EP0511231A1

EP0511231A1 - Hydraulic control device

Info

Publication number: EP0511231A1
Application number: EP91901688A
Authority: EP
Inventors: Helmut Rembold; Martin Müller
Original assignee: Robert Bosch GmbH
Current assignee: Robert Bosch GmbH
Priority date: 1990-01-16
Filing date: 1991-01-09
Publication date: 1992-11-04
Also published as: KR920703973A; WO1991010813A1; JPH05503562A; DE4037824A1

Abstract

Dispositif de réglage hydraulique présentant un cylindre différentiel (11) à piston différentiel (12) dont la chambre de pression (13) est alimentée en permanence par une pompe (16) du côté de la petite surface du piston, la chambre de pression (18) du côté surface du piston (18) étant alimentée par une soupape de commande (20) électromagnétique. Celle-ci présente un recouvrement négatif en position neutre. Etant donné que la pompe est actionnée à vitesse variable et que le piston différentiel doit être alimenté par une pression partielle constante en état stationnaire, la soupape de commande est conçue pour que les sections transversales de l'écoulement au niveau de la soupape puissent être réglées en permanence ou non par l'excitation correspondante de l'électro-aimant.Hydraulic adjustment device having a differential cylinder (11) with differential piston (12), the pressure chamber (13) of which is permanently supplied by a pump (16) on the side of the small surface of the piston, the pressure chamber (18 ) on the surface side of the piston (18) being supplied by an electromagnetic control valve (20). This has a negative overlap in the neutral position. Since the pump is actuated at variable speed and the differential piston must be supplied by a constant partial pressure in the stationary state, the control valve is designed so that the cross sections of the flow at the valve can be adjusted permanently or not by the corresponding excitation of the electromagnet.

Description

Hydraulic actuator

State of the art

The invention is based on a hydraulic actuator according to the preamble of the main claim. In such a known actuator, for. B. according to DE-OS 36 16 234, the control unit consists of a pump, memory, pressure control valve and solenoid valve. Exact positioning of the differential piston is often difficult under certain conditions, and the expenditure on components is high and costly.

Advantages of the invention

The actuating device according to the invention with the characterizing features of the main claim has the advantage that it works with significantly lower losses if no adjustment is made, that it works particularly precisely and is nevertheless inexpensive. Further advantages of the invention emerge from the subclaims.

drawing

The invention and embodiments of the invention are explained in more detail in the following description and drawing. Show it:

Figure 1 is a simplified representation of a hydraulic actuator;

Figure 2 shows a longitudinal section through a control valve;

FIG. 3 shows an equivalent circuit diagram of the hydraulic actuating device;

4a and 4b are diagrams showing the flow cross sections as a function of the speed on the pump shaft;

FIG. 5 shows a simplified illustration of a modified embodiment of the hydraulic actuating device;

6 to 11 each show longitudinal sections through further embodiments of a control valve;

12 and 13 each show longitudinal sections through only partially shown further embodiments of a control valve.

Description of the embodiments

In FIG. 1, 10 denotes a hydraulic actuating cylinder which has a differential cylinder 11 with differential pistons 12, 14. The pressure chamber 13 on the small piston surface of the piston 14 is always acted upon by a pump 16 via a line 15, which is driven by a drive shaft 17, for example, the camshaft of an internal combustion engine.

A line 19, which starts from a control valve 20 and which continues to line 15 of the pump, opens into the pressure chamber 18 on the large piston surface. A pressure limiting valve 21 is also connected to this line. The control valve is designed as a 3/2-way valve and is shown in more detail in FIG. It is actuated by an electromagnet 22 against the force of a spring 23. The control valve has a slide 24 and is designed in a so-called open center design, i. H. in the neutral position of the control slide 24 there is a connection between an inlet bore 25 to which the line 19 is connected and a bore 29 to which the space 18 is connected via the line 19 and the bore 27 which is connected to the return. The control slide has a continuous, stepped longitudinal bore 30 running in its longitudinal direction. The bore 27 is also designated with R, the bore 29 with A and the bore 25 with P. The pressure relief valve 21 is set to 120 bar, for example.

In the de-energized state, the control valve has the position shown in FIG. 1. The pressure chamber 18 is depressurized via the line 19 and the control valve cross section A. The pressure chamber 13 is supplied with the pump delivery pressure via line 15. The differential piston moves to the left. In the constantly energized state, both pressure chambers 13, 18 are acted upon by the pump delivery pressure. The larger effective area in space 18 results in a total force which moves the differential piston to the right.

