EP1497165A2 - Mecanisme de direction hydraulique - Google Patents

Mecanisme de direction hydraulique

Info

Publication number
EP1497165A2
EP1497165A2 EP03720170A EP03720170A EP1497165A2 EP 1497165 A2 EP1497165 A2 EP 1497165A2 EP 03720170 A EP03720170 A EP 03720170A EP 03720170 A EP03720170 A EP 03720170A EP 1497165 A2 EP1497165 A2 EP 1497165A2
Authority
EP
European Patent Office
Prior art keywords
control
pressure
steering
valve
spring
Prior art date
Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
Withdrawn
Application number
EP03720170A
Other languages
German (de)
English (en)
Inventor
Edwin Harnischfeger
Erhard Bergmann
Walter Scandella
Vincenzo Domenico Bollero
Current Assignee (The listed assignees may be inaccurate. Google has not performed a legal analysis and makes no representation or warranty as to the accuracy of the list.)
Bosch Rexroth AG
Original Assignee
Bosch Rexroth AG
Priority date (The priority date is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the date listed.)
Filing date
Publication date
Priority claimed from DE10307943A external-priority patent/DE10307943A1/de
Application filed by Bosch Rexroth AG filed Critical Bosch Rexroth AG
Publication of EP1497165A2 publication Critical patent/EP1497165A2/fr
Withdrawn legal-status Critical Current

Links

Classifications

    • B—PERFORMING OPERATIONS; TRANSPORTING
    • B62—LAND VEHICLES FOR TRAVELLING OTHERWISE THAN ON RAILS
    • B62D—MOTOR VEHICLES; TRAILERS
    • B62D5/00—Power-assisted or power-driven steering
    • B62D5/06—Power-assisted or power-driven steering fluid, i.e. using a pressurised fluid for most or all the force required for steering a vehicle
    • B62D5/09—Power-assisted or power-driven steering fluid, i.e. using a pressurised fluid for most or all the force required for steering a vehicle characterised by means for actuating valves
    • B62D5/093—Telemotor driven by steering wheel movement

