EP0179769A1 - Load responsive fluid control valve - Google Patents

Abstract

A load-sensitive valve (10) is capable of limiting, by the regulating action of its positive load compensator (40), the maximum pressure developed in the actuating device (11) which it controls to a level pressure preselected specific, while transmitting a charge pressure signal, limited to a specific maximum level, to the control (14) sensitive to the load of the pump (12) of the system.

Description

Description

Load Responsive Fluid Control Valve

Background of the Invention

This invention relates generally to direction and flow control valves capable of proportionally controlling a number of loads under positive and negative load conditions. In more particular aspects this invention relates to fluid control valves provided with positive and negative load compensation.

In still more particular aspects this invention relates to fluid control valves provided with a positive load compensator, throttling action of which limits maximum pressure supplied from the pump to the fluid actuator, to a certain predetermined level.

In still more particular aspects this invention relates to modulation of control pressure signals, transmitted from the actuators of the system to the valve compensating controls and to the load responsive control of the system pump, to limit the level of maximum load pressure in each of the actuators to a different preselected system level, while limiting the maximum system pressure, through the load responsive pump control, to the level, as dictated by any specific combination of loads being controlled at a time.

Closed center fluid control valves, pressure compensated for control of positive and negative loads, are desirable for a number of reasons. They permit load control with reduced power losses and therefore increased system efficiency. Such fluid control valves are shown in my patent 4,362,087 issued December 7, 1982. However, with such valves all of the system

OMPI actuators, especially when controlling inertia type loads, can be subjected to extremely high pressures, or at least to system pressure, equivalent to the pressure setting of the maximum pump pressure limiting control. This disadvantage is overcome by the control valve, disclosed in U.S. patent 3,592,216 issued to McMillen and U.S. patent 3,934,742 issued to Tennis, in which the load pressure signal, transmitted to the load sensing compensator, is limited to a certain predetermined maximum value, thus limiting the maximum inlet pressure to the actuator, while the load sensing pump control responds directly to an unmodified control signal, introducing high throttling losses in control of system loads. This disadvantage can be overcome in part by the control valve disclosed in U.S. patent 3,987,623 issued to Blanchetta, in which the load pressure signal, transmitted to load sensing compensator and to the load sensing pump control is limited to the same certain predetermined maximum pressure level.

It is very desirable to limit the load pressure signals transmitted to the system- pump to any specific desired different levels for optimum efficiency and performance of the system, while not affecting the load pressure signals transmitted to load sensing compensators, so that such signals can be maintained at their unmodified levels. Then the load pressure signals, transmitted to the load sensing compensators, can be selectively limited to different levels, either as dictated by the optimum maximum pressure levels, dictated by the structural limitations of the actuators, or by the optimum performance of the controls.

OMPI Summary of the Invention

It is therefore the principal object of this invention to limit the maximum pressure that can be supplied from the system pump to a specific actuator to a specific maximum level, as established by the structural integrity of the actuator.

Another object of this invention is to provide a pressure compensated valve, which would automatically limit the maximum pump pressure transmitted to an actuator through the throttling action of its positive load compensator.

It is a further object of this invention to individually limit the maximum pressure level of the load pressure signal transmitted from the actuator controlling a load to the load responsive pump control, to establish maximum discharge pressure level of the pump.

It is a further object of this invention to transmit load pressure signals, -limited to specific maximum levels, through a check valve logic system, to the load responsive pump control, to limit its controlled maximum discharge pressure to a level, as dictated by any specific combination of loads being controlled at the time. It is a further object of this invention to transmit load pressure signals, limited to a first specific maximum level, to the positive load compensator, to limit the maximum pressure that can be transmitted to the actuator from the system pump and to transmit load pressure signals, limited to a second specific maximum level, to the load responsive pump control, to limit the maximum discharge pressure of the pump.

Briefly the foregoing and other additional objects and advantages of this invention are accomplished by providing novel control of maximum pump pressure, transmitted to the actuator, through the throttling action of the positive load compensator, control action of which is dictated by modification of the transmitted load pressure signals. In turn modified load pressure signals limit, through the load responsive pump control, the maximum discharge pressure of the pump.

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 Drawing The figure is a longitudinal sectional view of an embodiment of a flow control valve in a load responsive system provided with a single positive and negative load compensator, also showing a longitudinal sectional view of an embodiment of a pilot valve amplifying stage controlling the compensator, together with the schematically shown signal modifying control and pump flow control with system lines, second flow., control valve, system actuator, system pump and system reservoir shown diagrammaticall .

