Patent Application
20060081410
The present invention relates generally to power steering systems for vehicles, and more particularly to an energy efficient power steering system intended particularly for medium to large vehicles.
Virtually all present power steering systems comprise implementation means whose fundamental output is force based. By way of example, present art power steering systems generally comprise an open-center four-way valve that delivers differential pressure to a double-acting power cylinder as a function of torque applied to a steering wheel. This is accomplished via torque applied to the steering wheel progressively closing off return orifices comprised within the open-center four-way valve. Another example is an electric power steering system (hereinafter ‘“EPS system”) wherein a servomotor delivers torque to the steering gear as a function of current applied to it by a controller. An EPS system of particular interest herein is described in U.S. Pat. No. 6,152,254, entitled “Feedback and Servo Control for Electric Power Steering System with Hydraulic Transmission,” issued Nov. 28, 2000 to Edward H. Phillips, wherein differential pressure is directly delivered to a double-acting power cylinder from a servomotor driven reversible fluid pump. In view of continued reference hereinbelow to the '254 patent, the whole of that patent is also expressly incorporated in its entirety by reference herein.
While the EPS system described in the incorporated '254 patent has optimum performance characteristics, it like all EPS systems is limited in utilization to relatively small vehicles because of limited available electrical power. All vehicle manufacturers limit electrical current availability for EPS systems to a value that can be supplied directly from an alternator. A limiting value of perhaps 70 Amperes from a 12 Volt electrical system is typical. At a lower limiting voltage value of 10 Volts and an overall EPS system efficiency of perhaps 60 percent this results in a net maximum power delivery from the steering gear of only 420 Watts. This low value stands in stark contrast to known future power steering system requirements ranging as high as 3,500 Watts.
Various so-called “closed-center” power steering systems have been proposed as a solution to this problem. Such closed-center power steering systems utilize an accumulator to store power steering fluid at relatively high pressure. Some form of closed-center valving is then used to meter a flow of pressurized fluid to one end of a double-acting power cylinder while concomitantly permitting a similar return flow of low-pressure fluid from the other end thereof to a reservoir. Generally, pressurized fluid is supplied to the accumulator from the reservoir by a relatively small displacement pump driven by a simple (e.g., non-servo) motor controlled by a pressure-activated switch.
To date however, none of the proposed closed-center power steering systems has provided acceptable on-center steering “feel” and they have not gained acceptance in the industry. It is believed herein that the primary problem with the closed-center power steering systems proposed to date is that their fundamental output is fluid flow or rate-based rather than force-based as is described above with reference to currently accepted power steering systems. The fundamental problem with the rate-based closed-center systems is that they provide nominally linear control of system velocity with inherent discontinuities in system acceleration. It is believed herein that these discontinuities in system acceleration are the root cause of the unacceptable on-center steering feel in the closed-center power steering systems. By way of contrast, all force-based systems provide direct quasi-linear control of system acceleration.
Therefore, it would be highly advantageous to provide an accumulator enabled power steering system that has the acceptable on-center steering “feel” provided by a force-based power steering system. Such a force-based power steering system was disclosed in U.S. Pat. No. 6,945,352 entitled “FORCE-BASED POWER STEERING SYSTEM,” issued Sep. 20, 2005 to Edward H. Phillips, which is hereby incorporated by reference in its entirety. Since the application for that patent was filed however, some have suggested that they would prefer a system with improved failsafe characteristics wherein unwanted steering forces are not possible regardless of failure mode.
An accumulator enabled power steering system according to the present invention functions as a force-based power steering system in an inherently failsafe manner.
The accumulator enabled power steering system of the present invention includes a directional control open-center four-way valve having an input port, a return port fluidly connected to a reservoir, and left and right output ports respectively fluidly connected to left and right cylinder ports of a power cylinder. An electronically controlled slightly over-lapped normally open three-way valve has an input port fluidly connected to an accumulator and a return port fluidly connected to the reservoir. An output port of the three-way valve is fluidly connected to the input port of the four-way valve. A valve spool in three-way valve is spring-loaded in accordance with the three-way valve's designation of being “normally open” such that the output port and therefore the input port of the four-way valve are normally fluidly connected to its return port and therefore the reservoir.
