WO1998015740A1

WO1998015740A1 - Control valve apparatus for pressure and substantially flow matched electro-hydraulic power steering systems

Info

Publication number: WO1998015740A1
Application number: PCT/US1997/018678
Authority: WO
Inventors: Edward H. Phillips
Original assignee: Techco Corporation
Priority date: 1996-10-10
Filing date: 1997-10-10
Publication date: 1998-04-16

Abstract

A control valve apparatus for determining pressure between a delivery line (18) and a return line (26), and concomitantly, for selectively directing a flow of power fluid at that pressure through a load device (64) from the delivery line (18) to the return line (26). The control valve apparatus comprises sets of first (190), second (194), third (196), and fourth (200) metering ramps sequentially connecting first transverse slots (184) with second transverse slots (186), and then the second transverse slots (186) with third transverse slots (188), for selectively determining pressure between a delivery line (18) and a return line (26) as a function of metering fluid flow therethrough. In addition, the control valve apparatus comprises a similar number of sets of four-way flow control orifice arrays for selectively enabling the flow of power fluid between the delivery line (18) and the return line (26) via the load device (64), which is fluidly coupled across the sets of four-way flow control orifice arrays.

Description

CONTROL VALVE APPARATUS FOR PRESSURE AND SUBSTANTIALLY FLOW MATCHED ELECTRO-HYDRAULIC POWER STEERING SYSTEMS CROSS-REFERENCE TO RELATED APPLICATIONS I. Field of the Invention

The present invention relates generally to electro-hydraulically energized vehicular power steering systems and, more particularly, to control valve apparatus for a pressure and substantially flow matched electro-hydraulic power steering system having extraordinarily high efficiency.

II. Background of the Invention

In order to implement the method and apparatus for controlling the operation of the pressure and substantially flow matched electro-hydraulic power steering system described in the preferred and alternate preferred embodiments of my co-pending U.S. Patent Application entitled "Method and Apparatus for Controlling Pressure and Substantially Flow Matched Electro-Hydraulic Power Steering Systems" (hereinafter "first and second EHPS Systems"), it is necessary to provide control valve apparatus therefor. Generally, the first and second EHPS Systems operate in the following manner:

Selected values of metering fluid flow are applied to a high impedance metering circuit whose flow impedance value is continually under physical control of a host vehicle's driver. Generally this is accomplished in a known manner via the driver's manipulation of a steering wheel driven valve input shaft with reference to instant positions of a valve sleeve which is physically coupled to the vehicle's dirigible wheels. Suitable first and second valve input shaft and valve sleeve sets for this purpose are respectively comprised in control valve portions of control valve apparatus described hereinbelow in preferred and alternate preferred embodiments of the present patent application. In either case, the resulting instant system pressure is substantially determined by the product of the metering fluid flow value and the instant flow impedance value.

In the first EHPS System, all pressurized system fluid flow is provided by a single electrically driven pump unit comprising a servo motor. Selected values of metering fluid flow are derived as small portions of the pressurized system fluid flow. A signal representative of instant values of fluid flow through the high impedance metering circuit is provided by a flow measuring transducer section of a metering flow regulation module such as that described in my co-pending U.S. Patent Application entitled "Apparatus for Regulating Metering Fluid Flow in a Pressure and

Substantially Flow Matched Electro-Hydraulic Power Steering System". The signal is compared with a selected value therefor in order to provide an error signal. The error signal is amplified by a power amplifier and utilized to drive the servo motor of the electrically driven pump unit in such a manner as to continually minimize the error signal regardless of instant values of power fluid flow.

In the second EHPS System, metering fluid flow is provided for the high impedance metering circuit by a separate metering pump in an open- loop fashion. Power fluid flow is provided by an electrically driven pump unit comprising a servo motor. A signal representative of instant differential pressure values therebetween is derived from a null value differential pressure transducer and compared with a selected value in order to provide an error signal. Then the error signal is amplified by a power amplifier and utilized to drive the servo motor of the electrically driven pump unit continually toward volumetric output values which minimize the error signal regardless of instant values of power fluid flow. Further associated with the control valve of either EHPS System is a relatively low impedance flow direction switching power circuit.

Orientation of the switching power circuit is controlled, concomitantly with the high impedance metering circuit, by the driver via the direction of the driver's application of torque to the steering wheel. This enables directionally controlled application of the instant value of system pressure to a power cylinder which is also coupled to the dirigible wheels for achieving the required directionally applied steering force. Finally, because the valve sleeve and power cylinder are linked to the valve input shaft and steering wheel in a known follow-along manner, instant values of driver imposed rotational motion of the steering wheel determine instant values of power fluid flow value and direction through the low impedance flow direction switching power circuit and the power cylinder.

In either EHPS System, the percentage of pressurized fluid utilized by the high impedance metering circuit is quite small. For instance, it would typically be only about 15 percent of the fluid delivered to the power cylinder when parking a host vehicle at a hand wheel rotation value of 400 degrees per second. Thus, hydraulic system efficiencies in excess of 80 percent can be achieved under high pressure parking conditions.