Description of the Preferred Embodiment

Referring now to the single drawing, an embodiment of a flow control valve, generally designated as 10, is shown interposed between diagrammatically shown fluid motor 11 driving load and a pump 12, of a fixed displacement or variable displacement type, driven by a prime mover, not shown. Fluid flow from the pump 12 to flow control valve 10 and circuits of diagrammatically shown flow control valves 13 and 13a are regulated by pump flow control.

^

O H generally designated as 14. If pump 12 is of a fixed displacement type, pump flow control 14 is a differential pressure relief valve, which, in a well known manner, by bypassing fluid from pump 12 to a reservoir 15, maintains discharge pressure of pump 12 at a level, higher by a constant pressure differential, than load pressure developed in fluid motor 11. If pump 12 is of a variable displacement type, pump flow control 14 is a differential pressure compensator, well known in the art, which by changing displacement of pump 12, maintains discharge pressure of pump 12 at a level, higher by a constant pressure differential than load pressure developed in fluid motor 11.

The flow control valve 10 is of a fourway type and has a housing 16 provided with a bore 17, axially guiding a valve spool 18. The valve spool 18 is equipped with lands- 19, 20 and 21, which in neutral position of the valve spool 18, as shown in the drawing, isolate a fluid supply chamber 22, load chambers 23 and 24 and outlet chambers 25 and 26.

Lands 19, 20 and 21, of valve spool 18, are provided with metering slots 27, 28, 29 and 30 and timing slots 31, 32, 33 and 34. Negative load sensing ports 35 and

36 are positioned between load chambers 23 and 24 and outlet chambers 26 and 25. Positive load sensing ports

37 and 38 are located between supply chamber 22 and load chambers 23 and 24. Negative load throttling slots 39, of control spool 40, equipped with throttling edges 41, connect the outlet chamber 26 with an exhaust chamber 42, which in turn is connected to reservoir 15, while negative load throttling slots 39a, equipped with throttling edges 41a, connect the outlet chamber 25 with an exhaust chamber 42a, which in turn is connected to reservoir 15.

OMPI The pump 12, through its discharge .line 43, is connected to an inlet chamber 44. The inlet chamber 44 is connected through positive load throttling slots 45, on control spool 40, provided with throttling edges 46, with the fluid supply chamber 22. Bore 47 axially guides the control spool 40, which is biased by control spring 48, contained in control space 49, towards position as shown. The control spool 40 at one end projects into control space 49, the other end projecting into chamber 50, connected to the reservoir 15. A pilot valve assembly, generally designated as 51, comprises a housing 52, provided with a bore 53, slidably guiding spool 54 and free floating piston 55. The spool 54 is provided with lands 56, 57 and 58, defining annular spaces 59 and 60. Annular space 61 is provided within the housing 52 and communicates directly with bore 53. The free floating piston 55 is provided with a land 62, which defines annular spaces 63 and 64 and is provided with extension 65, selectively engageable with land 58 of the spool 54.

The spool 54 at one end projects into control space 66 and engages, with land 56 and spring retainer 67,_ a pilot valve spring 68. Control space 66 communicates through line 69 with check valves 71b, 71 and 71a. The check valve 71b is connected by passage 72 with positive load sensing ports 37 and 38. The check valves 71 and 71a communicate through line 73 with the outlet chambers 25 and 26. Annular space 61, of the pilot valve assembly 51, communicates through line 74 with control space 49 and also communicates through leakage orifice 75, with annular space 60, which in turn is connected to reservoir 15. Annular space 59 communicates through lines 76 with discharge line 43. Annular space 64 is connected by line 81 with the supply chamber 22. Annular space 63 is connected by