If the differential piston 12 is to assume a stationary position, this can be done by clocking the control valve 20. The operating pressure is predetermined by the pressure relief valve 21. If the pump 16 were to work continuously against 120 bar, the power consumption would be too high at high speed. This will now thereby prevents the control valve from being clocked during stationary operation in such a way that, on average over time, only a partial pressure of about 20 to 30 bar is established in rooms 13 and 18. At this pressure there is still no actuating movement of the differential piston, since a restoring force acts on the differential piston from the object to be actuated. FIG. 3 shows an equivalent circuit diagram in which P represents the inflow from the pump to A (consumer) or R (return). The inlet cross-section according to P and the outlet cross-section according to R are shown as adjustable throttles that are set by the control slide. The above-mentioned, resulting partial pressure results from the delivery flow of the pump 16 and the cross sections from P to A and A to R. If the delivery flow changes, for example, as a function of the speed, a different partial pressure would occur with a constant position of the control spool, which again, an undesirable piston movement would result. The time-averaged cross section from A to R must therefore be changed with increasing speed in accordance with the physical laws so that the partial pressure remains approximately constant. The time-averaged cross-section from P to A should be so large that the pressure drops as little as possible. Figures 4a and 4b show the flow cross-sections depending on the speed at the pump shaft for stationary operation. From this it can be seen that the outflow cross section A increases with the speed. The cross section A inevitably becomes smaller with increasing valve speed. The ratio A ", = 3 A-."(-> P ,,, - 9 P ",.

PA AR AR PA).

Deviations in the position of the differential piston, the z. B. resulting from temperature changes or flow forces in the spool are corrected via a target-actual comparison. This is done by correspondingly controlling the electromagnet 22 via an electronic control circuit, to which the corresponding characteristic values are supplied. The embodiment of the hydraulic actuating device according to FIG. 5 differs from that according to FIG. 1 by the use of a throttle and a spring-biased check valve in the return line from the control valve. A line 31, which is divided into two line sections 31 ', 31'', leads from the pressure connection R of the control valve 20 to a container 32 is inserted. The spring-biased check valve 34 acts as a pressure control valve and maintains a system pressure of 20 bar here in the central valve position. The throttle 33 ensures that in the event of a supply pressure failure and control valve in position I, the actuating cylinder is moved to the left by external forces in order to ensure an engine emergency operation even in the event of a supply pump failure.

FIG. 6 shows a further embodiment of the control valve, in which the valve body is designed as a seat valve body. The control valve 20A has a hollow cylindrical housing 35, on the one end side of which an inwardly pointing annular shoulder 36 is formed. An approximately cup-shaped magnet housing 37 with a reduced inner diameter, a piston guide disk 38 and a flange plate 39 are inserted into the housing one behind the other. The magnet housing 37 made of hard, non-magnetic material lies with its open end face against the ring shoulder 36, the flange plate 39 closes off the opposite end face of the housing. A central, conical bore 41 is arranged in the bottom 40 of the magnet housing and widens in the direction of the open side of the magnet housing. An approximately cup-shaped outer magnet insert 43, which consists of soft magnetic material, is also fitted into the interior section 37a of the magnet housing with a smaller diameter facing the bottom 40. The bottom 44 of the magnet insert has a conical extension 45 which projects into the bore 41 and is flush with the underside of the bottom 40 of the magnet housing 37. From the bottom inside of the outer magnetic insert 43 a likewise conical bore 47 ^• • is which merges into a continuous cylindrical bore 47b. A corresponding sealing cone 49 made of non-magnetic material is inserted into the bore 47a and is flush with the inside of the base 44. In the interior of the outer magnet insert 43, a magnet coil 22A with a coil body made of plastic is inserted. The magnet coil extends into the interior section 37b of the larger diameter of the magnet housing 37. Electrical leads 50 lead from the magnet coil out of the housing 35. The magnet coil 22A together with the coil body is fixed in its position by a cover 52 made of soft magnetic material inserted into the interior section 37b. A cylindrical extension 53 with a stepped outer diameter projects through the open end face of the housing 37 from the cover. Another extension 54 projects from the cover into the interior of the hollow cylindrical magnet coil 22A. The outside diameter of the extension 54 corresponds to the inside diameter of the coil body of the magnet coil 22A. The cover 52 lies flush with the end face of the housing against the end face of the outer magnet insert 43.

The sealing cone 49 has a central bore 56, the diameter of which corresponds to the inside diameter of the coil body of the magnet coil 22A. An approximately cup-shaped inner magnet insert 57 made of soft magnetic material projects through this bore 56, so that its bottom rests on the extension 54 of the cover 52. The open end face of the inner magnet insert 57 is flush with the outer side of the bottom 40 of the magnet housing 37 and the conical extension 45 of the outer magnet insert 43. An annular groove 58 is thus formed through the bore 47b in the extension 45 of the outer magnet insert and the sealing cone 49 and the inner magnet insert 57.

The inner magnet insert 57 and the sealing cone 49 as well as the sealing cone and the outer magnet insert 43 are fixed to one another connected, e.g. B. soldered. The cover 52 and the magnet housing 37 are also firmly connected to one another, for. B. glued, soldered or crimped. Through these connections, the assembly of the sealing cone, the outer and inner magnet insert and the magnetic coil 22A together with the coil body are fixed in their position. The extension 53 of the cover 52 or the end face of the housing 37 and the electrical feed lines 50 are extrusion-coated with plastic in the form of a plug.

Arranged in the piston guide disk 38 is a continuous, central bore 60, the diameter of which is larger than that of the bore 47b in the outer magnet insert 43. In each of the two end faces of the piston guide disk 38, a flat cylindrical depression 61, 62 is provided which extends around the bore 60 whose axes are offset parallel to that of the bore 60. The axis of the recess 61 facing the flange 39 lies above the axis of the bore 60, while that of the opposite recess 62 lies below. In addition to the bore 60, an axially parallel blind bore 64 runs in the piston guide disk and is connected to the depression 62.