In fact, either EHPS system is competitive with many known pure electric power steering (or EPS) systems in terms of efficiency. This is fortunate since it avoids two serious shortcomings of EPS systems as follows:

1. Safety concerns relating to shutdown modes comprising possible sudden and uncontrollable application of steering force concomitant with a system failure, commonly known in the industry as "auto-steer"; and

2. Undesirable tactile characteristics due to the very high reflected moment of inertia of severely geared electric drive motors, which characteristics are especially noticeable during backup manual operation following a system failure.

Summary of the Invention As mentioned above, the present invention is directed to providing control valve apparatus for pressure and substantially flow matched electro-hydraulic power steering systems. As implied above, preferred and alternate preferred embodiments thereof respectively comprise first and second sets of valve input shaft and valve sleeve members for implementation of first and second high impedance metering circuits and a commonly utilized low impedance flow direction switching power circuit. Physically, the high impedance metering circuits and the low impedance flow direction switching power circuit are each implemented in two identical arrays of slots and/or metering ramps located oppositely from one another for hydrostatically balancing the valve sleeve about the valve input shaft. And in either of the first and second control valves, each set of the two identical arrays of slots and/or metering ramps is located orthogonally from the other in order to fully utilize the circumferentially mating surfaces of the valve input shaft and valve sleeve members.

Each of high impedance metering circuit array of slots and metering ramps comprises first and second spaced axially oriented slots formed within the valve sleeve and four transversely oriented metering ramps formed in the valve input shaft. The first and second transversely oriented metering ramps are formed in an underlapped (i.e., open-center) manner with reference to the first axially oriented slot while the third and fourth transversely oriented metering ramps are formed in an underlapped manner with reference to the second axially oriented slot. In addition, the second and third metering ramps are formed co-linearly in a fluidly coupled manner. In the first control valve, a passage is formed in the valve sleeve for fluidly coupling an external circumferential input groove, which defines an input node, to the first metering ramp, while in the second control valve, a similar passage is used for fluidly coupling an additional external high pressure metering input groove to the first metering ramp. In either case, the metering fluid flow then passes through the first metering ramp to the first axially oriented slot and along that slot to the second metering ramp. It then passes through the second and third metering ramps to the second axially oriented slot and along that slot to the fourth metering ramp. In the first control valve, it then passes through the fourth metering ramp to a passage also formed in the valve sleeve for fluidly coupling it to an additional external metering sensing pressure groove. It is then conveyed to the flow measuring transducer section of metering flow regulation module. In the second control valve, it simply passes through the fourth metering ramp to a return passage formed in the valve input shaft from which it returns to a reservoir via a low pressure valve housing cavity, which defines a return node, and a return port formed in a valve housing.

With either control valve, application of torque to the valve input shaft results in a differential rotational orientation of the valve input shaft with reference to the valve sleeve in a known manner via deflection of a torsion bar. Thus, application of torque in a first direction will result in a reduction of the amount of underlap between the first and second metering ramps and the first axially oriented slot while application of torque in the other direction will result in a reduction of the amount of underlap between the third and fourth metering ramps and the second axially oriented slot. Because of the series orientation of the first, second, third and fourth metering slots, application of torque to the valve input shaft in either direction results in their combined flow impedance value rising and the resulting instant system pressure value also rising as called for above.

Each of the low impedance flow direction switching power circuit arrays of slots comprises slots and passages formed in either of the valve input shaft and valve sleeve sets, which slots and passages generally function in the manner of a closed-center four-way valve. While this could be accomplished in inverted fashion, one suitable array is depicted herein as follows: A passage is formed in the valve sleeve for fluidly coupling the external circumferential input groove to an input slot formed in the valve input shaft. Slightly overlapped input orifices are formed between the input slot and left and right turn slots formed in the valve sleeve. Left and right turn passages are respectively formed in the valve sleeve for fluidly connecting the left and right turn slots and external circumferential left and right turn grooves which respectively define left and right turn nodes. And, the left and right turn nodes fluidly communicate with left and right ports of the power cylinder. Finally, slightly underlapped return orifices are respectively formed between the left and right turn slots and first and second return slots formed in the valve input shaft from which expended power fluid returns to the return port via first and second return passages also formed in the input valve sleeve. It is desirable to form these various slots such that the opening one of the slightly overlapped input orifices achieves its incipient opening point slightly before the closing one of the slightly underlapped return orifices achieves its incipient closing point. Thus, an on-center fluidic fault is present between left and right ports of the power cylinder via the slightly underlapped return orifices. This enables non-pointing on-center behavior of the host steering system. Further, power fluid flow is not possible until the opening one of the slightly overlapped input orifices achieves its incipient opening point. And, although there is a sight bypassing of power fluid during the slight overlap between that opening and the closing of the series connected return orifice, the result is a smooth transition to a positive steering assist condition.