3 PI line 82 and passage 83 with negative load sensing ports 36 and 35. Positive load sensing ports 37 and 38 are connected through passage 72, line 84 and a check valve 85 and a signal line 86 with the pump flow control 14. Positive load sensing ports 37 and 38 are also connected through passage 72, line 84, a control module, generally designated as 78, check valve 71b and line 69 with control space 66. The control module 78 is provided with a signal throttling variable orifice 79, the throttling action of which is controlled by the force developed on piston 80 by pressure developed in positive load sensing port 37 or 38 and conducted by line 77, against biasing force of spring 80a. Control space 66 is connected through a leakage device 87 with the reservoir 15. Leakage device 87 may be of a straight leakage orifice type, or may be a flow control device, passing a constant flow from control space 66 to the reservoir 15. The leakage device 87 comprises a housing 88, provided with bore 89 guiding a spool 90, which defines spaces 91, 92 and 93. The spool 90 is provided with throttling slots 94, leakage orifice 95 and is biased by a spring 96. The chambers 23 and 24 are connected, for one way fluid flow, by check valves 97 and 98, to schematically shown system reservoir 15, which also might be a pressurized exhaust manifold of the entire control system, as shown in the drawing. Line 84 is provided with fixed orifice 99 down stream of which is connected by line 100 with a pressure limiting module generally designated as 101. The pressure limiting module 101 in the form of a signal pressure relief valve consists of a housing 102 provided with port 103, engaging poppet 104 biased by a spring 105. Internal space of the housing 102 is directly vented to the system reservoir 15. The load pressure signals of flow control valves 13 and 13a are phased by check valves 106 and 107 to line 86 connected with pump flow control 14. The pump flow control, generally designated as 14, is provided with an actuator 108, biased by spring 109, directly operates flow changing mechanism 110 of the pump 12. If the pump 12 is of a variable displacement type the flow changing mechanism 110 becomes a variable displacement changing mechanism, well known in the art. If the pump 12 is of a fixed displacement type the flow changing mechanism 10 becomes a bypass valve, well known in the art. The actuator 108 is subjected to the control pressure in space 111, which is controlled by pilot valve spool 112. The pilot valve spool 112 is provided with lands 113 and 114 defining spaces 115 and 116. Space 115 communicates directly with the system reservoir 15. Space 116 is directly connected through line 43 with the discharge pressure of the pump 12. The land 114 projects directly into control space and is biased by spring 118. Control space 117 is directly connected to the check valve logic system transmitting positive load pressure signal by line 86.

The preferable sequencing of lands and slots of valve spool 18 is such, that when displaced in either direction from its neutral position, as shown, one of the chambers 23 or 24 is connected by timing slot 32 or 33 to the positive load sensing port 37 or 38, while the other load chamber is simultaneously connected by timing slot 31 or 34 with negative load sensing port 35 or 36, the load chamber 23 or 24 being isolated from the supply chamber 22 and outlet chambers 25 and 26. Further displacement of valve spool 18 from its neutral position connects load chamber 23 or 24 through metering slot 28 or 29 with the supply chamber 22, while simultaneously connecting the other load chamber through metering slot 27 or 30 with outlet chamber 25 or 26. As previously described the pump flow control 14, in a well known manner, will regulate fluid flow, delivered from pump 12, to discharge line 43, to maintain the pressure in discharge line 43 higher, by a constant pressure differential, than the highest load pressure signal transmitted through the check valve system to signal line 86. Therefore, with the valve spool 18 of flow control valve 10, in its neutral position blocking positive load sensing ports 37 and 38, signal pressure input to pump flow control 14 from signal line 86 will be at minimum pressure level, corresponding with the minimum standby pressure of the pump 12.

Assume that the load chamber 23 is subjected to a positive load and that the control pressure differential of the pilot valve assembly 51 is higher than the control pressure differential of the pump flow control 14. The pilot valve assembly 51 is shown in the drawing with the spool 54 in its equilibrium modulating position and with land 57 blocking the annular space 61. With the control system at rest the pilot valve spring 68 will move the spool 54 all the way to the left, connecting annular space 60 with annular space 61 and therefore connecting control space 49 with system reservoir. Under those conditions the control spool 40 will be maintained by the control spring 48 in the position as shown. The initial displacement of the valve spool 18 to the right will connect, in a manner as previously described, the load chamber 23, subje.cted to positive load pressure, with positive load sensing port 37, while also connecting the load chamber 24 with negative load sensing port 35. The positive load pressure signal from positive load sensing port 37 will be transmitted through passage 72, line 84, check valve 85 and signal line 86 to the pump flow control 14 and, in a manner as previously described, will raise the discharge pressure of the pump 12 to a level, higher by a constant pressure differential, than the positive load pressure existing in the load chamber 23.