In the bore 60, a piston 66 is guided in a tightly sliding manner. The piston 66 consists of an outer piston sleeve 66a made of non-magnetic, hard material and an inner cylinder 66b made of soft magnetic material, which are connected to one another by soldering, welding or pressing. The end face of the piston 66 facing the flange plate 39 has a conical depression 68, which extends from the outer edge, so that it is designed in the form of an annular cutting edge 67 serving as a sealing edge. An axial, circular depression 69 extends from this depression. In the opposite end face of the piston, which has a conical recess 75 extending from the outer circumference, there is an axially extending blind bore 70, from the bottom of which a continuous bore 71 extends, which connects the blind bore 70 with the recess 69. The recess 75 is formed such that an annular cutting edge 74, which acts as a sealing edge, also remains on the outer circumference of the piston. At the bottom of the blind bore 70, a compression spring 23A is supported, the opposite end of which rests against the bottom of the cup-shaped, inner magnet insert 57.

In the vicinity of the outer circumference of the piston 66, a plurality of symmetrically arranged continuous, axially parallel bores 73 run in it, which are connected to an annular groove 76 in the end face of the outer piston sleeve 66a facing the magnet housing. The outer diameter of the annular groove 76 is smaller than the outer diameter of the piston 66.

An axial blind bore 78 is arranged in the end face of the flange plate 39 facing the piston guide disk 38, the outside diameter of which is slightly smaller than that of the bore 60. A blind bore 79 which runs in the same axis and which is connected to the line 19 extends from the opposite end face. The connection of the line 19 to the blind bore corresponds to the control cross-section A. The two blind bores 78, 79 are connected to one another by a throttle bore 80. In addition to the blind bores 78, 79, a through bore 82 is made which is aligned with the bore 64 in the piston guide disk 38. The bore 82 corresponds to the control cross-section R on the outlet side.

In addition to the blind bores 78, 79, two axially identical blind bores 83, 84 run opposite the bore 82, each of which originate from the opposite end faces of the flange plate 39 and are connected to one another by a throttle bore 85. The blind bore 84 is connected to the recess 61 in the piston guide disk 38, while the bore 83 corresponds to the control cross section P. Between the bore 82 and the outer edge, the flange plate is penetrated by a blind bore 86, which likewise penetrates the piston guide disk 38 and extends into the magnet housing 37. A pin 87 is inserted into the blind bore 86, which secures the correct position of the flange plate 39, the piston guide disk 38 and the magnet housing 37. The outer circumference of the housing 37 is provided in the region of the piston guide disk 38 and the flange plate 39 with an external thread 88, which is used to connect the control valve to a suitable connection surface and to clamp the magnet housing 37, the piston guide disk 38 and the flange plate 39.

In the de-energized state, the control valve 20A has the position I as shown in FIGS. 1 and 5. Due to the action of the spring 23A, the piston 66 rests with its conical sealing edge (ring cutter 67) on the flange plate 39, so that the pressure medium connection P is closed on one side. From the pressure medium connection A, pressure medium passes through the blind bores 79 and 78 and the throttle bore 80 arranged between them into the recess 69 of the piston. From there, there is a connection via the bores 73 and the annular groove 76 past the annular cutting edge 74 to the depression 62, which is connected to the pressure medium connection R via the bores 64 and 82. At the same time, the recess 69 is connected via the bores 71 and 70 to the opposite end of the piston. Through the gap 89 between the piston 66 and the magnet housing 37 there is a connection to the annular groove 76. The differential piston 12, 14 of the differential cylinder 11 moves to the left.

In the energized state, the piston 66 acting as a magnet armature is pulled against the force of the spring 23A to the magnet housing and thus closes off the pressure medium connection R on one side. At the same time, the pressure medium connection P is via the bores 83, 85, 84 and the depression 61 past the ring cutter 67 with the depression 69 and the blind bore 78 connected. Pressure medium can thus reach the pressure medium connection A via the throttle bore 80 and the blind bore 79. At the same time, the gap 89 is pressurized via the bores 71 and 70 and the bores 73. As a result, the delivery pressure of the pump 16 is applied to both pressure chambers. The larger effective piston area in the pressure chamber 18 results in a force which moves the differential piston to the right.

The control valve 20A has a negative overlap, i. H. when the piston 66 is tightened, the sealing edge formed by the annular cutting edge 67 opens before the second sealing edge formed by the annular cutting edge 74 is closed. During the movement phase of the piston 66, all three pressure medium connections A, P, R are connected to one another. In addition, the pressure forces on the piston are completely balanced, i. H. the piston is free of axial forces in any position due to the pressure of the pressure medium.

A further exemplary embodiment of a control valve is shown in FIG. 7, in which larger magnetic forces and thus better switching time constancy are possible through a larger electromagnet. The valve has a modular structure so that differently dimensioned components can be combined with one another. The same parts are labeled with the same numbers. Parts with the same function are identified with the same numbers and additional capital letters.