A steering wheel torque transducer provides an applied torque signal Vat indicative of values of torque applied to the steering wheel (hereinafter “applied torque”). A pressure transducer provides a pressure signal Vp indicative of pressure values present at the input port of the directional control open-center four-way valve. A controller provides a power control signal Vc to the three-way valve at values determined via filtering and amplifying an error signal Ve. The error signal Ve is generated by the difference between a control function signal Vcf determined by a control algorithm from at least the applied torque signal Vat and the pressure signal Vp issued by the pressure transducer. The power control signal Vc is for controlling the three-way valve such that pressurized fluid is supplied to the input port of the four-way valve at fluid pressure values that continually reduce the error signal Ve. Thus, pressurized fluid is provided by the four-way valve to one of the ports of the double-acting power cylinder as determined by the rotational direction of the applied torque at a value in accordance with the magnitude of the applied torque and the resulting control algorithm determined control function signal Vcf.
The accumulator is initially and then intermittently charged with fluid such that the accumulator fluid pressure is always greater than a selected threshold value exceeding that required for executing any likely steering load. Operationally, whenever torque is applied to the steering wheel, an applied torque signal Vat is sent to the controller by the torque transducer. First, the absolute value of the applied torque signal Vat is multiplied by a control function constant Kcf to form the control function signal Vcf, wherein the control function constant Kcf is determined by the above mentioned control algorithm as a selected function of the applied torque value, and in addition, most likely at least the vehicular speed in accordance with procedures fully explained in the incorporated '254 patent. The pressure signal Vp from the pressure transducer is then subtracted from the control function signal Vcf whereby the resulting algebraic sum forms the error signal Ve. The error signal Ve is then filtered and amplified to form the power control signal Vc that is then used to control the three-way valve such that appropriately pressurized fluid is provided to the appropriate power cylinder port as directed by the directional control open-center four-way valve in accordance with the rotational direction of the applied torque. Thus, steering force is applied to the dirigible (steerable) wheels of the host vehicle in accordance with the rotational direction and magnitude of the applied torque. Such three-way slightly over-lapped servo valves and their operative characteristics are thoroughly described in a book by Herbert E. Merritt entitled “Hydraulic Control Systems” and published by John Wiley & Sons, Inc. of New York.
It is desirable for working pressures in the double-acting power cylinder to always be kept at the lowest pressure values possible. This keeps pressure values applied to various power cylinder seals to a minimum thereby reducing leakage problems and minimizing Coulomb friction. The directional control open-center four-way valve, wherein at least one of the left and right output ports is always fluidly connected to return port and thus the reservoir, automatically accomplishes this task of course. In addition however, it is also desirable to fluidly couple both of the left and right cylinder ports to the reservoir during “on-center” steering conditions. This improves overall system efficiency by allowing small on-center steering motions to be effected without using any accumulator-sourced fluid. In the accumulator enabled power steering system of the present invention this is automatically accomplished by configuring the control algorithm such that the control function constant Kcf has zero values for small near on-center values of applied torque. This in turn results in the normally open slightly over-lapped three-way servo valve having zero valued power control signals for small near on-center values of torque applied to the steering wheel whereby both cylinder ports are fluidly connected to the reservoir.
A primary failsafe shutdown procedure is implemented via precluding current from being applied to the three-way valve whereby the spring-loaded valve spool again causes its output port and therefore the input port of the directional control open-center four-way valve to be fluidly connected to the reservoir thus imposing manual steering regardless of steering load. Furthermore, a redundant failsafe feature is provided via the four-way valve directly controlling fluid flow to the ports of the power cylinder in the manner of the present power steering systems mentioned above.
Overall system accuracy and stability is provided during normal operation via a feedback control loop implemented with reference to the pressure signal Vp representative of actual fluid pressure values present at the input port of the directional control open-center four-way valve. Because this type of control technique is described in detail in the incorporated '254 patent, it will not be repeated in full detail herein.
Because of its improved steering feel and ability to service known future power steering systems whose net hydraulic power requirements range as high as 3,500 Watts, a power steering system configured according to the present invention possesses distinct advantages over known prior art power steering systems able to handle such large steering loads. For example, the power steering system of the present invention provides dramatically improved system efficiency when compared to standard hydraulic power steering systems utilizing engine driven pumps. Further, the power steering system of the present invention provides dramatically improved tactile feel when compared to known prior art accumulator and closed-center valve enabled power steering systems. Thus, the accumulator enabled power steering system of the present invention enables both efficient and tactilely acceptable power steering for medium to large vehicles.