Definitions

  • the invention relates to a hydraulic steering device according to the preamble of patent claim 1.
  • Such a steering device known for example from US Pat. No. 3,566,749, is used in hydraulic steering circuits of vehicles and mobile work machines with high axle loads and comparatively low driving speeds ( ⁇ 50 km / h).
  • pressure medium is conveyed to steering cylinders via a metering device operating according to the gerotor principle, depending on a steering angle set via a steering wheel or a joystick.
  • pressure medium is conveyed from a pump via a control block to the steering cylinders, the control block being activated as a function of the outlet pressure of the metering device, i.e. a control pressure is present in the control rooms of the control block which corresponds approximately to the pressure which is applied to the steering cylinders to apply the required steering torque.
  • the invention is based on the object of avoiding leakage in the area of the control rooms of a control block of a steering device with a minimal expenditure on device technology.
  • the steering device has a control device by means of which the control pressure for actuating a control block which controls the pressure medium supply to a steering cylinder can be reduced substantially below the pressure which is required for actuating steering cylinders of the steering.
  • This pressure reduction in the control pressure of the control block means that the use of high-pressure covers can be dispensed with, so that the hydraulic steering device can be produced much more cost-effectively than the known solution described above, with the same functional reliability.
  • the control block used in the steering device contains a proportional valve which determines the pressure medium direction and speed and an associated pressure compensator, via which the pressure difference across the measuring orifice formed by the proportional valve is kept constant regardless of fluctuations in the load or pump pressure.
  • the control rooms which are effective for controlling the proportional valve are supplied with pressure medium from a control pump via a control valve, the control position of the control valve being set as a function of the pressure downstream of the metering device.
  • the control pump delivers a much lower pressure as the pump supplying the steering cylinder, ie according to the invention a control circuit is used which is operated at a comparatively low pressure.
  • the pressure difference for actuating the control valve is tapped at a nozzle which is arranged in a working line between a working connection of the metering device and the associated steering cylinder.
  • control rooms described above are also supplied with control oil, which is guided in a control circuit with a lower pressure level than the main circuit supplying the steering cylinders.
  • This control circuit is acted upon by a control pump with pressure medium, the control pressure difference for actuating the control block being applied by means of a control metering device arranged parallel to the metering device.
  • the metering device works according to the gerotor principle, the control metering device being formed by a conveying element which is placed on a shaft of the metering device.
  • the output pressure provided by the control pump can be changeable, for example via an adjustable pressure relief valve.
  • the operating behavior can take place by varying the spring preload of the pressure compensator of the control block.
  • this prestressing of the pressure compensator can be changed by nozzles connected in parallel, a nozzle with a specific cross section being effective, depending on the desired prestressing and the pretension is set as a function of the pressure drop across the effective nozzle.
  • the continuously adjustable directional control valve of the LS control block of the steering device is usually designed with a positive overlap, so that a certain minimum stroke of the control slide is required to drive over this overlap. This is biased into its zero position by means of spring arrangements.
  • the spring rate of the spring arrangement should be chosen to be as flat as possible so that the pressure range required to drive over the overlap is as small as possible.
  • the spring rate of the spring arrangement should be selected so that the largest possible pressure range can be covered for fully opening the directional control valve and the steering can be adequately supplied with pressure medium even with fast steering movements.
  • a steeper spring characteristic is set by connecting at least two springs in parallel.
  • the engagement of the second spring increases the overall spring rate and thus the relative stroke of the valve spool as a function of a control pressure difference smaller than would be the case if only one spring was active.
  • the first spring of the spring arrangement which is effective from the start, is received with a preload that is less than that of the known solutions described at the beginning.
  • the spring rate of the second spring is preferably carried out with a lower spring rate than the first spring.
  • the second spring is guided coaxially with the first spring and is carried along by a driving shoulder after a predetermined stroke of the control slide.
  • the steering quantity increase can be adapted to a certain extent to the present operating conditions by means of a valve arrangement.
  • This valve arrangement for changing the steering quantity increase can comprise, for example, a proportionally adjustable pressure relief valve and two check valves which act in opposite directions and are each assigned to a control side of the control block. The maximum control pressure difference applied to the control block - and thus the steering quantity increase - can be changed via the pressure relief valve.
  • control block is designed with a directional valve spring-biased into a central position
  • this can be preceded by a device which ensures, even at low steering speeds, that the directional valve with a predetermined one minimal control pressure difference is applied and the preload of the centering springs is overcome.
  • FIG. 1 shows a first exemplary embodiment of a hydraulic steering device in which a control block can be acted upon by a control pressure difference by means of a control metering device;
  • Figure 2 shows an alternative embodiment in which the control block can be acted upon by a control pressure difference via control valves