The control valve 20B according to FIG. 7 has an approximately cup-shaped magnet housing 37B, from the bottom inside of which a cylindrical core 93 extends and which is flush with the open end face of the magnet housing. A magnet coil 22B is inserted into the annular space 94 between the outer wall of the magnet housing 37B and the core 93. In its free end face, the core 93 has an axial, cylindrical recess 96 into which an approximately cup-shaped insert 97 made of hard material is fitted. On the outer circumference of the magnet housing run in the region of the open end face two adjoining ring grooves 99, 100, of which the ring groove 100 starting from the end face has the smaller inside diameter. A ring 101 made of non-magnetic material is inserted into this annular groove, the outer diameter of which corresponds to that of the magnet housing 37B.

On this end face of the magnet housing, one end face bears a cylindrical piston guide disk 38B, on the opposite end face of which a cylindrical flange plate 39B bears. The piston guide disk 39B has a cylindrical recess 106, 107 in each of its two end faces. The diameter of the larger depression 106 facing the magnet housing is larger than the inside diameter of the annular groove 100. The two depressions are connected by a central bore 108. A piston 110 made of hard material is slidably guided in this bore. On its front side facing the magnet housing, the piston has an extension 111 which projects into the interior of the cup-shaped insert 97. This extension is supported by the compression spring 23B, which rests with its opposite end on the bottom of the insert 97 and presses the piston against the flange plate. Around the extension 111 there is an annular groove 112 in the end face of the piston 110, from the bottom of which a plurality of symmetrically arranged longitudinal bores 113 extend, which open into a cylindrical depression 109 on the opposite end face of the piston. The outer edges of the end faces of the piston are each recessed conically inwards from the outer circumference, so that annular cutting edges 114, 115 are formed, of which the annular cutting edge 115 faces the flange plate 39B. A ring 117 is inserted into the annular groove 112, the outer diameter of which corresponds to that of the annular groove. The end face of the ring 117 facing the insert 97 is beveled in the opposite direction to the ring cutting edge 114, so that there is a V-shaped cross section - shown very exaggerated. The opposite end face of the ring is in the area of the longitudinal bores 113 also bevelled to reduce the flow resistance when the pressure medium flows. A ring 118 corresponding to the ring 117 is inserted into the recess 109. The shape of the cutting edges 114, 115 and the rings 117, 118 serves to balance the flow forces on the end faces of the piston. The V-shaped cross section formed by the shape of the ring cutters and the rings can also be formed by appropriate machining of the piston end faces without the use of rings.

On the end of the piston 110 facing the magnet housing, a plate-shaped flat armature 119 is fastened with an axially arranged sleeve 120 such that the piston 110 penetrates the sleeve 120. As a result, a residual air gap remains between the flat armature 119 and the magnetic housing 37B or the magnetic coil 22B when the annular cutting edge 114 of the piston is applied to the edge of the insert 97. The outer diameter of the flat armature corresponds approximately to the inner diameter of the annular groove 100. A plurality of symmetrically arranged bores 122 penetrate the flat armature and serve for the pressure medium flow.

In the flange plate 39B there is a blind bore 123 of smaller diameter which extends the bore 108 in the piston guide disk 38B and from the bottom of which a bore 124 which corresponds to the control cross section A extends. The annular shoulder 125 formed by the transition of the bores 108 and 123 interacts with the annular cutting edge 115 as a sealing edge. A longitudinal bore 126 corresponding to the bore 82 or the control cross-section R extends from the free end face of the flange plate and continues in the piston guide disc and opens into the recess 106. Another longitudinal bore 127 corresponding to the pressure medium connection P through the flange plate opens into the recess 107. When the electromagnet is not energized, the spring 23B presses the piston 110 downward, and the annular cutting edge 115 of the piston 110 lies against the flange plate 39B, so that the bore 127 is closed on one side. At the same time, pressure medium can flow from the pressure chamber 18 through the bores 124 and 123 through the piston and the recess 106 and through the bore 126, so that the pressure chamber 18 is depressurized.

If the electromagnet 22B has current flowing through it, the piston 110 is pulled upward by the flat armature 119, and the ring cutting edge 114 of the piston lies against the insert 97, so that the bore 126 and the recess 106 are closed on one side. At the same time, the bore 127 and the bore 124 are connected to one another via the depression 107 and the blind bore 123.

The ring 101 on the magnet housing 37B largely prevents the magnetic flux from passing between the magnet housing and the piston guide disk 38B. This undesirable magnetic flux would be almost completely lost for the magnetic force build-up. Magnetic contamination could also be drawn into the piston guide between the piston and the piston guide disk.

The cross sections of the piston end faces formed by the rings 117, 118 and the shape of the ring cutting edges 114, 115 serve to balance the flow force and act mainly in the case of a flow from the outside inwards. With currents in the opposite direction, they are not as effective. In order to achieve a flow force compensation here, too, the shape of the flat anchor 119 and the opposite piston end face can be adapted accordingly.

An exemplary embodiment of a control valve modified in this way is shown in FIG. 8. The sleeve 120 of the flat armature 119 has an annular extension on the end facing the magnet housing 37B 128, the end of which is chamfered inwards. The sheathing of the electromagnet 22B has an annular groove 128 'cooperating with this extension on the end face pointing towards the extension 128, the outer edge of which is also beveled.

