow from the third pressure chamber 71 and its pressure level can be obtained. Assume that the flow control valve section 63b provides a constant flow from the third pressure chamber 71 and therefore from the second control chamber 37, irrespective of its pressure level. Then in a well known manner, the flow control valve section 63b could be substituted by a conventional flow control valve, well known in the art. Constant flow to the third pressure chamber 71 is supplied from fluid motor 15 through the metering valve section 63c, second control chamber 37 and port 70. The metering valve section 63c, upstream of the flow control valve section 63b is provided with a bore 66, guiding the stem 55, provided with metering slots 68. Displacement of metering slots 68 past bore 66 creates an orifice, the effective area of which can be varied by positioning of stem 55 by the actuator 58, in response to external control signal 59. With stem 55 engaging bore 66 the flow area of the metering valve section 63c becomes zero. Therefore, in response to external control signal 59, the effective flow area through the metering valve section 63c can be varied from zero to a selected maximum value. Since the flow through the metering valve section 63c is maintained constant by the flow control valve 63b, each specific area of flow through the metering valve section 63c, in a well known manner, will correspond to a specific constant pressure drop Δ Px, irrespective of the variation in the load pressure Pw. Therefore the load pressure signal can be modified in the second control chamber 37 of the throttling section 63a, each value of pressure drop ΔPx, maintained constant by the flow control valve section 63b, corresponds to a specific value of pressure differential Δ Py, fol- lowing the basic relationship of Δ Py = Δ P - Δ Px. Therefore the control characteristics of the load responsive control of Fig. 3 will be identical to those described, when referring to Fig. 1, the pressure differential Δ Py being varied and maintained constant, at each specific level, by the control action

of the flow control valve section 63b and the metering valve section 63c, in response to external control signal 59, between maximum value equal to ΔP and zero. In a manner as previously described the shape of metering slot 76 and the biasing force char¬ acteristics of spring 78 can be so selected, that any desired relationship between pressure in the third pressure chamber 71 and the fluid flow through the metering valve section 63c can be obtained. Assume that in response to a specific external control signal 59 a specific flow area was created through the metering valve section 63c. Then controlled increase in flow through the metering valve section 63c, with increase in load pressure, will proportionally increase the pressure differential ΔPx and therefore pro¬ portionally decrease the pressure differential ΔPy, effectively decreasing the gain of the load responsive control with increase in the load pressure. Conversely, a controlled decrease of flow through the specific orifice area of the metering valve section 63c, with increase in the load pressure will proportionally decrease the pressure differential Δ Px and therefore proportionally increase the pressure differential ΔPy, effectively increasing the gain of the load responsive control, with increase in the load pressure. As is well known in the art, the stability margin of most fluid flow and pressure controllers decreases with increase in system pressure. Therefore the capability of adjusting the system gain, in respect to system pressure, is of primary importance. With the flow control valve section 63b the rate of change of pressure differential ΔPy, in respect to load pressure, does not have to be constant and can be varied in any desired way.

C. ?I ^Wi?ϋ Referring now to Fig. 4, the load responsive system is similar to that of Fig. 1. The throttling control 79, of the differential throttling control of Fig. 4 is identical to differential throttling section 13a of Fig. 1. However, the differential .valve 82 is different from the signal modifying section 13b of Fig. 1, although it performs the same function and provides identical performance. The differential valve, gener¬ ally designated as 82, contains the solenoid, generally designated as 85, which consists of coil 83, secured in the housing 21 and the armature 84, slidably guided in the coil 83. The armature 84 is provided with conical surface 86, which in cooperation with sealing edge 87, regulates the pressure differential Δ Px between inlet port 88 and port 80. A comparatively weak unnumbered spring can be interposed between the armature 84 and the housing 21, to permit a back flow under deenergized condition of the coil 83 from port 80 to inlet port 88. This feature may be of importance, when using a shuttle valve logic system instead of a check valve system, for trans ittal of control signals. The sealed connector 92, in the housing 21, well known in the art, connects the coil 83 with external terminals, to which the external signal 59 can be applied. A solenoid is an electro-mechanical device, using the principle of electromagnetics, to produce output forces from electrical input signals. The force developed on the solenoid armature 84 is a function of the input current. As the current is applied to the coil 83, each specific current level will correspond to a specific force level, transmitted to the armature. Therefore, the contact force between the conical surface 86 of the armature 84 and sealing edge 87 of housing 21 will vary and be controlled by the input current. This arrangement will then be equivalent to a