Other advantages of the present invention can be understood by reference to the following detailed description when considered in connection with the accompanying drawings wherein:
The present invention is directed to simplified method and apparatus for enabling an accumulator enabled power steering system to function in the manner of a force-based power steering system. With reference first to
The accumulator 26 is initially and then intermittently charged with pressurized fluid such that the accumulator fluid pressure is greater than a selected threshold value exceeding that required for meeting any likely steering load. Operationally, whenever torque is applied to the steering wheel 12, an applied torque signal Vat is sent to the controller 30 by a torque transducer 32 operatively connected thereto. Then as will be further described below, the absolute value of the applied torque signal Vat is multiplied by a control function constant Kcf to form a control function signal Vcf, where the control function constant Kcf is generated by the controller 30 as a function of at least the applied torque value, and most probably vehicular speed, in accordance with procedures fully explained in the incorporated '254 patent. A pressure signal Vp from a pressure transducer 34 provided for measuring pressure values in the fluid line 18 is then subtracted from the control function signal Vcf whereby the resulting algebraic sum forms an error signal Ve. The error signal Ve is then filtered and amplified to form a power control signal Vc that is then continuously applied to the three-way valve 28 in such a manner as to cause the error signal Ve to decrease in value. As will be further described hereinbelow, it is desirable for the control function constant Kcf generated by the controller 30 to have a zero value to relatively low initiating values of applied torque (i.e., +/−7.5 in.lbs.) and then blend into a selected linear control characteristic over perhaps twice that range in order to effect a preferred on-center steering characteristic.
With particular reference now to
With particular reference now to
As a design choice, either one of the valve sleeve 48 and input shaft 50 comprises multiple input slots 56 and return slots 58 while the other one of the valve sleeve 48 and input shaft 50 comprises multiple left output slots 60a and right output slots 60b (i.e., as depicted in
The directional control open-center four-way valve 20 is formed in an open-center manner as a consequence of the input slots 56 and return slots 58, and left output slots 60a and right output slots 60b all being formed with greater widths than juxtaposed lands 68 whereby input orifices 70a and 70b, and return orifices 72a and 72b are all enabled for freely conveying fluid in the on-center position as illustrated in
Optimum performance of the three-way valve 28 can be obtained by optimizing its flow gain. As depicted in
As further explained in detail in the book entitled “Hydraulic Control Systems,” flow values in either of the delivery flow or return flow directions can be determine by
Q=70w×Sqrt[deltaP]
where Q is flow rate, w is circumference and x instant valve stroke of the spring-loaded valve spool 38, and deltaP is pressure drop across the valve orifice. In addition, stroking force can be found by
F=0.006/Sqrt[deltaP]Q+kx+F0
where F is the stroking force, and k and F0 are the spring constant and force associated with the valve null position of the spring-loaded valve spool 38. Finally, the valve flow gains in either direction can be defined as the ratio of flow to variable portions of the stroking force or
Kq=Q/(F−F0)=1/(0.0061Sqrt[deltaP]+k/(70wSqrt[deltaP]))
where Kq is valve flow gain. Thus, the flow-sourced portion becomes dominant at high values of pressure drop and the spring rate-sourced portion becomes dominant at low values of pressure drop. This results in minimum valve flow rate gain values occurring at the extremes and larger values perhaps 2 to 3 times larger occurring at moderate pressure drop values in between.