In the flange plate 39B, an annular groove 129 is made in the end facing the piston 110, the outer diameter of which corresponds to that of the opposite recess 107 in the piston guide disk 38B. The inner diameter of the annular groove 129 is larger than the diameter of the blind bore 123. A deflecting screen 129 ′ which cooperates with the annular groove 129 is fastened on the piston 110. With its end facing the flange plate 39B, which has a conical depression 129 ″, this protrudes into the annular groove 129. The deflecting screen 129 ′ is designed such that its outer circumference tapers on the side facing the flat anchor.

In the exemplary embodiment according to FIG. 9, the undesired magnetic stray flux is prevented by other measures. The magnet housing lacks the non-magnetic ring and the corresponding ring groove. Instead, the magnet housing and the piston guide disk have a plurality of annular grooves, which are described in more detail below and which hinder or guide the magnetic flux.

The magnet housing 37C has an annular groove 130, which extends from the end facing the piston guide disk 38C and is arranged between the coil 22C and the outer circumference of the magnet housing. Two further annular grooves 131, 132 extend from the recess 106C in the piston guide disk, of which the annular groove 131 is arranged on the outer circumference of the recess 106C and the annular groove 132 between the bores 126 and 108. A further annular groove 133 extends from the opposite end face, which is arranged in the region of the bore 126 and extends as far as between the annular grooves 131, 132, so that the annular grooves - seen perpendicular to the axial direction - overlap without being connected to one another. FIG. 10 shows another embodiment of a switching valve that differs from the one described above in that a second compression spring acts on the piston. This changes the spring force on the piston depending on its stroke.

For this purpose, the magnet housing 37D has a recess 96D in the core 93D which projects to the bottom of the magnet housing. The insert 97D has a corresponding extension 135 on its bottom underside. A blind bore 136 penetrates this extension from the bottom side of the magnet housing, from the bottom of which an axial bore 137 extends, which penetrates the bottom of the insert 97D and whose diameter is somewhat smaller than is the inner diameter of the spring 22D.

A plunger 138 protrudes through the bore 137 and rests with its plate-shaped end 139 on the bottom of the blind bore 136. The other end 140 of the plunger extends into the vicinity of the extension 111. The distance A between the end of the plunger and the extension 111 is smaller than the distance A between the sealing edge 114 and the magnet housing 39D. The end of the plate-shaped end 139 of the plunger 138 facing away from the base of the blind bore 136 is provided with a central, conical recess 142. A pressure spring 143 located in the blind bore 136 acts on the plunger 138 and is supported with its ends on a connecting plate 144, 145 in each case. The support plate 144 bears against the base of the recess 96D, the opposite support plate 145 interacts with the plate-shaped end of the plunger via a ball 146. For this purpose, the side of the end plate 145 facing the plunger also has a central, conical depression 147, so that the ball 146 is guided through the two opposite depressions 142, 147. The ball connection largely prevents the transfer of non-axial forces. Two cross bores 149, 150 extend through the insert 97D. The cross bore 149 penetrates the blind bore 136 in the vicinity of the base of the bore, while the cross bore 150 penetrates the walls of the insert 97D near the bottom.

The two transverse bores 149, 150 are connected to one another by an annular channel 151. This ring channel is formed by the wall of the recess 96D and an annular groove 152 on the outer circumference of the insert 97D, which is formed between the two transverse bores 149, 150. The blind bore 136 is connected to the depression 106 via the transverse bores 149, 150 and the ring channel 151 via the interior of the insert 97D.

The restoring force on the piston 110 and the flat armature 119 is in this. Embodiment generated by the two springs 22D and 143. The spring 22D acts over the entire stroke. After a partial stroke, which corresponds to the distance A, the extension 111 bears against the plunger 138 and the spring 143 also acts.

In this embodiment, the central position of the piston can be achieved in a stable manner when the electromagnet is partially excited. The partial excitation can be implemented with low losses by clocking the control voltage at a higher frequency. The valve according to FIG. 10 is particularly suitable for use in the hydraulic circuit according to FIG. 5. The cross sections from A to R and from A to P should be large enough that only a low pressure drops at them, and therefore in all three, at maximum speed Connections have approximately the same pressure.

The exemplary embodiment according to FIG. 11 differs from the ones described above by a modified piston, by means of which this control valve can be used as a pressure-maintaining valve. For this purpose, the piston 110E has a section 155 in the region of the depression 107 of smaller diameter, which starts from the end face of the piston and whose axial extent corresponds to the length of the recess 107. The opposite end of the piston has a shoulder 156 of larger diameter in the region of the recess 106, which starts from this end face of the piston and extends into the vicinity of the base of the recess 106 without reaching it.

In the de-energized state, the piston 110E abuts the flange plate 39E and closes the pressure connection P. By suitably coordinating the force of the spring 23D and the annular surface 157 on section 155, the piston is lifted off at a certain pressure (here about 120 bar) .

If the axial extension of the shoulder 156 in the direction of the flange plate 39E is reduced, the annular surface 158 can serve as a stop for an annular shoulder (not shown) of the sleeve 120 of the flat anchor, so that it is additionally fixed on the piston 110E.