In addition, it is necessary to provide for reverse flow which typically occurs when exiting a turn. Thus, the control valve apparatus comprises left and right turn return check valves for respectively limiting reverse flow losses between the return node and the left or right turn nodes. This also serves to reduce the possibility of bulk cavitation in the left or right turn nodes. Similarly, the control valve apparatus comprises left and right turn input check valves for respectively limiting reverse flow losses between the left or right turn nodes and the input node. The left and right turn input check valves are typically located in the valve housing in common fluidic communication with an input port thereof. However, it has been found that it is preferable to locate the left and right turn return check valves in respective close fluidic communication with the left and right power cylinder ports. In this case they are fluidly connected via volumetrically compliant hose means also positioned in close fluidic communication. Then they are fluidly connected to the return node. Utilizing the closely coupled volumetrically compliant hose means substantially eliminates any tendency for bulk cavitation in the lower pressure side of the power cylinder as well as the left or right turn node itself. A full description of this feature is provided in my co-pending U.S. Patent Application entitled "Grunt Suppression and On-Center Enhancement Device".

Brief Description of the Drawings

The foregoing and other objects and advantages of the present invention will become readily apparent to those skilled in the art upon studying the following detailed description, when considered in connection with the accompanying drawings, in which:

Fig. 1 is a schematic flow circuit illustrative of apparatus for controlling operation of a pressure and substantially flow matched electro- hydraulic power steering system;

Fig. 2 is a schematic flow circuit additionally descriptive of a control valve comprised in a preferred embodiment of the invention;

Fig. 3 is a graphical diagram depicting the relative power requirements of a motor used in a prior art EHPS and a servo motor utilized in conjunction with the preferred embodiment of the invention;

Fig. 4 is a schematic flow circuit illustrative of apparatus for controlling operation of an alternate pressure and substantially flow matched electro-hydraulic power steering system;

Figs. 5A, 5B and 5C are sectional and plan views of control valve elements in accordance with both the preferred and alternate preferred embodiments of the invention;

Fig. 6 is a partially schematic sectional view of a control valve utilized in the preferred embodiment of the invention;

Figs. 7A and 7B are sectional views of fixed and variable orifices utilized in conjunction with the preferred embodiment of the invention;

Fig. 8 is a sectional view of a control valve utilized in the alternate preferred embodiment of the invention. Figs. 9A and 9B are sectional and plan views of an input check valve utilized in conjunction with both the preferred and alternate preferred embodiments of the invention; and

Fig. 10 is a sectional view of return check valves utilized in conjunction with both the preferred and alternate preferred embodiments of the invention.

Detailed Description of the Preferred Embodiments

Depicted in Fig. 1 is a schematic flow circuit illustrative of a first apparatus for controlling a first pressure and substantially flow matched electro-hydraulic power steering system (hereinbelow described as "first

EHPS System 80"). More particularly, Fig. 1 discloses, in block diagram form, a test flow circuit 10 actually used in development of the first apparatus for controlling. In test flow circuit 10, an electrically driven pump unit 12, comprising a pump 14 driven by a servo motor 16, provides a flow of pressurized fluid to a delivery line 18. Selected values of metering fluid flow derived from the flow of pressurized fluid pass through a needle valve

20 to, and through, a flow measuring transducer 22 and on to a return line

26. All returning fluid flow is delivered via return line 26 to a system reservoir 28 from which the pump 14 draws fluid via suction line 30 as required.

The flow measuring transducer 22 comprises either a fixed orifice 32 or a variable orifice 34 through which the flow of metering fluid passes. In practice, the fixed orifice 32 would be utilized in a non-speed sensitive version of the first EHPS System while the variable orifice 34 would be utilized in a speed sensitive version of the first EHPS System. In the case of a speed sensitive version of the first EHPS System, the variable orifice 34 can be manipulated to enable a selected value of metering fluid flow according to a selected function of vehicle speed, engine speed, road conditions and/or any other desired parameter by a control signal issued by the host vehicle's electronic control module 40. In addition, a temperature transducer 36 provides a temperature signal representative of the instant temperature of the metering fluid to a control module 38 along with vehicle speed, engine speed, road conditions and/or any other desired parameter signals.

The flow measuring transducer 22 also comprises a differential pressure transducer 42 which provides a metering fluid flow signal which is actually representative of the differential pressure associated with passage of the flow of metering fluid from a sensing pressure node 44 to the return line 26 via either of the fixed or variable orifices 32 or 34. Utilizing differential pressure measurements across an orifice is well known as a method for providing flow measurement information. For instance, the method is taught in a Chapter entitled "Fluid Measurements" of a well known textbook entitled "Elementary Fluid Mechanics" by John