0" :?ι type of differential pressure throttling valve, varying automatically the pressure differential Px between inlet port 88 and second control chamber 37, in proportion to the force developed in the armature 84, in respect to the area enclosed by the sealing edge 87 and therefore proportional to the external signal 59, of the input current supplied to the solenoid 85. The pressure forces acting on the armature 84, within the housing 21, are completely balanced with the exception of the pressure force due to the pressure differ¬ ential Δ Px acting on the enclosed area of sealing edge 87. This force is partially balanced by the reaction force, developed on the cross-sectional area of the reaction pin 91, guided in a bore 90, which is connected through venting passage 89 with inlet port 88. The cross-sectional area of the reaction pin 91 must always be smaller than the area enclosed by sealing edge 87, so that a positive force, due to the pressure differential ΔP , opposes the force developed by the solenoid 85. The reaction pin 91 permits use of a larger inlet port 88, while also permitting a very significant reduction in the size of solenoid 85, also permitting the solenoid 85 to work in the higher range of Δ Px. The second control chamber 37 is connected by leakage orifice 93 and passage 94 with the system reservoir.

Referring now to Fig. 5, the load responsive system is identical to that as shown in Fig. 1 with identical differential throttling controls being used, but the variable orifice 14 of Fig. 1 was substituted in Fig. 5 by a load responsive four way type direction control valve, generally designated as 95. The performance of the control embodiments of Figs. 1 and 5 is identical, the only difference being the con- struction of the variable orifice. The differential

PI throttling control and specifically the second load chamber 37 are connected through orifice 20 and line 115 with the load sensing ports 113 and 114 of four way valve 95. With the valve spool 103 in its neutral position, as shown in Fig. 5, load pressure sensing ports 113 and 114 are blocked by the land 105 and therefore effectively isolated from load pressure, existing in load chamber 98 or 99. Under those conditions, in a well known manner, the differential throttling control 13, automatically maintains minimum pressure in the supply chamber 97 and equal to Δ P of the throttling section 13a. Displacement of the valve spool 103 from its neutral position in either direction, first connects with signal slot 111 or 112 load chamber 99 or 98 with load pressure sensing port 113 or 114, while load chambers 99 and 98 are still isolated by the valve spool 103 from the supply chamber 91 and exhaust chambers 100 and 101. Then the load pressure signal is transmitted through load pressure sensing port 113 or 114, line 115 and orifice 20 to the second control chamber 37, permitting the differential throttling control 13 to react, before metering orifice is open to the load chamber 99 or 98. Further displacement of valve spool 103, in either direction, will create, in a well known manner, through metering slot 107 or 108 a metering orifice between one of the load chambers and the supply chamber 97, while connecting the other load chamber, through metering slot 109 or 110 with the exhaust chambers, in turn connected to system reservoir. The metering orifice can be varied by displacement of valve spool 103 each position corresponding to a specific flow level into one of the load chambers, irrespective of the magnitude of the load controlled by four way valve assembly 95. Upon this control, in a manner as previously described when referring to Fig. 1, can be superimposed the control action of the signal modifying section 13b. With valve spool 103 displaced to any specific position, corresponding to any specific area of metering orifice, the flow into load chambers can be proportionally controlled by the differential throttling control 13 with its signal modifying section 13b, each value of pressure differential Δ Py being automatically maintained at a constant level by the throttling section 13a and corresponding to a specific flow level into load chambers irrespective of the magnitude of the load controlled by the four way valve assembly 95.

Referring now to Fig. 6, the stem 55 of the actuator 58 of Figs. 1 to 5 is biased by spring 118 towards position of zero orifice and is directly operated by a lever 117, which provides the external signal 59 in the form of manual input.

Referring now to Fig. 7, the stem 55 of actuator 58 of Figs. 1 to 5 is biased by spring 118 towards position of zero orifice and is directly operated by a piston 119. Fluid pressure is supplied, in a well known manner, to the piston 119 from a pressure generator 120, operated by a lever 121. Therefore the arrangement of Fig. 7 provides the external signal 59 in the form of a fluid pressure signal.

Referring now to Fig. 8, the stem 55 of the actuator 58 of Figs. 1 to 5 is biased by a spring 122 towards position of zero orifice and is directly operated, in a well known manner, by a solenoid 123, connected by a line to an input current control 124, operated by a lever 125 and supplied from an electrical power source 126. Therefore the arrangement of Fig. 8 supplies the external signal 59 in the form of an

OMPI electric current, proportional to displacement of lever 125.