The transition between dissimilar flow delivery and flow return curve slopes is eased however, by virtue of the three-way valve 28 being configured in a slightly over-lapped manner. As depicted in the book entitled “Hydraulic Control Systems,” this would result in a zero slope, and thus zero valve gain, between so bifurcated critical positions of an “ideal” such slightly over-lapped three-way servo valve. This is not the case with a practical slightly over-lapped three-way valve 28 however, because of its finite leakage characteristics. Thus, there is a smooth transition of valve gain through the bifurcated critical position region in the manner depicted in
In most cases adequate control can be achieved without tailoring feedback filtering in accordance with instant deltaP values, or alternately, by limited such tailoring achieved through interpretation of which one of the input grooves 40 or return grooves 44 is instantly in use via a combination of signals indicative of solenoid current and output pressure value. However, such tailoring may in some cases be desirable. In such cases, it is necessary to additionally provide the controller 30 with a signal indicative of the direction of fluid flow through the three-way valve 28 in order for it to interpret which of the input groove 40 or return groove 44 is instantly being utilized. This of course requires additional means for determining the direction of fluid flow. Perhaps the easiest way to determine the direction of fluid flow is to take advantage of the obvious correlation between fluid flow direction and steering wheel motion by utilizing a steering wheel motion direction sensor 74 to determine the direction of rotational motion of the steering wheel and then convey a signal so indicative to the controller 30. As shown in
It is desirable for working pressures in the double-acting power cylinder 16 to always be kept at the lowest pressure values possible. This keeps pressure values applied to various power cylinder seals to a minimum thereby reducing leakage problems and minimizing Coulomb friction. The directional control open-center four-way valve 20, wherein at least one set of the left output slots 60a and right output slots 60b is always fluidly connected to the return slots 58 and thus the reservoir 24, automatically accomplishes this task of course.
In addition, it is also desirable to fluidly couple both of the left output slots 60a and right output slots 60b (and thus the left cylinder port 14a and the right cylinder port 14b) to the reservoir 24 during “on-center” steering conditions. This improves overall system efficiency by allowing small on-center steering motions to be effected without using any accumulator-sourced fluid. In the accumulator enabled power steering system 10 this is automatically accomplished by configuring the control algorithm such that the control function constant Kcf has zero values for near on-center values of applied torque (i.e., such as +/−7.5 in.lbs.). This in turn results in the normally open slightly over-lapped three-way valve 28 having zero valued power control signals for small near on-center values of torque applied to the steering wheel whereby both the left cylinder port 14a and the right cylinder port 14b are fluidly connected to the reservoir 24.
In the accumulator enabled power steering system 10 a primary failsafe shutdown procedure is implemented via precluding current from being applied to the three-way valve 28 whereby the spring-loaded valve spool 38 again causes its output groove 42 and therefore the fluid line 18 and the input slots 56 of the directional control open-center four-way valve 20 to be fluidly connected to the reservoir 24 thus imposing manual steering regardless of steering load. Furthermore, a redundant failsafe feature is provided via the directional control open-center four-way valve 20 directly controlling fluid flow to the left cylinder port 14a and the right cylinder port 14b of the double-acting power cylinder 16 in the manner of present power steering systems as mentioned above.
A fluid source must of course be provided for charging the accumulator 26 with pressurized fluid. An electrically driven fluid source can be utilized for this purpose as is indicated in alternate forms in
On the other hand, it may be desired to maintain the supply pressure in the accumulator 26 at a nominally constant value in order to maintain the consistent gain characteristics for the three-way valve 28. In this case, the drive motor 224 is configured as a variable speed drive motor driven by a controlled power signal issuing from the controller 30 such that the drive motor 224 and pump 228 function as part of a relatively simple servo system for maintaining the supply pressure at a preselected nominal value.
On the other hand, an accessory drive train 236 of the engine 238 of the host vehicle can be directly utilized to mechanically drive the pump 228 in either of the manners depicted in
With reference again to
As is conventional, application of an applied steering torque Ts to the steering wheel 12 results in application of an assisted steering force to the dirigible wheels 84. More particularly, the rack 90 is partly contained within a portion of the steering gear housing 88 comprising the double-acting power cylinder 16. The steering gear housing 88 is in turn fixed to a conventional steering assembly sub-frame 94. The steering assembly sub-frame 94 includes a plurality of mounts 96 for connecting the steering assembly sub-frame 94 to the vehicle chassis (not shown). The dirigible wheels 84 are rotatably carried on wheel spindles 98 connected to the rack 90 via steering knuckles 100 and tie rods 102, and pivotally connected to the host vehicle's chassis and/or steering assembly sub-frame 94 via vehicle struts 104 and lower control arms 106. A portion 108 of each steering knuckle 100 defines a knuckle arm radius about which the assisted steering force, comprising both mechanically derived steering force and powered assist to steering as respectively provided by a pinion-rack interface (not shown) and the double-acting power cylinder 16, is applied.