When the electromagnet is fully excited, the piston 110E abuts the insert 97E, so that the pressure medium connection R is closed. If the force generated by the pressure on the end face of the piston enlarged by the annular surface 158 of the shoulder 156 exceeds a value which is greater than the difference between the magnetic force and the force by the spring 23D, the piston is moved downward so that it does not more is present and pressure medium can flow through the pressure port R.

As a result of this redesign, the control valve 20E can be used as a pressure-maintaining valve. When using this valve in the hydraulic actuating device according to FIGS. 1 and 5, the preloaded check valve 21 can be dispensed with.

In order to reduce or prevent contamination of the piston guide, the axial extension of the section 155. be shortened by an amount that corresponds to half the piston stroke. The resulting edge of the section 155 then slides past the mouth of the bore 108 into the depression 107 when the piston moves. The pressure medium flow in this area and the lifting effects at the edges remove and wash away any impurities that are already in the guide gap.

By suitable modification of the magnet housing, the various embodiments of the control valve can also be used with proportional magnets. For this purpose, the magnet housing in the area of the armature can be changed by additional or modified magnet guides. The shape of the magnet housing should then be chosen so that the magnetic flux across the axial working air gaps remains constant regardless of the stroke at constant current. This ensures that the magnetic force remains constant over the stroke. The use of a proportional magnet in the embodiment of the control valve as a pressure-maintaining valve is particularly advantageous.

A control valve modified in this way is shown in FIG. The coil of a proportional magnet 22F is inserted into the magnet housing 37F, the dimensions of which are chosen such that an annular space 160 remains on the open end face of the magnet housing facing the piston guide disk 38F. In this annulus three ring elements 161 to 163 are used, which serve to guide the magnetic flux. The outer ring element 163 bears against the inner circumference of the magnet housing 37F and projects into the recess 106F with a section 164 of smaller outer diameter. The inner ring element 161 lies against the outer circumference of the core 93F and extends into the recess 106F. In the end face of the ring element facing the recess, a circular recess 165 is provided, from the outer circumference of which an annular groove 166 extends. Several open into this ring groove radially extending, symmetrically arranged bores 167, which in turn emanate from the outer circumference of the ring element 161. Between the ring elements 161, 163, the ring element 162 is fitted, which is designed as a flat disc made of non-magnetic material. An annular channel 168 is formed by the three ring elements 161 to 163 and is open to the recess 106F. The bores 167 connect this annular channel 168 to the annular groove 166 and serve for the pressure medium flow.

A cup-shaped insert 170 made of non-magnetic material is fitted into the recess 96F in the core 93F. On the bottom inside of this insert, one end of the compression spring 23F is supported with an end plate 171 and a compensating disk 172, the opposite end of which rests on the piston 110E. The plate-shaped flat anchor 119F is fastened to the piston 110E in the region of the shoulder 156. The inside diameter of the flat anchor is somewhat larger than the outside diameter of the section 164 of the ring element 163, so that the edge 173 of the flat anchor extends partially around the ring element 163. The sleeve 120F of the flat anchor projects into the recess 165 of the inner ring element 161. The outer diameter of the sleeve 120F is slightly smaller on this side than the diameter of the recess 165.

The outer circumference of the piston guide disk 38F consists of a non-magnetic ring 174 in the region of the recess 166F, which reduces magnetic stray fluxes. This ring extends into a corresponding ring groove 175 in the magnet housing 37F.

When the proportional magnet 22F is actuated, the flat armature 119F approaches the magnet housing 37F. The radial magnetic flux, which does not generate a usable magnetic force, is increased via the ring elements 161 and 163. By appropriate adjustment of the dimensions of the ring elements and the flat armature, the magnetic flux that is the axial working air gap between the magnet housing 37F and the flat armature 119F flows, kept constant regardless of the stroke. This also keeps the magnetic force constant over the stroke.

The control valve 20F according to FIG. 12 is advantageously used in the hydraulic actuating device according to FIG. 1, the pressure control valve 21 being able to be omitted. The function of the throttle 33 and the pressure holding function by the prestressed retaining valve 34 in FIG. 5 are maintained by this control valve with a corresponding magnetic flow.

In the embodiments according to FIGS. 11 and 12, the end faces of the piston 110E can also be flat. In order to achieve a pressure maintenance function with the control valves 20E, 20F, the sealing diameter of the end face of the piston facing the magnet housing 37E, 37F is larger and the sealing diameter of the end face facing the flange plate is smaller than the diameter of the bore 108. The diameters can affect the desired pressure control behavior be adjusted. The sealing diameters are given by the diameter of the blind bore 123E or the blind bore (not shown) in the flange plate of the control valve 20F and by the inner diameter of the inserts 97E, 170. These diameters determine the narrowest flow cross-section, which largely determines the pressure build-up on the piston ends.