  • Figure 3 is a sectional view of a control valve of the embodiment of Figure 2;
  • FIG. 4 flow characteristics of the control block from FIG. 2 with different preloads of a pressure compensator of the control block;
  • FIG. 5 shows a side view of a metering device of the exemplary embodiment from FIG. 1;
  • FIG. 6 shows a section through an LS control block which can be used in a steering device according to FIGS. 1 and - in modified form - FIG. 2;
  • FIG. 7 shows a detailed representation of the control block from FIG. 6;
  • FIG. 8 spring characteristics of a spring arrangement of the control block from FIG. 6;
  • FIG. 9 shows a circuit diagram of a third exemplary embodiment of a hydraulic steering device.
  • FIG. 10 is a circuit diagram of a fourth embodiment of a hydraulic steering device.
  • FIG. 1 shows a circuit diagram of a first exemplary embodiment of a steering device 1.
  • This has a steering cylinder 2 designed as a synchronous cylinder, via which the steering torque required to turn the steered wheels is applied.
  • the pressure medium supply to the steering cylinder 2 takes place via a steering unit 4 which in principle consists of a metering pump and a manually operated servo valve of the rotary slide type.
  • the servo valve and the metering pump are actuated via the steering column connected to a steering wheel of the vehicle. Since the steering unit 4 is a standard component, reference is made to the relevant prior art with regard to the structure, for example DE 199 28 530 AI.
  • the steering unit 4 has a pressure connection P which is connected to a steering pump (not shown in FIG. 1). This pressure port P is shut off when the steering wheel is not actuated.
  • Two output connections L, R of the steering unit 4 are connected to the two cylinder spaces 10, 12 of the steering cylinder 2 via working lines 6, 8.
  • pressure medium is then conveyed into the cylinder space 10 via the connection L and the working line 6, while the pressure medium is guided out of the cylinder space 12 via the working line 8 and the connection R back to the steering unit 4.
  • the steering unit 4 also had an LS connection via which a load signal for controlling a priority valve or a pump can be tapped.
  • the steering assembly 4 is often designed such that it alone is not able to deliver the amount of pressure medium required to actuate the steering cylinder 2.
  • an additional amount of pressure medium is led to the steering cylinders 2 via an LS control block 14 in order to increase the steering quantity, the control block 14 being activated as a function of the actuation of the steering unit 4.
  • the pressure medium supply is designed so that three to five Steering wheel turns from Stop to stop of the steering cylinder can be steered.
  • the control block 14 used in the steering device according to FIG. 1 has a proportionally adjustable directional valve, via which the pressure medium speed and direction of the pressure medium flow are specified.
  • the pressure medium speed is adjusted via a measuring orifice, to which a pressure compensator of the control block is assigned.
  • the control block 14 can also be a standard component, as described, for example, in the Bosch Rexroth data sheet RE 64 282 / 10.99 or in DE 197 15 021 AI.
  • Pressure changes at the hydraulic consumers of the hydraulic system or at the pump are corrected by means of the pressure compensator so that the pressure medium volume flow to the consumer is kept constant even with different loads.
  • the pump is controlled depending on the load pressure. If the steering and the other working hydraulics are supplied by a common pump, a priority valve is used to ensure the preferred supply of the steering unit with pressure medium.
  • a variant of the control block 14 is explained with reference to FIGS. 6, 7 and 8.
  • the control block 14 shown in Figure 1 has a pressure connection P connected to the steering pump (or the central pump for supplying the working hydraulics and the steering), a tank connection T, an LS connection for tapping the load pressure on the steering cylinder 2 and two
  • Consumer ports A, B. are connected to the cylinder rooms 10 and 12 via consumer lines 16, 18.
  • the control block 14, or more precisely, the proportional valve which determines the flow direction and size of the pressure medium volume flow, is actuated via a control circuit 20 which is connected via a supply pump 22 is supplied with control oil.
  • the control circuit 20 is connected to the two control or spring spaces 24, 26 of the proportional valve of the control block 14, so that the displacement of a main slide of the proportional valve takes place as a function of the pressure difference applied to the control spaces 24, 26.
  • the steering assembly 4 has a metering pump which works on the gerotor principle.
  • a second gear set is placed, which forms a control metering device 28, via which, depending on the steering wheel speed, control oil can be conveyed in the direction of one of the control spaces 24, 26.
  • Control lines 30, 32 which lead to the control rooms 24 and 26, are connected to the two connections A, Ai of the control metering device 28. From these two control rooms 24, 26 return lines 34, 36 each lead back to the associated connection A or Ai. In each of these return lines 34, 36 there are arranged a nozzle 38 and 40 as well as a check valve 42, 44 which opens in the direction of the connections A, Ai.
  • the oil pumped by the feed pump 22 is fed in via a connecting line 46 connecting the two return lines 34, 36, so that when the steering wheel is not actuated, both branches of the control circuit 20 and thus the control spaces 24, 26 are acted upon by the pressure supplied by the feed pump - that Proportional valve of the control block 14 remains in its spring-biased basic position.
  • control oil is shifted within the control circuit 20 in accordance with the steering speed and the number of steering wheel rotations.
  • control oil is conveyed via the connection A of the control metering device 28 into the control line 30 and from there to the control chamber 24.
  • the control oil is then returned to the control line 32 via the return line 34, the nozzle 38, the connecting line 46 and the check valve 44.
  • the feed pump 22 is assigned an adjustable pressure relief valve 48, via which the pressure in the control circuit 20 can be set.