With reference now to
The block diagram 110 is also useful in that it allows an assessment of the response to a perturbation arising anywhere between the system input (here, the applied steering wheel torque Ts) at input terminal 112 and the system output (here the steering angle or dirigible wheel tire patch angle Thetatp) at output terminal 114. Therefore, while the block diagram 110 will be described in a forward direction from the input terminal 112 to the output terminal 114 (a direction associated with actually steering the vehicle), concomitant relationships in the other directions should be assumed to be present. However, detailed descriptions of such opposite, concomitant relationships are omitted herein for the sake of brevity.
In any case, an applied steering torque T present at terminal 116 and representative of actual torque applied to the torsion bar 52 is subtracted from Ts at a summing point 118. That algebraic sum yields an “error torque” Te, which in this case is the available torque for accelerating the moment of inertia of the steering wheel 12. Te is then divided by (or rather, multiplied by the reciprocal of) the sum of a moment of inertia and damping term (Jss2+Bss) of the steering wheel 12 at block 120 where Js is the moment of inertia of the steering wheel, Bs is steering shaft damping and s is the Laplace variable. The multiplication at the block 120 yields a steering wheel angle Thetas which serves as the positive input to another summing point 122. The negative input to the summing point 122 is a pinion feedback angle Thetap derived in part from the linear motion Xr of the rack 90 at a terminal 124 described below. The summing point 122 yields an error angle Thetae, which when multiplied by the stiffness Ks (at block 126) of the combined steering shaft 78 and torsion bar 52 connecting the steering wheel 12 to the pinion 54 gives the applied steering torque T (at terminal 116) that is substantially present anywhere along the steering shaft 78, input shaft 50 and at the pinion 54. Ks can be considered as a series gain element in this regard. T is fed back from terminal 116 for subtraction from Ts at the summing point 118 in the manner described above. Division of T by the pitch radius Rp of the pinion 54 at block 128 (or rather, multiplication by its reciprocal) gives the mechanical force Fm applied to the rack 90 via the pinion 54.
The total steering force Ft applied to the rack 90 is generated at summing point 130 and is the sum of the mechanical force Fm applied to the rack 90 via the pinion 54 and a hydraulic force Fh provided by the hydraulic assist of the particular system modeled by the block diagram 110. The hydraulic force Fh is derived from the applied steering torque T (again, supplied from terminal 116) in a manner described in more detail below. In any case, the hydraulic force Fh is summed with the mechanical force Fm at summing point 130 to yield the total force Ft in the manner indicated above.
Force applied to the effective steering linkage radius, Fr, taken at terminal 132 is subtracted from the total force Ft at a summing point 134. The resulting algebraic sum (Ft−Fr) from the summing point 134 is divided by (or rather, multiplied by the reciprocal of) a term (Mrs2+Brs) at block 136, where Mr relates to the mass of the rack 90 and Br is a parallel damping coefficient term associated with motion of the rack 90. The resulting product is the longitudinal motion Xr of the rack 90 at terminal 124. Xr is supplied as the positive input to a summing point 138, from which the lateral motion Xh of the steering gear housing 88 is subtracted. The algebraic sum (Xr−Xh) taken at terminal 140 is divided by (or rather, multiplied by the reciprocal of) the pinion radius Rp at block 142 to yield a rotational feedback angle Thetap which serves as the negative input to the summing point 122 as described above.
A time based derivative of the algebraic sum (Xr−Xh) is taken at block 144 and then multiplied by power cylinder piston area A at block 146 to obtain a damping fluid flow Qd which is supplied as a negative input to summing point 148. Concomitantly, the applied steering torque T present at terminal 116 is detected by the torque transducer 32 (at block 150) to obtain an applied torque signal Vat. The applied torque signal Vat is then multiplied by a control function constant Kcf at block 152 to obtain a control function signal Vcf that in turn is supplied as the positive input to summing point 154.