K. Vennard and published by John Wiley & Sons, Inc. of New York. In any case, the metering fluid flow and temperature signals are respectively transmitted from the flow measuring transducer 22 to the control module 38 via signal transmission lines 46a and 46b. Present versions of the differential pressure and temperature transducers 42 and 36 respectively issue the metering fluid flow and temperature signals as analog signals. Thus, the control module 38 comprises A-to-D converters 48 and 50 which respectively convert those analog signals to multi-bit digital metering fluid flow and temperature signals. Then a digital controller 52 subtractively compares the multi-bit digital metering fluid flow signal with a multi-bit number that is derived from a first look-up table in response to the multi-bit temperature signal to create a multi-bit error signal representative of the difference between instant and selected values therefor. The multi-bit error signal is then processed in a proportional and differential (e.g., "PD") manner to generate a compensated multi-bit error signal. Finally, the compensated multi-bit error signal is applied to a second look-up table from which a linearized multi-bit error signal is delivered to a power amplifier 54 via linearized multi-bit error signal line 56. The power amplifier 54 then issues a suitable amplified power signal to the servo motor 16 via power buss 58. And, the servo motor 16 drives the pump 14 continually toward volumetric output values which minimize the various error signals mentioned above. Actually, the above description of the operation of the digital controller 52 is considerably simplified. For instance, the PD function additionally comprises complex filtering to provide a selected servo "roll- off' characteristic while compensating for such things as mechanical and electrical time constants of the servo motor 16 and volumetric compliance of the delivery line 18 as indicated in Fig. 1 by capacitor 60. And looking forward to an actual application of the methods and apparatus for controlling in an EHPS System installed in a vehicle, further compensation would be required for various hydro-mechanical features thereof, perhaps even including one or more notch filters for eliminating frequencies associated with physical resonances. A full discussion of these problems can be found in my U.S. Patent No. 5,544,715 dated August 13, 1996 and entitled "Method and Apparatus for Enhancing Stability in Servo systems Comprising Hydro-Mechanically Driven Actuators".

In addition, power fluid flow is conveyed between delivery line 18 and return line 26 via power fluid supply line 62 and a load valve 64.

Further, the metering fluid flow regulation module 24 includes dynamic braking section 66 which selectively conveys dynamic braking fluid from delivery line 18 to itself and then on to return line 26 via dynamic braking line 68. The dynamic braking section 66 comprises a spring biased, differentially activated, normally closed two-way valve 70. Its control port 72 is fluidly connected to sensing pressure node 44 while its biasing port 74 is fluidly connected to return line 26. Thus, opening of the spring biased, differentially activated, normally closed two-way valve 70 is also controlled by the sensing pressure. Whenever the force provided by spring 76 is overcome by excessive sensing pressure values, dynamic braking fluid flows through dynamic braking line 68. In normal static operation, selected values for the digital metering fluid flow signal are slightly less than the value that would cause the sensing pressure to overcome the force provided by spring 76.

However, there are two exceptions to this general operational principle. A first exception demonstrates an additional possible operational feature wherein it is possible to speed up the electrically driven pump unit 12 in order to assist in warming up the fluid. This is accomplished at very cold temperatures by setting up the first look-up table such that the multi-bit number derived therefrom in response to very low values of the multi-bit temperature signal results in the sensing pressure partially overcoming the force provided by spring 76. Then the spring biased, differentially activated, normally closed two-way valve 70 is partially opened and additional fluid flow is delivered from the electrically driven pump unit 12.

A second exception can be encountered during rapid closure of the load valve 64 (which is equivalent to hitting the end of travel in either EHPS System described herein or just stopping rotation of the steering wheel for that matter). For reasons of simplicity and cost reduction, the power amplifier 54 is normally configured as a single quadrant amplifier (e.g., one that only puts out unidirectional real power). This means that the electrically driven pump unit 12 is always slowed via back pressure on the pump 14 slowing the pump 14 and servo motor 16 against their combined moment of inertia. While this is happening, normal operation intermittently continues under hydro-mechanical control provided by the spring biased, differentially activated, normally closed two-way valve 70 instead of by the electronic means described above. As will become apparent below with reference to the more detailed system described with reference to Fig. 2, this type of control can also be encountered when a host vehicle using either EHPS System exits a turn. In the case of the first EHPS System 80 described below with reference to Fig. 2, power fluid flows in reverse into delivery line 18 and its energy is dissipated in dynamic braking section 66 under control of the spring biased, differentially activated, normally closed two-way valve 70.

Other than when operating in that particular dynamic braking induced control mode, most operational aspects of the first apparatus for controlling can be demonstrated with the test flow circuit 10. Instant values of system pressure are substantially determined by the product of the metering flow value and the instant flow impedance value encountered by the flow of metering fluid through the needle valve 20. Thus, manipulation of the needle valve 20 results in changes in the system pressure. This is achieved in linearly related fashion to the associated changes in the flow impedance encountered by the flow of metering fluid as it passes through the needle valve 20. In a similar manner, instant values of power fluid flow (i.e., the non- metering fluid portion of instant pressurized fluid flow value) can be selected via manipulation of the load valve 64 and measured by a flow meter 78.