  • the pressure limiting valve 48 By suitable actuation of the pressure limiting valve 48, different pressures in the control circuit 20 can be selected, so that different steering speeds can be achieved depending on the operating state of the vehicle.
  • the control oil volume flow led via the control block 14 can be changed with the same steering wheel lock, so that the steering characteristic is adapted, for example, to the driving speed of the vehicle
  • Vehicle can be adjusted.
  • Characteristic curves of this type are shown, for example, in FIG. 4, different control oil delivery quantities Q being adjustable as a function of the speed of an engine of the vehicle (for example diesel engine), the characteristics can be preselected by changing the maximum control pressure limited by the pressure relief valve 48.
  • the pressure limiting valve 48 is adjusted by means of a pressure sensor 50.
  • the feed pump 22 is designed in such a way that the control pressure supplied by it is substantially below the steering pressure generated by the steering unit 4, so that the pressure applied to the control rooms 24, 26 is comparatively low and the sealing of the control rooms / spring chambers is thus much easier than with conventional ones Solutions in which high-pressure covers must be used, since a control pressure corresponding to the steering pressure (more than 100 bar) is present in the control rooms 24, 26. It was found that the variant according to FIG. 1 can already be operated with control pressures in the range of 4 bar, so that the feed pump 22 can be made very small.
  • the steering assembly 4 is in principle a standard product, the housing 52 of which is shown schematically in FIG.
  • the servo valve, designed as a rotary slide valve, and the gear set of the associated gerotor (metering pump), indicated by 54, are accommodated in this.
  • the second gear set which forms the control metering device 28 is placed on the elongated shaft of the gerotor via an intermediate disk 56, in which the connections A, Ai are also formed.
  • An end plate 58 is then marked on this intermediate disk 56, so that the steering unit 4 and the control metering device 28 can be produced very simply and compactly by converting a conventional steering unit.
  • FIG. 2 shows a variant of the concept according to the invention.
  • the connections L, R of the steering assembly 4 are connected to the steering cylinders 2 'and 2' 1 via working lines 6, 8, similar to the exemplary embodiment described above.
  • working lines 6 not a single synchronous cylinder 2, but two differential cylinders 2 ', 2''arranged in parallel are used.
  • the working line 6 opens into a cylinder space 60 ′′ of the steering cylinder 2 ′′ and into an annular space 62 ′ of the steering cylinder 2 ′, while the working line 8 leads into the annular space 62 ′′ of the steering cylinder 2 ′ 1 and into the cylinder space 60 ′ of the steering cylinder 2 'opens out.
  • an LS control block 14 is again provided, via which additional pressure medium can be conveyed from a steering pump to the steering cylinders 2 ', 2' 1 .
  • the control of the proportional valve (not shown) of the control block 14 takes place via its own control circuit 20, in which a control pressure can be built up which is substantially below the load pressure on the steering cylinders 2 ', 2' 1 .
  • a control pressure that is comparatively low is controlled via the control circuit 20, so that the sealing of the control spaces 24, 26 is not difficult.
  • the control circuit has two control lines 30, 32 which are connected to the pressure connection of the control pump 22.
  • the maximum pressure in the control circuit 20 is increased to a maximum value via a control pressure limiting valve 64, for example limited to 35 bar (the maximum steering pressure is over 100 bar).
  • a control valve 66 or 68 is connected in each control line 30, 32, via which the control pressure in the respectively assigned control chamber can be regulated depending on the direction of the control oil flow.
  • the two control valves 66, 68 are each biased by a control spring into a basic position in which an input port P of the control valve is connected to a working port A.
  • a connection to a tank connection T is opened.
  • control valves 66, 68 are each controlled via two control lines 70, 72, so that the valve slide is shifted into a control position as a function of the pressure difference in the control lines 70, 72.
  • a nozzle 74, 76 is arranged in each of the two working lines 6, 8, the control line 72 opening into the section between the nozzle 74 and the associated working connection L, R, while the control line 70 opens into a line section downstream of the nozzles 74, 76 , That is, the control valves 66, 68 are acted upon by a control pressure difference which corresponds to the pressure drop across the nozzles 74 and 76, respectively.
  • FIG. 2 shows the pressure compensator 78 assigned to the proportional valve (not shown) of the control block 14.
  • this pressure compensator 78 makes the pressure drop across the measuring orifice formed by the proportional valve independent of fluctuations.
  • the pressure on the steering cylinders 2 or the steering pump is kept constant.
  • the bias of a control spring 80 of the pressure compensator 78 can be changed via a biasing device 82.
  • This has a piston 84 which can be acted upon in the "increase preload” direction by the pressure in a preload line 86 and in the "decrease preload” direction by the pressure in a further preload line 88.
  • the pressure difference in the two prestressing lines 86, 88 is varied by means of a prestressing valve device 90, so that different prestressing of the control spring 80 of the pressure compensator 78 can be set depending on the operating state.
  • the same characteristic curve characteristic as shown in FIG. 4 can be selected in this way, the individual characteristic curves being variable, for example, as a function of the selected driving gear.
  • the biasing valve device 90 is acted upon by a pump 92 with pressure medium.
  • the preload valve device has three parallel preload nozzles 94, 96, 98, each of which has different nozzle diameters. It is assumed that the effective diameter of the nozzle 94 is the smallest and that of the nozzle 98 the largest.
  • the nozzles 94, 96, 98 are each arranged in a parallel branch, in each of which a switching valve 100, 102, 104 is arranged upstream of the respective nozzle, so that when a switching valve is switched into its open position, the respectively assigned nozzle in the control oil flow flow d is switched.