The fluid pressure P (e.g., that is present in the fluid line 18 and at the input slots 56 of the directional control open-center four-way valve 20) at terminal 156 is detected by the pressure transducer 34, which pressure transducer is represented at block 158, in order to obtain feedback pressure signal Vp which is then supplied as the negative input to summing point 154. The error signal Ve formed by the algebraic sum (Vcf−Vp) is filtered (which operation involves multiplying by the inverse of the instant servo valve gain as is preferably accomplished via software control means within the controller 30) at block 160 and amplified at block 162 to obtain a power control signal Vc. The power control signal Vc is then multiplied by the instant valve flow gain factor Kq (e.g., in accordance with the discussion relating to
The lateral motion Xh of the steering gear housing 88 depends upon Ft. More particularly, Ft is a negative input to a summing point 170, from which a force Fhsf present at terminal 172 (e.g., applied to the steering assembly sub-frame 94 as a housing-to-sub-frame force) is subtracted. The lateral housing motion Xh is then determined by the product of the algebraic sum (−Ft−Fhsf) and a control element 1/(Mhs2) at block 174, where Mh is the mass of the steering gear housing 88. Xh is taken from terminal 176 as the negative input to summing point 138 to yield the algebraic sum (Xr−Xh) in the manner described above.
The output tire patch steering angle Thetatp at output terminal 114 is determined by tire patch torque Ttp applied to the tire patches 178 (shown in
The torque Tw applied to the dirigible wheels 84 is determined by the force Fr applied at the effective steering linkage radius (located at terminal 132) multiplied by a control element Rw shown at block 194, where Rw is the effective steering linkage radius of the portion 108 of the steering knuckles 100 defined above. The force Fr is determined in three steps. First, (f Xsf) is subtracted from Xr at summing point 196 with (f Xsf) having been obtained by multiplying (at block 198) the lateral motion Xsf of the steering assembly sub-frame 94 present at terminal 200 by a coupling factor f between the steering assembly sub-frame 94 and mounting points 202 (shown in
The balance of the block diagram 110 models the structural elements disposed in the path of reaction forces applied to the steering gear housing 88, and provides the lateral motion Xsf of the steering assembly sub-frame 94 (at terminal 200) and the housing-to-sub-frame force Fhsf (at terminal 172) mentioned above. Ultimately, the reaction force is applied to the mounting points 202 (at terminal 210) of the dirigible wheels 84 as a sub-frame reaction force Fsf. More particularly, Fsf is determined by the product of a control element (Bsfmps+Ksfmp) shown at block 212 and Xsf at terminal 200, where Ksfmp and Bsfmp stiffness and series damping coefficient terms, respectively, associated with the interface between the steering assembly sub-frame 94 and the mounting points 202. Xsf at terminal 200 is determined by the product of control element 1/(MsfS2+Bsfs) shown at block 214, where Msf is the mass of the sub-frame as well as connected portions of the host vehicle's structure and Bsf is damping associated with coupling the steering assembly sub-frame 94 to the structure, and an algebraic sum (Fhsf−Fsf) generated by summing point 216, where Fhsf is the force applied to the steering assembly sub-frame 94 as the housing-to-sub-frame force located at terminal 172. Fhsf is determined by the product of a control element (BhsfS+Khsf) shown at block 218, where Khsf and Bhsf are stiffness and damping terms associated with the interface between the steering gear housing 88 and the steering assembly sub-frame 94, and an algebraic sum (Xh−Xsf) generated by summing point 220. The positive input to summing point 220, Xh, is taken from terminal 176 while the negative input, Xsf, is taken from terminal 200.
The following values and units for the various constants and variables mentioned above can be considered exemplary for a typical power steering system, and a conventional host vehicle on which it is employed:
1/(Btps+Ktp)=1/(20s+8,000)[rad./in.-lb.]
Bsws+Ksw=30s+500,000[in.-lb./rad.]
1/(Jws2)=1/(8s2)[rad./in.-lb.]
1/(Bss+Jss2)=1/(0.1s+0.5s2)[rad./in.-lb.]
Rw=5[in/rad.]
Kk=8,000[lb./in.]
1/(Mrs2+Brs)=1/(0.02s2+0.1s)[in./lb.]
1/Rp=1/0.315[in.−1]
K=500[in.-lb.]
f=0.7(dimensionless)
A=1.5[in.2]
1/(Mhs2)=1/(0.05s2)[in./lb.]