Further displacement of the valve spool 18 to the right will create a metering orifice through metering slot 29, between the load chamber 23 and the supply chamber 22, while also creating through metering slot 27 a similar metering orifice between the load chamber 24 and the outlet chamber 25. Therefore, fluid flow from the supply chamber 22 to the load chamber 23 will take place at a constant pressure differential, automatically maintained by the pump flow control 14, with the control spool 40 remaining in the position as shown and with spool ^'54 in a position all the way to the left. Therefore the flow into the load chamber 23 will be proportional to the area of the metering orifice and therefore to the displacement of the valve spool 18 from its neutral position and independent of the magnitude of the load W.

Assume that while controlling positive load W through the flow control valve 10, a higher load pressure signal is transmitted from the schematically shown flow control valve 13 or 13a through the check valve 106 or 107 and signal line 86 to the pump flow control 14. The discharge pressure of the pump 12 will proportionally increase increasing the pressure differential between the supply chamber 22 and the load chamber 23. The spool 54, of the pilot valve assembly 51, is subjected to the pressure differential between supply chamber 22 and the load chamber 23, since the annular space 64 is connected by line 81 to the supply chamber 22 and the control space 66 is connected by

OMPI lines 69, check valve 71b, the control module 78, line 84, passage 72 and positive load sensing port 37 to the load chamber 23. The increasing pressure differential between the pressure in the supply chamber 22 and the pressure in the load chamber 23 will move the spool 54 from left to right, against the biasing force of the pilot valve spring 68, into a modulating position, as shown, increasing pressure in the control space 49, which will move the control spool 40 from right to left, into a position in which it will throttle fluid flow between the inlet chamber 44 and the supply chamber 22. Therefore, the spool 54, in its modulating position, will automatically throttle, by control spool 40, the fluid flow from the inlet chamber 44 to the supply chamber 22 to maintain the pressure differential between the supply chamber 22 and the load chamber 23, at a constant predetermined level, equivalent to preload in the pilot valve spring 68 and higher than the constant pressure differential of the pump flow control 14. Therefore, irrespective of the pump pressure level, the pilot valve assembly 51 will automatically control the throttling action of the control spool 40, to maintain a constant pressure differential between the supply chamber 22 and the load chamber 23, and across the metering orifice, created by displacement of the metering slot 29. During this control action the free floating piston 55 will be subjected to the pressure differential between the supply chamber 22 and the load chamber 24, which is subjected to minimum pressure and therefore it will be maintained in a position all the way to the left, out of contact with the spool 54.

The leakage device 87 connects control space 66 with the system reservoir. The leakage device 87 may take the form of a simple orifice, or may be of a compensated type, as shown in Fig. 1, permitting a constant flow, at a very low flow level, from control space 66. Such a leakage flow is necessary to permit the spool 54 to move from left to right. Such a movement will close the check valves 71b, 71 and 71a, the displaced fluid from control space 66 being passed by the leakage device 87. The spool 90 of the leakage device 87, in a well known manner throttles, by throttling slots 94, the fluid flow from space 92 to space 93, to maintain space 93 at a constant pressure, equivalent to preload in the spring 96. Since space 93 is maintained at constant pressure and since space 91 is connected to system reservoir, there exists constant pressure differential across leakage orifice 95, corresponding to a constant flow from control space 66 and independent of the pressure level in control space 66.

During control of positive load this constant flow, passing through the leakage device 87, is supplied at positive load pressure from positive load sensing port 37 or 38, through variable throttling orifice 79 of the control module 78, to control space. 66. If the flow area of variable throttling orifice 79 is large, a minimum throttling pressure drop will occur across it and the pressure in control space 66 will approximately equal the positive load pressure. Then, in a manner as explained above, the pilot valve assembly 51 will control the throttling action of the control spool 40, to maintain a constant pressure differential between pressure in the supply chamber 22 and pressure in load chamber 23 or 24, as dictated by the preload of the pilot valve spring 68, irrespective of the magnitude of the discharge pressure of the pump 12. An increase in pressure drop across signal throttling variable orifice 79 will proportionally

OMPI lower the pressure in control space 66, in turn proportionally lowering the pressure in the supply chamber 22 and therefore also decreasing the pressure differential between the pressure in the supply chamber 22 and pressure in the load chamber 23 or 24.

Therefore, any increase in the pressure differential across signal throttling variable orifice 79 will automatically result in a proportional decrease in the pressure level in the supply chamber 22. In this way, by controlling the flow resistance of variable orifice 79, the pressure level in the supply chamber 22 can be controlled.