FIG. 13 shows a modification of the exemplary embodiment described above, which differs from it in that it has a modified magnetic flux guide. The control valve 20G has a magnet housing 37G, in which the coil of an electromagnet 22G is inserted. In contrast to the exemplary embodiment described above, the magnet housing 37G has no non-magnetic ring. The length of the coil of the electromagnet is selected so that a flat annular groove 178 remains on the end face facing the piston guide disk 38G. The piston guide disk 38G is composed of three piston guide rings 180-182 arranged axially one behind the other. The piston guide ring 180 facing the flange plate 39G has a sleeve-shaped extension 183 which extends in the direction of the magnet housing 37G without reaching it. The piston guide ring 182 arranged on the magnet housing also has a sleeve-shaped extension 184, the outer diameter of which is larger than that of the extension 183. Between the piston guide rings 180 and 182, the piston guide ring 181 made of non-magnetic material is arranged. Its inner diameter corresponds to the outer diameter of the extension 183. On the end facing the magnet housing 37G, the piston guide ring 181 has a cylindrical depression 185, the outer dimensions of which correspond to those of the extension 184 which projects into the depression. The piston guide ring 180 has on its end facing the flange plate 39G a cylindrical, off-center recess 107G which corresponds to the previously described recesses in the piston guide disks. The cylindrical interior 186 of the piston guide ring 180, which also extends through the sleeve-shaped extension 183, corresponds to the bore 108. In this interior, the piston 110E - already described above - is slidably guided and extends into the interior 187 of the piston ¬ guide ring 182. The piston 110E can be flat on its end faces - as described above - the sealing edges can also - as shown - be machined in the form of ring edges. The diameter of the cylindrical interior 187 of the piston guide ring 182 is slightly larger than the inside diameter of the coil of the electromagnet 22G. On its end facing the magnet housing 37G, the piston guide ring has a flat, cylindrical depression 188, the diameter of which corresponds to the outer diameter of the coil of the electromagnet 22G. An annular armature 190 is arranged on the piston 110E in the area of the interior 187 with little play to the wall thereof. The axial dimension of the anchor is chosen so that a flat annular space 191 remains between its end face facing the flange plate 39G and the end face of the extension 183 and the bottom of the depression 185. The armature 190 protrudes with its one end face into the depression 188 and there has a shallow recess 192 in the form of a truncated cone, so that axial flow forces on the piston are compensated for when the pressure medium flows.

On the outer circumference of the armature, three annular grooves 194-196 are formed, the spacing of which corresponds approximately to the maximum stroke of the armature. The groove width and the groove depth correspond approximately to twice the groove spacing. Three ring grooves 194 '- 196' are also arranged in the wall of the interior 187, the dimensions of which correspond to those of the ring grooves 194 - 196. The ring grooves 194 - 196 and 194 '- 196' are flush with each other when the piston 110E abuts the flange plate 39G.

Through the piston guide plate 38G, i.e. H. A bore 198 runs through the piston guide rings 180-182, which corresponds to the bore 126 and connects the depression 188 and the annular space 178 with the bore 126 'in the flange plate.

The shape of the armature 190 and the piston guide ring 182 influences the course of the magnetic force as a function of the armature stroke. When the piston bears against the flange plate 39G (rest position), the radial magnetic flux cross section on the outer circumference of the armature is very much larger than the magnetic flux cross section on the working air gap. When the armature or the piston approaches the magnet housing, the magnetic voltage at the working air gap between the magnet housing and armature is reduced by the decreasing covering length between armature and piston guide ring 182. The magnetic force characteristics Depending on the piston stroke and excitation current through the coil of the electromagnet, the number and size of the slots, the outside diameter of the armature and the residual air gap between the armature and the magnet housing can be significantly influenced. Constant magnetic forces or constant increases in magnetic force can thus be achieved when the armature approaches the magnet housing. The increase in magnetic force should preferably correspond approximately to the increase in spring force over the piston stroke.

The design of the electromagnet and the compression spring is chosen, for example, so that the spring force and the magnetic force are balanced when a current flows through the electromagnet which corresponds to approximately 40% of the nominal current. This means that at higher or lower currents there is a constant excess of the magnetic force or the force exerted by the spring on the piston regardless of the stroke, which is advantageous for the control or regulation of the pressures. The non-magnetic piston guide ring 181 prevents unwanted magnetic flux from entering the piston.

When the piston moves, there is a change in volume in the annular space 191. Since pressure medium can only be displaced via the gaps between the piston and the piston guide ring 180 and its extension 183 and between the armature and the piston guide ring 182, there is very good hydraulic damping of the piston . Short-term pressure differences on the piston end faces therefore do not have a major effect on the piston.

Due to the oscillating movement of the piston, pressure medium is pumped back and forth from the annular space 188 into the annular space 191. This can contaminate the gap between the armature and piston guide ring 184. Via the piston guide gap between the piston and the piston guide ring 180 together with the extension 183, a cleaning fluid flow, which is filtered through the narrow gap, flows through the annular space 191 through the gap between the armature and the piston guide ring 184. If an RC element is connected in parallel to the coil of the electromagnet in the electrical / electronic circuit for controlling the electromagnet, a decaying, alternating alternating current acts on it when it is switched off, which demagnetizes all components carrying magnetic flux. As a result, magnetic impurities are no longer attracted and can be flushed out of the valve by the pressure medium.

In both end positions of the piston 110E it is possible to set the operating pressure and thus the adjustment speed of the differential piston 12 by correspondingly controlling the proportional magnet. If the maximum permissible adjustment speed is included in the control concept of the control electronics, tolerance-related variations in the operating pressure or the flow cross sections can be compensated. This has the advantage that the component tolerances can be larger. The manufacturing costs can thus be reduced considerably.