Bhsfs+Khsf=100s+150,000[lb/in.]
1/(Bsfs+Msfs2)=1/(0.05s+0.4s2)[in./lb.]
Bsfmps+Ksfmp=10s+20,000[lb.in.]
Vt=12[in.3]
Be=100,000[lb./in.2]
Kc=0.1[in.5/lb.-sec.]
Pl, Pc, Pd,=[lb./in.2]
Xf, Xh, Xsf, Xf=[in.]
Fhsf, Fh, Fsff, Ft, Fm, Fh, Fr=[lb.]
T, Ts, Ttp=[in.-lb.]
θs, θe, θp, θw, θtp=[rad.]
It should be noted that the block diagram 110 is a minimal block diagram presented herein for enabling a basic understanding of dynamics of the accumulator enabled power steering system 10. In particular, a more complete representation would include various electronic resistance, electronic inductance, mass and stiffness elements associated with internal operation of the three-way valve 28. It is believed herein however, that these factors can be controlled in an inner feedback control loop separate from the overall feedback loop implemented with reference to the torque transducer. Preferably, the inner feedback control loop would be implemented with reference to the pressure signal Vp representative of actual fluid pressure values present in the fluid line 18 as provided by the pressure transducer 34. This type of control technique is described in detail in the incorporated '254 patent. In addition of course, pertinent servo valve design and control technologies are fully described in the book entitled “Hydraulic Control Systems.”
In passing however, it should be noted that functioning of the three-way valve 28 differs fundamentally from that of a common open-center control valve because the three-way valve 28 is fundamentally flow control device whereas open-center control valves are pressure control devices. In fact, their version of a gain constant Kq′ is actually a pressure gain constant with dramatically differing values that relate valve output pressures to input error angles. In any case, procedures for determining appropriate values for Kq and Kc as utilized herein are fully described in the book entitled “Hydraulic Control Systems.” On the other hand, procedures for determining appropriate values for Kcf over a range of input steering wheel torque and vehicle speed values are fully described in the incorporated '254 patent. Also, a description of procedures for evaluating stability criteria for power steering systems such as the accumulator enabled power steering system 10 as depicted in the block diagram 110 can be found in the incorporated '254 patent and so will not be repeated herein.
In addition, a possible problem wherein foam could form in the fluid due to rapid cycling of the steering wheel 12 should be addressed. This problem could arise due to pressure drop within either side of the double-acting power cylinder 16 relative to reservoir pressure. Such pressure drop could result from backflow through a respective one of the return orifices 72a and 72b of the directional control open-center four-way valve 20 when rapidly recovering from a turn. Although this problem could theoretically be solved by slightly pressurizing the reservoir 24, that possible solution is discounted herein because the reservoir 24 would likely have to be vented to the atmosphere in view of relatively large exchanges of fluid between the reservoir 24 and the accumulator 26 occurring during normal operation of the accumulator enabled power steering system 10. A more practical solution is to provide a pair of check valves 252 fluidly connected between the reservoir 24 and each of the left turn tube 22a and the right turn tube 22b as shown in
Finally as depicted in the flow chart of
Having described the invention, however, many modifications thereto will become immediately apparent to those skilled in the art to which it pertains, without deviation from the spirit of the invention. For instance, the three-way valve 28 could be formed with multiple holes defining input and return “ports” in place of the input grooves 40 and return grooves 44, thereby almost certainly lowering fabrication costs. Thus, such an over-lapped servo valve could, albeit with possibly some degradation of performance, be used in place of the three-way valve 28 having the input grooves 40 and return grooves 44 as depicted in
The instant system is capable of providing accumulator enabled power steering systems intended for medium through large vehicles, and accordingly finds industrial application both in America and abroad in power steering systems intended for such vehicles and other devices requiring large values of powered assist in response to torque applied to a steering wheel, or indeed, any control element functionally similar in nature to a steering wheel. Alphanumeric identifiers on method steps in the claims are for convenience in reference by dependent claims and do not signify a required order of performance of the method steps unless explicitly stated in the claims.
This application claims priority to U.S. Provisional Application Ser. No. 60/620,079 filed Oct. 18, 2004.
| Number | Date | Country | |
|---|---|---|---|
| 60620079 | Oct 2004 | US |