  • a control line 106 containing the nozzle 94 and the switching valve 100 and a control line 108 containing the nozzle 96 and the switching valve 102 are connected to the inputs of a shuttle valve 110 between the nozzle and the switching valve, whose output is in turn connected to the input of a further shuttle valve 112.
  • the other input of this shuttle valve 112 is correspondingly connected to the control line 114 containing the nozzle 98 and the switching valve 104.
  • the prestressing line 88 opens into the outlet of this shuttle valve 112.
  • the pressure in the prestressing line 86 connected downstream of the nozzles is limited to an adjustable fixed value via a pressure limiting valve 122.
  • the switching valve 100 is brought from its blocking position into its open position so that the control oil delivered by the pump 92 flows through the nozzle 94 to the pressure limiting valve 122. downstream of the nozzles and thus in the prestressing line 86, the pressure is kept at the predetermined value by the pressure limiting valve 122.
  • the pressure upstream of the nozzle 94 is present at the input of the shuttle valve 110 and is reported via this and the shuttle valve 112 into the bias line 88. That is, the piston 84 is subjected to a pressure difference which corresponds approximately to the pressure drop across the respectively active nozzle 94, 96, 98.
  • the nozzle 94 is designed with a comparatively small diameter, so that the pressure acting in the direction of relaxation of the control spring 180 is comparatively large and the preload of the pressure compensator 78 and thus the pressure difference across the metering orifice of the LS control block is lower than when it was activated of the other nozzles 96, 98 - the flatter characteristic curve is obtained in the illustration according to FIG. 4.
  • the nozzles 96 or the nozzle 98 are then opened. This change in the preload can understandable can also be controlled via other suitable valve devices.
  • variable steering speed by changing the preload of the pressure compensator 78 can also be used in the concept shown in FIG. 1 in order to implement different control characteristics.
  • FIG. 3 shows a schematic section through a control valve 66, as can be used in the circuit according to FIG. 2.
  • the control valve 66 has a valve spool 124 which is biased by a spring 126 into a basic position, not shown, in which the pressure port P is connected to the working port A via a connecting bore 128 of the valve spool 124.
  • the valve spool 124 has a radially projecting piston collar 130 which is guided in a radially widened annular space.
  • the partial space 132 located at the top in FIG. 3 and acting in the direction of the spring 126 is acted upon by the pressure upstream of the respective measuring orifice 74 (76) via the control line 72, while the lower partial space 134 active in the direction "connection to the tank" is applied via the control line 70 is subjected to the pressure downstream of the nozzle 74, 76.
  • valve spool 174 sets itself into a control position in which the connection from P to A and / or T is opened.
  • the pressure supplied by the control pump 22 is thus reduced to a control pressure via the control valve 66, 68, via which the valve slide of the proportional valve of the control block 14 is brought into its position dependent on the steering wheel angle. That is, when the steering wheel is turned, a quantity of pressure medium dependent on the steering wheel turn and the steering wheel speed is conveyed to the respective steering cylinder 2 ', 2' 1 via the steering unit 4.
  • control block 14 is acted upon by a control pressure as a function of the pressure difference across the associated nozzle 74, 76 via the control valve 66 or 68 and brought into an open position in which an additional quantity of pressure medium from the steering pump P to the steering cylinders 2 ', 2 ′′ is promoted so that a steering torque corresponding to the steering wheel angle and the steering wheel speed is built up as a function of the set steering characteristic (preload of the pressure compensator 78).
  • FIG. 6 shows a section through an LS control block 14.
  • Such an LS control block has at least one valve disk or valve housing 136, in which the two working connections A, B assigned to a consumer and the connections T, P (not shown), the LS connection and control connections are formed.
  • the valve housing 136 has a valve bore 138, in which a control slide 140 of the continuously adjustable directional valve 142 is guided. This is preceded by a pressure compensator 144, the control piston 146 of which is prestressed in the closing direction by a comparatively weak control spring 148.
  • the control piston 146 is further acted upon in the closing direction by the load pressure present downstream of the directional valve 142 forming a measuring orifice with a constant cross-section and in the opening direction by the pressure acting at the inlet of the pressure compensator 144, which corresponds approximately to the pressure at the pump connection P.
  • This pressure connection P is connected via a pressure channel 150 a pressure compensator inlet chamber 152 of a pressure compensator bore 154 receiving the control piston 146.
  • This pressure present in the pressure compensator inlet chamber 152 acts via an inner bore of the control piston 146 on the right-hand end face of the control piston 146 in FIG. 6.
  • the valve bore 138 is provided with a plurality of annular spaces which form an inlet chamber 156, two outlet chambers 158, 160 and two tank chambers 162, 164 connected to the tank connection T.
  • the control slide 140 is designed with a plurality of ring collars 166, 168 and 170, via which the connection between the aforementioned chambers can be opened or closed.
  • the control slide 140 also has an axial bore 172 which is closed at the end and which opens out via two transverse bores 180, 182 in the outer circumference of the control collar 166 and 168, respectively.
  • the control slide is biased into its basic position via two spring arrangements 174, 176.
  • the two drain chambers 158 and 160 are connected to the adjacent tank chambers 162 and 164, respectively.
  • the transverse bores 180 and 182 opening in the axial bore 172 are blocked off. Furthermore, the connection from the inlet chamber 156 to the adjacent outlet chambers 160 and 158 is also controlled.