Shown in Fig. 2 is a schematic flow circuit which depicts the function of a first EHPS System 80 comprising most of the elements of the test flow circuit 10. They are similarly numbered and will not be further described herein. In first EHPS System 80, however, the functions of needle valve 20 and load valve 64 are respectively implemented by high impedance metering valve 82 and a four-way low impedance flow switching valve 84 of a control valve assembly 86 comprised in a preferred embodiment of the invention. In first EHPS System 80, the four-way low impedance flow switching valve 84 is used to enable directionally controlled application of instant values of system pressure to a power cylinder 88 for achieving required directionally applied steering force. The four-way low impedance flow switching valve 84 is implemented by two input orifices 90a and 90b, and two return orifices 92a and 92b. Although they are all nominally closed-center orifices, it is desirable for either set of input orifices 90a and 90b or return orifices 92a and 92b to be configured in a slightly underlapped manner and the other of them to be configured in a slightly overlapped manner. Usually, the set of return orifices 92a and 92b is configured in a slightly underlapped manner. This results in an on-center fluidic fault, at the low pressure value of the return line 26, between ports 94a and 94b of the power cylinder 88. This, in turn, enables non-pointing behavior of the first EHPS System 80 when installed in a host vehicle. Providing low pressure for the power cylinder 88 while in its on-center mode is important because it reduces on- center friction via minimized seal drag.

Conversely, the set of input orifices 90a and 90b is usually configured in a slightly overlapped manner in order to preclude any passage of power fluid from the power fluid supply line 62 until a preselected level of torque is applied to the host vehicle's steering wheel. This, in combination with the slightly underlapped set of return orifices 92a and 92b, enables an ideal on-center feel of pure mechanical steering without any flow of power fluid from the power fluid supply line 62. Further, it is desirable to form the four-way low impedance flow switching valve 84 such that the incipient opening point of the slightly overlapped set of input orifices 90a and 90b occurs at a slightly higher value of applied torque than the incipient closing point of the slightly underlapped set of return orifices 92a and 92b. This ensures a smooth transition from purely mechanical on-center steering to power assisted steering after the closing of the slightly underlapped set of return orifices 92a and 92b.

In addition to enabling directionally controlled application of instant values of system pressure to the power cylinder 88, the control valve assembly 86 must be able to effectively extract regenerative hydraulic power from the power cylinder 88 whenever the host vehicle exits a turn. This is because the dirigible wheels then drive the power cylinder 88 in a reverse direction. This would be problematic for a first EHPS System 80 comprising the so far described apparatus only, because such dirigible wheel powered motion of the power cylinder 88 can be independent of the instant disposition of the four-way low impedance flow switching valve 84. And in general, resulting instant values of flow impedance of the input and return orifices 90a, 90b, 92a and 92b might yield excessive (and non- linear) pressure value changes.

Thus, in addition to the four-way low impedance flow switching valve 84, the control valve assembly 86 comprises input check valves 96a and 96b respectively positioned in parallel with input orifices 90a and 90b, and return check valves 98a and 98b respectively positioned in parallel with return orifices 92a and 92b. Whenever the power cylinder 88 moves with significant velocity in the reverse direction, either set of return and input check valves 98a and 96b, or 98b and 96a, open and convey a regenerative flow of power fluid to the power fluid supply line 62. And whenever either of the input check valves 96a and 96b returns a flow of regenerative power fluid to the delivery line 18 via the power fluid supply line 62, that regenerative power is handled in a dynamic braking manner by the dynamic braking section 66 of the metering fluid flow regulation module 24 as is fully described above.

Shown in Fig. 3 is a graphical diagram depicting the relative speed/torque requirements of a drive motor comprised in a prior art EHPS system and the servo motor 16 of a similarly sized first EHPS System 80 during identical parking maneuvers. In this example both motors are of the field weakening type and could, for instance, be either switched reluctance or induction motors. Identical maximum motor performance envelopes are depicted in Fig. 3 by line 100, which depicts maximum motor torque, hyperbolic curve 102, which depicts maximum motor power, and line 104, which depicts maximum motor speed. Generally, the motor of the prior art EHPS system operates somewhere along the line 104 or hyperbolic curve 102. On the other hand, instant torque/speed operating points of the servo motor 16 are respectively determined by system pressure and steering wheel speed as is described above.

As depicted by point 106, a typical parking maneuver that starts on-center might begin at an operating point of full speed and about 40% torque for the drive motor of a prior art EHPS system. As tie rod and knuckle arm geometry become less favorable, a locus of instant operating points follows line 104 and hyperbolic curve 102 to maximum steering deflection operating point 108. Then, assuming that the driver forcibly holds the steering system at its end point or stop, the locus of instant operating points follows hyperbolic curve 102 to maximum power and torque operating point 110 where it remains as long as the driver continues to apply torque. In the first EHPS System 80 on the other hand, a parking maneuver at 400 degrees per second steering wheel velocity that similarly starts on-center begins at an operating point of about 45% of full speed and about 40% torque for the servo motor 16 as depicted by point 112. As the tie rod and knuckle arm geometry become less favorable, a locus of instant operating points follows line 114 to maximum steering deflection operating point 116. Then, assuming that the driver forcibly holds the steering system at its end point or stop, the locus of instant operating points follows curve 118 to maximum torque and metering flow operating point 120 where it remains as long as the driver continues to apply torque.