Assume that the load W, controlled by the fluid motor 11 is of an inertia type and that the valve spool 18 was suddenly moved to the right, creating a larger metering orifice through slot 29. The force necessary to accelerate the load W may result in very high load pressures in the load chamber 23 and even higher pressures in the supply chamber 22. This high load pressure conducted through line 84, will react on the cross-sectional area of the piston 80, generating a force, higher than the preload in the spring 80a, which will result in an upward movement of the piston 80. This movement will reduce the flow area and therefore increase the throttling action of varible orifice 79. In a manner, as previously described, the increased pressure differential, across variable orifice 79, will proportionally lower the pressure in the supply chamber 22, lowering the load pressure in the load chamber 23. In this way, with proper selection of the preload of the spring 80a and the cross-sectional area of the piston 80, the maximum pressure in the supply chamber 22 and therefore the maximum pressure in the load chamber 23 or 24 can be limited to any specific preselected level, in turn establishing the maximum

OMPI pressure, to which the fluid motor 11 can be subjected. This control of maximum pressure, to which the fluid motor 11 can be subjected, in a manner as previously described, is obtained through the throttling action of the control spool 40, which in turn is controlled, in response to the control signal generated by the pilot valve assembly 51. Therefore in a fluid power system employing a plurality of load responsive control valves, the maximum load pressure generated in the fluid motor 11, can be limited to any preselectable maximum value to match the structural integrity of each fluid motor of the system, while working through the existing positive load throttling controls. As previously described the flow through the leakage device 87 is very small. Therefore, throttling of this flow, by the control module 78, takes place at a very _.low energy level, making possible a very accurate and fast responding control. By control of the pressure in the supply chamber 24 the pressure differential between the supply chamber 22 and the load chamber 23 or 24 can be positive or negative. In the. zone of the negative pressure differential the flow between the supply chamber and the load chambers would reverse. Such flow reversal is automatically prevented by the control spool 40 moving to the left and isolating the inlet chamber 44, connected to the pump 12, from the supply chamber 22. This action of the control spool 40 could be compared to the action of a load check valve, well known in the art, and usually located between the load chambers and the supply chamber.

The control of negative load is accomplished through the throttling action of negative load throttling slot 39 or 39a of the control spool 40, the position of which is controlled in response to control signal, generated by the pilot valve assembly 51. When controlling the negative load the pilot spool 54 on one end is subjected to the biasing force of control spring 68 and the pressure in the outlet chamber 25 or 26, while on the other end it is subjected, through the free floating piston 55, to the negative load pressure in negative load sensing port 35 or 36. When controlling a negative load the control will automatically maintain a constant pressure differential between the load chamber 23 or 24 and the outlet chamber 25 or 26. The pressure limiting module 101 is phased into the control circuit and connected down stream of fixed orifice 99 in line 84 and upstream of check valve 85. The control action of the pressure limiting module 101 is as follows. Once the load pressure in line 84 will reach the level, equivalent to the pressure setting of the pressure limiting module 101, the poppet 104 will move against the biasing force of the spring 105 inducing additional flow through the fixed orifice 99 and therefore increasing the pressure drop through the fixed orifice 99. In this way the pressure limiting module 101 will maintain a constant pressure upstream of the check valve 85 as dictated by the preload in the spring 105. The pressure limiting module 101 is phased into the control circuit in such a way, that it will limit, in cooperation with the fixed orifice 99, the load pressure signal transmitted through the check valve 85 and line 86 to the load responsive control 14 of the pump 12, to any preselected maximum level. Therefore, not only the pressure developed in the actuator 11 can be limited to any desired maximum level, by the throttling action of the control spool 40, but the signal transmitted to the pump control, can also be limited to the same or other