Referring now to Fig. 9, the stem 55 of the differential control 127 is biased by a spring 128 towards a position, where it isolates the third control chamber 46 from the exhaust chamber 47. The stem 55 is completely pressure balanced, can be made to operate through a very small stroke and controls such low flows, at such low pressures, that the influence of flow forces is negligible. In any event, if the area of metering slots 57 is so selected, that it provides a linear function in respect to displacement of the stem 55 and a constant pressure is maintained in front of the orifice, the flow forces will also be linear and will add to the spring force, changing slightly the combined rate of the spring. The stem 55 is directly coupled to a solenoid 129. A solenoid is an electromechanical device using the principle of electromagnetics, to produce output forces from electrical input signals. The position of solenoid armature, when biased by a spring, is a function of the input current. As the current is applied to the coil, the resulting magnetic forces generated move the armature from its deenergized position to its energized position. When biased by a spring, for each specific current level there is a corresponding particular position, which the solenoid will attain. As the current is varied from zero to maximum rating, the armature will move one way from a fully retracted to a fully extended position in a predictable^'fashion, depending on the specific level of current at any one instant. Since the forces, developed by solenoid 129 are very small, so is the input current, which is controlled by a logic circuit or a micro-processor 131. The micro-processor 131 will then, in response to

OMPI different types of transducers either directly control the system load, in respect to speed, force and position, or can superimpose its action upon the control function of an operator, to perform required work in the minimum time, with a minimum amount of energy, within the maximum capability of the structure of the machine and within the envelope of its horsepower.

Referring now to Fig. 10, the control signal from the logic circuit, or the micro-processor 135, in a similar way as described in Fig. 9, is directly transmitted through the amplifier 139 to the differential pressure control 82, where through a solenoid and throttling valve combination, in a manner as previously described, when referring to Fig. 4, regulates the pressure differential in response to input current.

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.

C PI

Claims

AMENDED CLAIMS

(received by the International Bureau on 30 December 1981 (30.12.81))

39. (New) A valve assembly comprising a housing (21) having an inlet chamber (22) connected to a pump (10) , a supply chamber (23) connected to a fluid motor (15) and exhaust means (12) , control orifice means (14,107,108) interposed between said supply chamber (23) and said fluid motor (15) , first valve means (13a,63a) having fluid throttling means (31,32) between said inlet chamber (22) and said supply chamber (23) controllable by a pilot valve means (38,42) and operable to throttle fluid flow from said inlet chamber (22) to said supply chamber (23) to maintain a constant pressure differential at a preselected constant level across said pilot valve means (38,42) and to maintain a constant pressure differential across said control orifice means (14, 107,108) .

40. (New) A valve assembly as set forth in claim 39 wherein said control orifice means (14,107,108) has variable area orifice means (V) .

41. (New) A valve assembly as set forth in claim 39 wherein said control orifice means (14,107,108) has means (103) operable to selectively interconnect said fluid motor (15) with said supply chamber (23) and with said exhaust means (100,101,12) .

42. (New) A valve assembly as set forth in claim 39 wherein second valve means (13b,63b,82) has means (48,55,75,84,83) operable through said first valve means (13a,63a) to vary the level of said constant pressure differential across said control orifice means (14,107,108) while said pressure differential across said pilot valve means (38,42) remains con¬ stant at said constant predetermined level.

43. (New) A valve assembly as set forth in claim 42 wherein said second valve means (13b) includes constant pressure reducing means (48,53), orifice means (20) upstream of said constant pressure reduc¬ ing means (48,53) and flow orifice means (57) down- stream of said constant pressure reducing means (48,53) .

44. (New) A valve assembly as set forth in claim 42 wherein said second valve means (63b) includes constant flow control means (75,76,78) connected to said exhaust means (73,74,25,12) and variable control orifice means (55,68) upstream of said con¬ stant flow control means (75,76,78).

45. (New) A valve assembly as set forth in claim 42 wherein said second valve means (82) includes fluid throttling means (83,84) and orifice means

(93) downstream of said fluid throttling means (83,84) communicable with exhaust means (12) .

46. (New) A valve assembly as set forth in claim 42 wherein said second valve means (13b,63b,82) has means (18,19,20,37,48,55) to vary the level of said constant pressure differential across said control orifice means (14) below the level of said pressure differential across said pilot valve means (38,42) maintained constant at said constant predetermined level.

-gUREAlT 0Λ.PI 47. (New) A valve assembly as set forth in claim 42 wherein said second valve means (13b) has means (20,36,43,48,55) to vary the level of said constant pressure differential across said control orifice means (14) above the level of said pressure differ¬ ential across said pilot valve means (38,42) main¬ tained constant at said constant predetermined level.

48. (New) A valve assembly as set forth in claim 42 wherein said second valve means (13b, 3b,82) has means (58) responsive to an external control signal

(59).

49. (New) A valve assembly as set forth in claim 48 wherein said means (58) responsive to an external control signal (59) includes mechanical actuating means (117) .

50. (New) A valve assembly as set forth in claim 48 wherein said means (58) responsive to an external control signal (59) includes fluid pressure actuating means (120,121).

51. (New) A valve assembly as set forth in claim 48 wherein said means (58) responsive to an external control signal (5£L- ncludes electro-mechanical actuating means (123,124) .

52. (New) A valve assembly as set forth in claim

48 wherein said means (127) responsive to an external control signal (132,133,134) includes logic circuit means (131) , amplifying means (130) and electro-mech¬ anical means (129) .

O PI