  • valve housing 136 two secondary valves 184, 186 are accommodated in the valve housing 136, via which a connection from the drain chambers 158, 160 to the tank chambers 162 and 164 can be opened.
  • this valve disk of the control block 14 also has an LS pressure limiting valve 188 and a shuttle valve 190, by means of which the highest load pressure that is effective at several consumers can be tapped.
  • the control of the proportionally adjustable directional control valve 142 takes place via two indicated pressure reducing valves 192, 194, via which spring spaces 178, 179 in the spring arrangements 176 and 174 can be acted upon with a control pressure difference, so that the control slide 140 is shifted from its basic position and a measuring orifice with a constant one Flow cross section is controlled.
  • Control piston is acted upon in the closing direction by the pressure in the axial bore 172.
  • the two control collars 166, 168 are designed such that the connection from the inlet chamber 156 to the adjacent outlet chamber 158 or 160 is only opened after a predetermined stroke - i.e. , In the basic position, the directional control valve is designed with a positive overlap.
  • connection between the inlet chamber 156 and the outlet chamber 158 is opened by the control collar 166.
  • connection between the tank chamber 162 and the drain chamber 158 is activated and the connection between the other tank chamber 164 and the drain chamber 150 controlled so that pressure medium can flow through the working port A to the cylinder chamber 10 of the steering cylinder 2 and from the cylinder chamber 12 via the working port B to the tank T.
  • the pressure compensator is acted upon by the pressure at the pressure port P in the opening direction (in the direction of a "direction" opening the connection of the pressure compensator inlet chamber 152 to a pressure compensator discharge chamber 198 connected to the inlet chamber 156), while in the closing direction the force of the control spring is applied 48 and the pressure in the discharge chamber 158 or when connecting several consumers, the highest load pressure acts ..
  • the control piston 146 sets itself in a control position in which the pressure drop across the open orifice (directional valve 142) with a constant orifice opening is kept constant regardless of load pressure.
  • the directional control valve 142 should already respond at very low control pressures, ie at correspondingly low speeds of the gerotor of the steering assembly 4, and thus enables a suitable increase in the steering quantity.
  • the spring arrangement 174, 176 could be designed, for example, with a low spring preload and / or a low spring rate, so that the directional control valve 142 is opened even with small control pressure differences.
  • the directional valve 142 would, however, already with a comparatively low control Open the pressure difference completely, so that the passable pressure range would be too small for the steering device to function adequately.
  • a spring arrangement 174, 176 is used in the exemplary embodiment shown in FIG. 6, as is shown in detail in FIG. 7, the left spring arrangement 176 being described by way of example.
  • This has two springs 200, 202 connected in parallel.
  • the first spring 200 is supported on an end face 206 of the valve housing 136 via a spring plate 204.
  • the other end of the first spring 200 is supported on an inner shoulder 208 of an attached housing 210 and engages around a bolt-shaped stroke limiter 212 screwed into this housing 210.
  • This first spring 200 is received with a predetermined preload between the spring plate 204 and the inner shoulder 208. In the event of an axial displacement of the control slide 140 to the left, after overcoming this prestress, the spring plate 204 is lifted off the end face 206 and the first spring 200 is compressed accordingly.
  • the second spring 202 is supported on a radial shoulder 214 of the stroke limiter 212 and extends beyond a radially recessed end section 216 of the stroke limiter 212 up to an axial pin 218 des
  • This axial pin 218 is from the end cut around the second spring 202, which ends at a predetermined axial distance h from a driving shoulder 220 of the spring plate 204.
  • this driving shoulder 220 runs onto the adjacent end of the second spring 202, so that after this stroke h, the displacement of the control slide 140 against the force of the first spring 200, which is loaded with preload and the force of the spring 202 takes place.
  • the stroke h described above is selected such that the second spring 202 is only effective when the stroke overlap described at the beginning has been exceeded.
  • the resulting spring characteristic is shown in FIG. 8.
  • the control pressure acting on the directional control valve 142 is shown over the stroke of the control spool 140.
  • the first spring 200 is received with a pretension, so that a stroke of the control slide 140 does not occur until the effective control pressure has reached a value po corresponding to this pretension.
  • this control pressure pg can be in the range of one to two bars in a steering device, for example, the control slide 140 is displaced depending on the control pressure in accordance with the approximately linear characteristic curve of the first spring 200.
  • the overlap of the directional valve 142 has been overrun and the driving shoulder 220 runs onto the second spring 202, so that both springs are effective from this stroke h.
  • a steeper characteristic curve is set, so that the maximum stroke h max is reached when a maximum control pressure Pmax is applied, which is substantially greater than in the case in which the spring 200 would have been effective alone dash-dotted line in Figure 8).
  • the control range can thus be expanded significantly compared to a simple spring with a low spring rate, only a comparatively small control pressure difference p ⁇ - pg being required to drive over the overlap.
  • the control characteristic of the control block can be optimally adapted to the requirements, in particular in the case of a steering device.
  • such a spring arrangement can also be used in directional valve arrangements for other applications.
  • the second spring 202 has a lower spring rate than the spring 200 - in principle, of course, the spring rate of the spring 202 could also be selected to be greater than that of the spring 200. Instead of two springs, several springs could also be effective to optimize the spring characteristic.
  • a certain, preset pressure is maintained in the lines 34, 36 via the pressure relief valve 48 located downstream of the nozzles 38, 40.