OMPI level. Assume that the pressure limiting module 101 is set for a certain selected low pressure level. Then the pressure limiting module 101 will not only limit the maximum pressure developed in fluid motor 11 to this level, but it will also limit the pump discharge pressure, which will be higher, by a constant pressure differential, than the pressure setting of the pressure limiting module 101. Assume that each load responsive valve of the system is provided with a pressure limiting module 101 and control module 78, that each module is set at a different pressure level and that each module not only limits the maximum pressure developed in the respective fluid motor, but also limits, to the same or other level, the pressure signal transmitted through the check valve logic system to the load responsive pump control 14. In this arrangement the maximum discharge pressure of the pump will be limited to the maximum controlled pressure level of any of the loads being simultaneously controlled, resulting in greatly increased system efficiency and longer life of the system pump. This action of the pressure limiting modules 101 is superimposed upon the action of the control modules 78, which can superimpose their control action upon the control action, resulting from the change in position of the valve spool 18, thus providing a unique dual control input while also limiting maximum pressures developed in the fluid motors of the system and limiting the maximum pressure signals transmitted to the load responsive pump control. The pressure limiting module 101 is shown in the form of a relief valve, well known in the art. Such a pressure limiting module must act in combination with fixed orifice 99 and limits the maximum pressure, upstream of the check valve 85, by bypassing some of the flow from the positive load sensing port 37 or 38

OMPI to the system reservoir. It should be noted that the pressure limiting module 101 can be substituted by the control module 78 and control the signal pressure, transmitted to the flow changing pump control 14 by throttling, as long as the pump flow control 14 is provided with the leakage device 87, cross-connecting control space 117, or line 86 with system reservoir. Then the control module 78, by throttling of the load pressure signal and by controlling much lower flows through lines 84 and 86, will limit the maximum load pressure signal transmitted to the pump flow control 14, providing a much faster responding and more precise control than the pressure limiting module 101.

Although the preferred embodiment of this invention has been shown and described in detail it is recognized that the invention is not limited to the precise form and structure shown and various modifications and rearrangements as will occur to those skilled in the art upon full comprehension of this invention may be resorted to without departing from the scope of the invention as defined in the claims.

Claims

Claims

1. A load responsive fluid power and control system comprising a pump (12) provided with a flow changing control (14) responsive to a load pressure signal and operable to maintain a relatively constant pressure differential between its discharge pressure and the pressure of said load pressure signal and a plurality of fluid motors (11) each connected to a load W, a valve assembly (10,51) interposed between said pump (12) and each of said fluid motors (11) having variable orifice means (29,28), fluid throttling means (45,46,40,51) operable to throttle fluid flow to maintain a relatively constant pressure differential across said variable orifice means (28,29), first load pressure signal transmitting means (37,38,72,84,69) interconnecting each of said fluid motors (11) and said fluid throttling means (45,46), second load pressure signal transmitting means (84,86) interconnecting eac of said fluid motors (11) and said flow changing control (14) of said pump (12), and means (101) operable to limit maximum pressure transmitted from each of said fluid motors (11) through said second load pressure signal transmitting means (84,86) to certain predetermined levels without affecting transmission of load pressure signals through said first load pressure signal transmitting means (37,38,72,84,69).

2. A load responsive -fluid power and control system as set forth in Claim 1 wherein check valve means (85) is positioned in said second load pressure signal transmitting means (84,86) down stream of said means (101) operable to limit maximum pressure.

OMPI

3. A load responsive fluid power and control system as set forth in Claim 1 wherein flow restrictor means (99) is positioned in said second load pressure signal transmitting means (84,86) upstream of said means (101) operable to limit maximum pressure.

4. A load responsive fluid power and control system as set forth in Claim 1 wherein said flow changing control (14) of said pump (12) includes pump displacement changing means (110).

5. A load responsive fluid power and control system as set forth in Claim 1 wherein said flow changing control (14) of said pump (12) includes pump discharge flow bypass means (110).

6. A load responsive fluid power and control system as set forth in Claim 1 wherein said variable orifice means (28,29) includes direction control spool means (18).

7. A load responsive fluid power and control system as set forth in Claim 1 wherein said fluid throttling means (45,46,40,51) includes fluid throttling spool means (40).

8. A load responsive fluid power and control system as set forth in Claim 1 wherein said fluid throttling means (45,46,40,51) includes amplifying pilot valve means (51).

9. A load responsive fluid power and control system as set forth in Claim 1 wherein said means (101) operable to limit maximum pressure includes pressure signal throttling means (104,105).

10. A load responsive fluid power and control system as set forth in Claim 1 wherein said means (101) operable to limit maximum pressure includes throttling bypass means (104,105,102,15).

11. A load responsive fluid power and control system as set forth in Claim 1 wherein signal pressure modifying means (78) are positioned in said first load pressure signal transmitting means (37,38,72,84,69).

12. A load responsive fluid power and control system as set forth in Claim 11 wherein said signal pressure modifying means (78) includes signal pressure throttling means (79,80,80a).