A comparison of the relative motor speeds required by the two systems in performing the same parking maneuver readily confirms the efficiency advantages of first EHPS System 80 over the prior art EHPS system. Instant power requirements of the servo motor 16 relative to those of the drive motor of a prior art EHPS system are linearly related to the horizontal fraction of the distance to line 104 or hyperbolic curve 102 depicted by positions along line 114 or curve 118. And, this is for the industry standard for maximum steering wheel velocity during parking. A more normally seen steering wheel velocity of 200 degrees per second would proportionately reduce these instant power levels to those depicted by line 114a and curve 118a. Overall, such a typical parking maneuver could be accomplished by the first EHPS System 80 with about 1/3 of the energy expended by the prior art EHPS system.

Shown in Fig. 4 is a schematic flow circuit which depicts the function of a second EHPS System 130 comprising many of the elements of the test flow circuit 10 and first EHPS System 80. They are similarly numbered and will not be further described herein. In second EHPS System 130, however, a separate metering fluid supply module 132 supplies a selected value of metering fluid flow directly to the high impedance metering valve 82. A temperature transducer 36 provides a temperature signal representative of the instant temperature of the metering fluid to a metering control module 134. The metering control module 134 also receives other desired parameters from the host vehicle's control module 40 (as described above with reference to control module 38) and selectively provides electric power to a metering motor/pump 136 via power buss 138. This results in the provision of the selected value of metering fluid flow to the high impedance metering valve 82 via metering fluid line 140 according to a selected function of fluid temperature, vehicle speed, engine speed, road conditions and/or any other desired parameter. The spent flow of metering fluid then returns to a system reservoir 142 via metering flow return line 144. Fluid is drawn via suction line 146 (from the system reservoir 142) by the metering motor/pump 136 as required. Instant values of metering pressure present in the metering fluid line 140 are applied to a biasing port 148 of a null value differential pressure transducer 150. Concomitantly, electrically driven pump unit 12 provides a flow of power fluid to power fluid line 152 wherefrom instant values of control pressure are applied to control port 154 of the null value differential pressure transducer 150. A signal representative of the differential pressure therebetween is delivered to a control module 156, via differential pressure signal line 158, wherein it is digitized and filtered as described above with reference to control module 38. A linearized multi- bit error signal is then conveyed from the control module 156 to a power amplifier 162 via linearized multi-bit error signal line 160. The power amplifier 162 then issues a suitable amplified power signal to the servo motor 16 via power buss 164. And, servo motor 16 drives the pump 14 continually toward volumetric output values which minimize the linearized multi-bit error signal.

Within null value differential pressure transducer 150, biasing pressure is conveyed to the spring loaded end of a spring biased, differentially activated, normally closed two-way valve 166, and control pressure is conveyed to its opposite end. Regenerative power can be dissipated in a dynamic braking manner by the null value differential pressure transducer 150 via dynamic braking line 165 similarly to the dynamic braking action described above with respect to dynamic braking section 66 of the metering fluid flow regulation module 24. Concomitantly, a position measuring transducer 168 provides a position signal indicative of the actual position of a moving element 170 of the spring biased, differentially activated, normally closed two-way valve 166 to the control module 156. In a manner similar to that described above, the control module 156 utilizes digitizing means, lookup tables and the like to provide a linearized multi-bit error signal conveyed to the power amplifier 162.

Finally, the flow of power fluid delivered to power fluid line 152 from electrically driven pump unit 12 is also conveyed to a control valve assembly 86 and thereby to a power cylinder 88. And, the second EHPS System 130 performs in a substantially identical manner to the first EHPS

System 80. Particularly, the operational power requirements of second

EHPS System 130 are substantially identical to those illustrated in Fig. 3B.

Depicted in Figs. 5A, 5B and 5C are common elements of two slightly differing versions of control valves embodied in the preferred embodiment of the invention. Comprised are two identical high impedance open-center circuits 180. As shown in Fig. 5A, pressurized fluid enters the high impedance open-center circuits 180 via input ports 182. Each high impedance open-center circuit 180 comprises a series arrangement of three transverse slots including input slot 184, transverse connecting slot 186 and output slot 188 as particularly shown in Fig. 5C wherein an input port 182 is indicated in phantom by dashed lines. In operation, metering fluid enters each high impedance open-center circuit 180 via input port 182 and flows over a first metering ramp 190 to a first axial connecting slot 192 as particularly shown in Fig. 5B. The metering fluid then flows along first axial connecting slot 192 and enters transverse connecting slot 186 by flowing over a second metering ramp 194. The metering fluid then flows over a third metering ramp 196 to a second axial connecting slot 198. Finally, the metering fluid flows along second axial connecting slot 198, enters output slot 188 by flowing over a fourth metering ramp 200, and exits via sensing pressure port 202.