  • a pressure is established in the line 30 or 32 such that there is a pressure difference corresponding to the metered quantity via the nozzle 38. If the setting value on the pressure limiting valve 48 is increased, the pressure in the line 30 or 32 also becomes correspondingly higher. This means that for a given metering quantity, the pressure difference between lines 30 and 32 and thus the pressure difference applied to control rooms 24 and 26 of directional control valve 40 remain essentially constant, regardless of the setting of pressure limiting valve 48. By changing the limit pressure at the pressure relief valve 48 the steering quantity increase cannot therefore be changed for a given metering quantity.
  • this increase in the steering quantity can be changed within certain limits for a given metered quantity.
  • the exemplary embodiment shown in FIG. 9 corresponds to the variant shown in the basic concept of FIG. 1 and differs from this solution by a valve arrangement 222 for changing the steering quantity increase and a device 224 for improving the response behavior at low steering speeds.
  • the valve arrangement 222 for changing the steering quantity increase has called a proportionally adjustable pressure relief valve in the following boost valve 226, which can be adjusted electromagnetically, for example depending on the signal from a sensor 50.
  • the input of the boost valve 226 is connected via a control channel 228 to a connecting channel 230 which runs parallel to the connecting line 46.
  • Two check valves 232, 234 are provided in the connecting channel 230, which enable a control oil flow from the line 30 or 32 to the reinforcing channel 228 and block them in the opposite direction.
  • the two counter-acting check valves 232, 234 ensure that the pressure in the two lines 30 and 32 and thus in both control rooms 24 and 26 can be limited with only one pressure relief valve via the boost valve 226.
  • the two check valves 42, 44 are arranged upstream of the check valves 232, 234 of the valve arrangement 222. By changing the limit pressure to be set in valve 226, the control pressure in control rooms 24, 26 can be limited to values of different heights.
  • the pressure limiting valve 48 in contrast to the exemplary embodiment described with reference to FIG. 1, is not designed as a proportionally adjustable valve, but rather as a simple valve that can be set to different values, for example in the range of 1-2 bar, depending on the application , so that it is ensured that the lines are always supplied with pressure medium.
  • the device 224 has nothing to do with the change in the steering quantity amplification as such, but serves to improve the response behavior of the control block 14.
  • the device 224 has two pressure differential valves 240, 242 arranged downstream of the nozzles 38 and 40, which are designed as check valves in the exemplary embodiment shown in FIG.
  • Such pressure difference valves 240, 242 produce a substantially constant pressure difference between their inlet and their outlet.
  • the closing springs of the check valves 240, 242 are set to a value equivalent to the pretension of the spring arrangements 174, 176, so that the check valves 240, 242 at a low speed of the metering device 28 accumulate the metered quantity so strongly that the Spring preload is overcome and the control slide 140 of the directional control valve 142 is pushed out of the central position against the force of the spring arrangement 174, 176 even at low steering speeds - the control block 14 reacts even at low speeds of the metering devices.
  • the pressure difference valves 240, 242 can also be arranged upstream of the nozzles 38 and 40, since, regardless of the arrangement of the pressure difference valves, the constant pressure drop across the pressure difference valve is added to the pressure drop across the upstream or downstream nozzle.
  • a somewhat different line routing is selected than in the exemplary embodiment according to FIG. 9.
  • the two return lines 34, 36 are replaced by a central channel 244, which extends between the two control lines 30, 32.
  • a nozzle 38 is arranged, which takes over the function of the nozzles 38, 40 from the exemplary embodiments described above.
  • a pressure differential valve 240, 242 is arranged on each side of the nozzle 38. However, these are not designed as check valves but as piston valves.
  • a pressure difference valve 240, 242 designed as a piston valve the dependence of the pressure difference on the flow rate is less than in the case of a design as a check valve.
  • Such a piston valve is described, for example, in the applicant's patent application DE 199 04 616 AI. With regard to the function of such a piston valve, reference is therefore made to these explanations for the sake of simplicity.
  • the two pressure difference valves 240, 242 can each be bypassed via a bypass channel 246, 248, in each of which a check valve 250, 252 is arranged, which are connected antiparallel to the pressure difference valve 240, 242.
  • These anti-parallel check valves 250, 252 make it possible to use a single nozzle 38 instead of 2 nozzles, as in the previously described exemplary embodiments.
  • the pressure difference valves 240, 242 are each acted upon in the opening direction by the pressure in the bypass channel 246 or 248 and in the closing direction by the force of a spring and the respective downstream pressure in the central channel 244, so that an essentially constant pressure difference is generated between their inlet and their outlet ,
  • valve arrangement 222 for changing the steering quantity increase has the same structure as the embodiment shown in FIG. game so that further explanations can be omitted.
  • a check valve can of course also be used as a pressure differential valve 240, 242 instead of the piston valve.
  • a piston valve could also be used in the variant shown in FIG.
  • a hydraulic steering device with increased steering quantity a metering device for supplying pressure medium to steering cylinders being actuated as a function of a steering wheel lock.
  • An additional amount of pressure medium can be fed to the steering cylinders via a control block, the control block being actuated according to the invention via a control circuit which has a control pressure which is substantially below the load pressure for transmitting the steering torque to the steering cylinder.
  • control line 32 control line