A further advantage is the possibility of adapting the operating pressure to the speed of rotation. B. used to adjust the camshaft of a motor vehicle relative to its crankshaft, the drive torque of the camshaft decreases with increasing speed and thus the required actuating force. By including the maximum permissible adjustment speed, a pressure adjustment or a reduction in the power consumption of the actuating device can be achieved automatically.

In such a control concept, e.g. For example, in addition to the control variable control piston travel S, the difference between the maximum adjustment speed and the actual adjustment speed can be included as a comparison variable. Another option for developing a control concept is to include the control parameters in a map. Control variables or control parameters can e.g. B. be: shaft speed, travel, Stell¬ speed, oil temperature. The control parameters can thus be adaptively changed, for example, when the maximum permissible adjustment speed is exceeded.

The rest position of the differential piston can also be maintained by comparing the target displacement with respect to the actual displacement by adjusting the corresponding pressure level. The design of the control valve as a pressure-maintaining valve enables the pressure required for this to be maintained, for example. B. 20 bar. The control variables adjustment path and adjustment speed can be recorded either directly via displacement sensors on the differential piston or indirectly via angle sensors on the crankshaft and camshaft.

Claims

Expectations

1. Hydraulic actuating device with a differential piston (12), the annular surface of which is always acted upon by pressure medium, while the pressure in the pressure chamber (18) on its large piston surface can be changed by an electromagnetically actuated control valve (20; 20A - 20G), the valve body of which (24; 66; 110; 110E) has a negative overlap in the middle position, a partial pressure being established in the pressure chambers (18 and 13) due to the partial flow of pressure medium, characterized in that the flow cross section of the control valve (20th ; 22A - 22G) can be changed continuously or discontinuously depending on certain criteria by appropriate control of the electromagnet (22; 22A - 22G) in such a way that the partial pressure in the pressure chambers (18 and 13) remains approximately constant.

2. Device according to claim 1, characterized in that the flow cross section of the control valve (20; 20A - 20G) from the consumer connection (A) to the return connection (R) is increased with the pump speed.

3. Device according to claim 1 and 2, characterized in that the electromagnet is controlled by electronics which is activated in the event of deviations from the target and actual value of the position of the differential piston.

4. Device according to claim 1 to 3, characterized in that. the valve body is a control slide (24).

5. Device according to claim 1 to 3, characterized in that the valve body is a Sitzventilk body (66, 110, 110E).

6. Device according to claim 1 to 5, characterized in that the pump (16) is driven by an internal combustion engine, in particular a vehicle, and the consumer is a device actuated by the differential piston for continuous adjustment of the camshaft relative to the crankshaft.

7. Device according to claim 1 to 6, characterized in that the control valve has a housing (35), in the one behind the other a magnetic housing (37 - 37G), a piston guide plate (38 - 38G) and a flange plate (39 - 39G) are used .

8. Device according to claim 1 to 7, characterized in that the seat valve body is designed as a piston (66, 110, 110E), de in a bore (60; 108; 186) is guided in the piston guide disc and that the piston pressure medium channels (70 , 71, 73; 113)

9. Device according to claim 1 to 8, characterized in that the piston (66; 110; 110E) cooperates with two valve seats.

10. The device according to claim 1 to 9, characterized in that the end faces of the piston cooperating with the valve seats are designed as ring cutting (67, 74; 114, 115).

11. The device according to claim 1 to 10, characterized in that the flow forces on the piston are reduced by compensating elements (117, 118, 128, 129 ', 192).

12. The device according to claim 1 to 11, characterized in that the piston simultaneously acts as an armature of the electromagnet and that the compressive forces on the piston are balanced.

13. The device according to claim 1 to 11, characterized in that a plate-shaped flat anchor (119; 119F) is attached to the piston.

14. Device according to claim 1 to 11, characterized in that an annular armature (190) is attached to the piston.

15. The device according to claim 1 to 14, characterized in that the piston guide disc consists of at least two piston guide rings (180 - 182).

16. The device according to claim 1 to 15, characterized in that on the outer circumference of the armature (190) and in the piston guide disc opposite annular grooves (194 - 196; 194 '- 196') are ausgebil¬ det.

17. The device according to claim 1 to 16, characterized in that a second compression spring (143) acts on the piston, which acts on the piston after exceeding a predetermined actuation path in addition to the first spring (23A - 23G).

18. Device according to claim 1 to 17, characterized in that the end faces of the piston have different diameters.

19. The device according to claim 1 to 18, characterized in that the electromagnet is a proportional magnet (22F, 22G).

20. Device according to claim 1 to 19, characterized in that the increase in magnetic force over the piston stroke corresponds to the increase in force by the spring over the piston stroke.

21. The device according to claim 1 to 20, characterized in that the electronics regulating the electromagnet includes the difference between the adjustment speed of the differential piston and a maximum allowable adjustment speed as a control parameter.

22. Device according to claim 1 to 21, characterized in that the electronics regulating the electromagnet accesses control parameters which have the shape of a characteristic diagram.

23. The device according to claim 1 to 22, characterized in that in the electronics regulating the electromagnet, an RC element is connected in parallel to the coil of the electromagnet.