Relative rotation of input shaft 204 with respect to valve sleeve 206 is attained in a well known manner by applying torque to input shaft 204 and twisting interconnecting torsion bar 208. Relative clockwise rotation of input shaft 204 with respect to valve sleeve 206 results in progressive closure of pressure control orifice 210a formed between first metering ramp 190 and corner 212 of first axial connecting slot 192, and a similar pressure control orifice 210b associated with second metering ramp 194. Conversely, relative counterclockwise rotation of input shaft 204 with respect to valve sleeve 206 results in progressive closure of similar pressure control orifices 210c and 21 Od associated with third and fourth metering ramps 196 and 200. In either case, there is progressive closure of two series connected pressure control orifices whereby the total pressure drop associated with the high impedance open-center circuits

180 is substantially equal to twice the pressure drop associated with each closing pressure control orifice. Depicted in Fig. 6 are additional details of a control valve 220 in accordance with the preferred embodiment of the invention. Fig. 6 is a sectional view taken along section lines respectively labeled A-A and B-B in Fig. 5A. Additional general details of control valve 220 include a positive attachment of torsion bar 208 to a pinion 222 with a pin 224 and a driving attachment of the pinion 222 to a valve sleeve 206a via pin 226. In addition, the first and second axial connecting slots 192 and 198 as well as respective left and right turn slots 228a and 228b are formed as axially oriented slots in the valve sleeve 206a and sealed off by rings 230 on each end. An input shaft 204 is supported for relative rotation with respect to the pinion 222 by bushing or needle bearing 232. Finally, the input shaft 204 and the pinion 222 are respectively formed with three shaft extension notches 234 and three arms 236 as described in my U.S. Patent Application Serial Number 08/577,415 filed on December 22, 1995 and entitled "Method and Apparatus for Forming a Control Valve for Hydraulic

Circuits".

In operation, metering and power fluid commonly enter power fluid annular groove 238 via power fluid supply line 62 and an input housing port (not shown). Then the metering fluid passes through input port 182 to input slot 184. Then the metering fluid passes on through transverse connecting slot 186 and output slot 188 to a sensing pressure port 202 as described above. Sensing pressure is conveyed to the sensing pressure node 44 (shown in Figs. 1 and 2) via sensing pressure annular groove 240 and a sensing housing port (not shown). Concomitantly (and with continued reference to Figs. 5A and 6 as well as Figs. 1 and 2), power fluid passes through an input port 242 to input slot 244 formed in input shaft 204. Application of torque to input shaft 204 enables power fluid to flow through either of the input power orifices 90a or 90b that may be open via respective superposition of input slot 244 and left or right turn slots 228a or 228b. Assuming that the input shaft 204 is rotating toward the direction of a turn, the power fluid flows from the open one of the input power orifices 90a or 90b through respective ones of left or right turn slots 228a or 228b, left or right turn ports 246a or 246b, left or right turn circumferential grooves 247a or 247b, and left or right turn passages (i.e., passages generally provided in a rack- and-pinion power steering gear by left and right turn tubes) 248a or 248b to left or right ports 94a or 94b of power cylinder 88. Then returning power fluid flows from the other of left or right ports 94a or 94b through the other of left or right turn passages 248a or 248b and left or right turn ports 246a or 246b to the other of left or right turn slots 228a or 228b. Power fluid flows through the concomitantly open opposite one of return power orifices 92a or 92b, formed via superposition of the appropriate ones of left or right turn slots 228a or 228b and return slots 250, and then through return ports

252 to annular passage 254 formed between input shaft 204 and the torsion bar 208. Finally, the spent power fluid exits through return ports 256 formed in the three shaft extension notches 234 and flows to the reservoir 28 via a return housing port (not shown). Depicted in Figs. 7A and 7B are respective sectional views of the flow restrictive portions of fixed and variable orifices 32 and 34 alternately utilized in first EHPS System 80. The fixed orifice 32 comprises a nozzle shape characterized by the flow of metering fluid serially passing through a radiused entry section 258 and a divergent passage 260 as indicated by flow direction arrow 262. On the other hand, variable orifice 34 is configured as a solenoid (not shown) operated needle valve 264 similarly characterized by the flow of metering fluid serially passing through a radiused entry section 266 and a divergent passage 268 as indicated by flow direction arrow 270. In this case, divergent passage 268 is formed between solenoid driven needle 272 and conical bore 274 of housing 276. Divergent geometries are utilized to avoid silting problems while divergence angles and lengths of divergent passages 260 and 268 are experimentally developed to substantially match low temperature flow characteristics to those of pressure control orifices 210a-d.

Depicted in Fig. 8 are alternate control valve details of a slightly modified control valve 280 in accordance with an alternate preferred embodiment of the invention. Control valve 280 is utilized in a second EHPS System 130. Fig. 8 is a sectional view taken along section lines respectively labeled A-A and B-B in Fig. 5A. Most details of control valve 280 are identical to those of control valve 220. They are similarly numbered and will not be further described herein. In this case, however, metering fluid is separately provided by the metering fluid supply module 132 to metering fluid annular groove 282. The metering fluid enters input slot 184 via input port 284 and exits output slot 188 by flowing through return port 286 formed in an input shaft 204a to annular passage 254 and on to reservoir 28 as described above.

Shown in Figs. 9A and 9B are sectional and plan views of an input check valve assembly 290 utilized in conjunction with either of left or right turn circumferential grooves 247a or 247b of control valves 220 and 280. Input check valve assembly 290 comprises a check valve seat 292, a ball 294 and a check valve stop 296. First check valve seat 292 is mounted and retained in a first counter bore 298 of a control valve housing 300 by an interference fit. Then the ball 294 is positioned upon the check valve seat 292 as shown in Fig. 9A, and finally, check valve stop 296 is mounted and retained by an interference fit in a second, and larger, counter bore 302 in common alignment with surface 304 of the control valve housing 300.