Abstract

L'invention concerne un mécanisme de direction assistée hydraulique, dans lequel une unité de dosage servant à acheminer un fluide sous pression à des vérins de direction est actionnée en fonction d'un angle de rotation du volant. Une quantité supplémentaire de fluide sous pression peut être acheminée aux vérins de direction par l'intermédiaire d'un bloc de commande qui est piloté, selon l'invention, par l'intermédiaire d'un circuit de commande sur lequel agit une pression de commande sensiblement inférieure à la pression de charge servant à transférer le couple de braquage au vérin de direction. L'invention concerne en outre un ensemble de ressorts destiné au bloc de commande, constitué de deux ressorts montés en parallèle, l'un desdits ressorts étant actif uniquement lorsque le tiroir de commande s'est déplacé selon une course prédéterminée. Il est ainsi possible d'établir une courbe caractéristique selon laquelle : le tiroir de commande est précontraint dans sa position de repos avec une précontrainte minimale ; un chevauchement éventuel peut être dépassé au moyen d'une faible différence de pression de commande ; et la plage de pression de commande servant à activer le bloc de commande est suffisamment large.
EP03720170A 2002-04-25 2003-03-13 Mecanisme de direction hydraulique Withdrawn EP1497165A2 (fr)

Applications Claiming Priority (7)

Application Number Priority Date Filing Date Title
DE10218639 2002-04-25
DE10218639 2002-04-25
DE10236557 2002-08-08
DE10236557 2002-08-08
DE10307943 2003-02-25
DE10307943A DE10307943A1 (de) 2002-04-25 2003-02-25 Hydraulisch Lenkeinrichtung
PCT/DE2003/000824 WO2003091082A2 (fr) 2002-04-25 2003-03-13 Mecanisme de direction hydraulique

Publications (1)

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EP1497165A2 true EP1497165A2 (fr) 2005-01-19

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EP03720170A Withdrawn EP1497165A2 (fr) 2002-04-25 2003-03-13 Mecanisme de direction hydraulique

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US (1) US7306071B2 (fr)
EP (1) EP1497165A2 (fr)
WO (1) WO2003091082A2 (fr)

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JP4698371B2 (ja) * 2005-10-04 2011-06-08 株式会社ショーワ 油圧式動力舵取装置
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WO2015037600A1 (fr) * 2013-09-10 2015-03-19 テイ・エス テック株式会社 Dispositif de siège pour véhicule
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Publication number Publication date
US20050161278A1 (en) 2005-07-28
US7306071B2 (en) 2007-12-11
WO2003091082A2 (fr) 2003-11-06
WO2003091082A3 (fr) 2004-01-29

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