In operation, regenerative fluid flow enters input check valve assembly 290 from either of left or right turn circumferential grooves 247a or 247b via passage 306 and lifts the ball 294 off of check valve seat 292 as indicated by the dotted outline 294a. The ball 294 is retained in a nominally central position above check valve seat 292 via containment by cavity 308 formed in check valve stop 296. The regenerative fluid flow then passes through longitudinal slots 310 formed in the outer periphery of check valve stop 296 which are more clearly depicted in Fig. 9B. The regenerative fluid flow then flows into an entry cavity formed by pocket 312 in an input housing 314 (which, in details not shown, contains dynamic braking section 66 and some components of flow measuring transducer 22). Fig. 10 is a plan view depicting placement of left and right return check valves 98a and 98b respectively juxtaposed to left and right ports 94a and 94b of the power cylinder 88. Actually, the left and right return check valves 98a and 98b are comprised within a connective assembly 318 which even includes left and right turn tubes 248a and 248b. Physically, the left and right return check valves 98a and 98b are comprised within banjo fittings 320 wherein their respective flow directions are indicated by arrows 322a and 322b. In addition, they are fluidly connected to the return line 26 via volumetrically compliant hose members 324a and 324b, each retained by crimp fittings 326 in a known manner. In greater detail, the return line comprises input tube 26a, cross-over tube

26b and output tube 26c as well as "tee" fittings 26d.

Utilizing volumetrically compliant hose members 324a and 324b in such close proximity to left and right ports 94a and 94b substantially eliminates any tendency for road shock induced bulk cavitation in a lower pressure side of the power cylinder. This is because such a lower pressure side of power cylinder 88 is then capacitively coupled to "ground" via the compliant wall of a volumetrically compliant hose whenever its respective check valve is open. The compliant wall of the volumetrically compliant hose acts as a low pass filter in effectively eliminating low pressure "spikes". A full description of this feature is provided in my co- pending U.S. Patent Application entitled "Grunt Suppression and On- Center Enhancement Device". While the present invention has been described in connection with the preferred embodiments of the various figures, it is also understood that other similar embodiments may be used, or modifications or additions may be made to the described embodiments for performing the same function of the present invention without deviating therefrom. Therefore, the present invention should not be limited to any single embodiment but, rather, construed in breadth and scope in accordance with the recitation of the appended claims. I claim:

Claims

1. Control valve apparatus for selectively determining pressure between a delivery line and a return line as a function of metering fluid flow therethrough comprising: a first valve member comprising at least one transverse slot; a second valve member comprising at least one set of first and second axial slots; at least one first metering ramp for fluidly connecting said at least one transverse slot to said at least one first axial slot in a selective manner such that its flow impedance increases with relative motion of said first valve member with respect to said second valve member in a first direction; at least one second metering ramp for fluidly connecting said at least one transverse slot to said at least one second axial slot in a selective manner such that its flow impedance increases with relative motion of said first valve member with respect to said second valve member in the other direction; and means for fluidly connecting said first and second metering ramps such that pressurized fluid flows therethrough in a series connected manner.

2. A control valve apparatus for selectively determining pressure between a delivery line and a return line as a function of metering fluid flow therethrough comprising: a first valve member comprising at least one set of first second and third transverse slots; a second valve member comprising an identical number of sets of first and second axial connecting slots; an identical number of first metering ramps for fluidly connecting each of said first transverse slots to a respective one of said first axial connecting slots in a selective manner such that its flow impedance increases with relative motion of said first valve member with respect to said second valve member in a first direction; an identical number of second metering ramps for fluidly connecting each of said first axial connecting slots to a respective one of said second axial connecting slots in a selective manner such that its flow impedance increases with relative motion of said first valve member with respect to said second valve member in the other direction; a identical number of fourth metering ramps for fluidly connecting each of said second axial connecting slots to a respective one of said fourth transverse slots in a selective manner such that its flow impedance increases with relative motion of said first valve member with respect to said second valve member in the other direction; and means for fluidly connecting said first, second, third and fourth metering ramps such that pressurized fluid flows therethrough in a series connected manner.

3. The control valve apparatus of claim 1 additionally comprising an identical number of sets of four-way flow control orifice arrays, each set comprising first and second input flow control orifices, and first and second return flow control orifices for selectively applying a flow of power fluid between said delivery line and said return line via a load device fluidly coupled across said sets of four-way flow control orifice arrays.

4. The control valve apparatus of claim 3 additionally comprising first and second input reverse flow check valves respectively fluidly connected in parallel with said first and second input flow control orifices, and first and second return reverse flow check valves respectively fluidly connected in parallel with said first and second return flow control orifices for selectively returning a flow of regenerative power fluid from said return line to said delivery line whenever an amount of said flow of regenerative power fluid in excess of that transferable by said sets of four- way flow control orifice arrays is produced by said load device.