EP0653014B1 - Dispositif electrohydraulique de positionnement - Google Patents

Dispositif electrohydraulique de positionnement Download PDF

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Publication number
EP0653014B1
EP0653014B1 EP93914613A EP93914613A EP0653014B1 EP 0653014 B1 EP0653014 B1 EP 0653014B1 EP 93914613 A EP93914613 A EP 93914613A EP 93914613 A EP93914613 A EP 93914613A EP 0653014 B1 EP0653014 B1 EP 0653014B1
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EP
European Patent Office
Prior art keywords
pressure
valve
line
camshaft
differential
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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.)
Expired - Lifetime
Application number
EP93914613A
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German (de)
English (en)
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EP0653014A1 (fr
Inventor
Manfred Ruoff
Helmut Rembold
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Robert Bosch GmbH
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Robert Bosch GmbH
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Publication of EP0653014A1 publication Critical patent/EP0653014A1/fr
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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01L—CYCLICALLY OPERATING VALVES FOR MACHINES OR ENGINES
    • F01L1/00—Valve-gear or valve arrangements, e.g. lift-valve gear
    • F01L1/34—Valve-gear or valve arrangements, e.g. lift-valve gear characterised by the provision of means for changing the timing of the valves without changing the duration of opening and without affecting the magnitude of the valve lift

Definitions

  • the invention is based on an electrohydraulic actuating device for actuating a device for adjusting at least one camshaft of an internal combustion engine relative to its crankshaft in accordance with the preamble of the main claim.
  • the pressure chambers of the differential cylinder serving as an actuator are acted upon by a high-pressure pump.
  • the larger of the two pressure chambers of the differential cylinder is acted upon directly by the pump pressure, the pressure in the smaller of the two pressure chambers can be changed via a control valve.
  • the pressure in the two pressure chambers can be changed by correspondingly controlling the control valve, holding pressures being set when the differential cylinder or differential piston is stationary, which are very much lower than the pressures necessary for adjustment.
  • valve control When using such an electrohydraulic actuating device for valve control of an internal combustion engine with stepless adjustment of the intake camshaft, the full load curve of the internal combustion engine can be significantly improved by an optimized valve closure.
  • This valve control requires a quick adjustment of the intake camshaft relative to the crankshaft over a crankshaft angle of approximately 50 °. High pressure pumps with maximum pressures in the range of approximately 100 bar are required for such a rapid adjustment.
  • a valve control with such an electrohydraulic actuating device reaches its limits.
  • a camshaft adjusting device for internal combustion engines in which a camshaft for the intake valve and a camshaft for the exhaust valve can be adjusted independently of one another by means of hydraulic actuators relative to the crankshaft, but this adjusting device has the Disadvantage that the two hydraulic adjustment devices work independently of one another and thus have two structurally complex, separate hydraulic circuits.
  • a hydraulic actuating device is known from international patent application WO-A 91 10 813, which has a differential piston, the smaller piston surface of which can be acted upon by a pump with pressure medium, while the pressure in the pressure chamber on its larger piston surface can be changed by an electromagnetically actuated control valve.
  • this adjusting device which can optionally be used as a camshaft adjusting device for internal combustion engines, has the disadvantage that for each single camshaft a single hydraulic actuating device must be provided, which causes high construction costs and thus high costs.
  • the camshaft adjusting device with the characterizing features of the main claim has the advantage that valve control of an internal combustion engine is possible with the aid of which a further substantial improvement in the exhaust gas quality of the internal combustion engine is possible. Due to the adjustability of the second camshaft (exhaust camshaft) of the internal combustion engine, the valve closure of the exhaust valves can also be optimized with regard to the speed and the load behavior of the internal combustion engine.
  • the features described in the subclaims allow both pure black-and-white adjustments (two control positions) of the exhaust camshaft and a continuous adjustment of the exhaust camshaft relative to the crankshaft.
  • the hydraulic components required for this are relatively simple, the additional components to be used are small.
  • both camshafts can be adjusted in a structurally simple manner by means of a common hydraulic circuit.
  • FIG. 1 shows an electrohydraulic actuating device with a differential cylinder 10, which does not form part of the present invention and whose pressure chamber 11 is separated from an annular chamber 14 by a differential piston 12 with a piston rod 13.
  • the annular space 14 is always acted upon via a pressure line 15 by a pump 16 which is driven by a drive shaft 17, for example the camshaft of an internal combustion engine.
  • the pressure chamber 11 on the larger, effective piston surface is connected via a line 19 to the inlet 20 of an active pressure control valve 21.
  • a valve seat 22 is formed which interacts with a valve member 23, which is designed here as a ball.
  • the tappet 24 of a proportional magnet 25, which projects into the pressure control valve 21, bears on the side of the valve member 23 opposite the valve seat 22.
  • the proportional magnet 25 is connected to a control unit 27 via an electrical control line 26.
  • the tappet 24 is brought into contact with the valve member 23 by a spring (not shown) which is only slightly pretensioned.
  • the magnetic force acting on the tappet 24 by appropriate energization acts in the closing direction of the valve member 23.
  • the outlet 28 of the pressure control valve 21 is connected to a container 29.
  • the line 15 between the pump 16 and the annular space 14 and the line 19 between the pressure space 11 and the pressure control valve 21 are connected to a connecting line 31 in which an overflow valve 32 is arranged. This is arranged so that its inlet 33 is acted upon by line 15, its outlet 34 leads via connecting line 31 to line 19. At the inlet 33 of overflow valve 32, a valve seat 35 is formed which interacts with a valve member 36. This is acted upon by a compression spring 37 in the closing direction. The bias of the compression spring 37 is adjustable - in a manner not shown.
  • a branch line 39 branches off from the line 15 and is connected to the annular space 40 of a second differential cylinder 41 with differential pistons 42 and piston rod 43.
  • the pressure chamber 44 on the larger effective piston side is connected to a 3/2-way valve 46 via a pressure line 45. From this a line connection 47 leads to the connecting line 39.
  • a third connection 48 of the 3/2-way valve 46 is connected to the container 29.
  • the valve member 50 of the 3/2-valve 46 is brought from the neutral position I into the switching position II by a switching magnet 51 against the action of a compression spring 52.
  • the switching magnet 51 is connected to the control unit 27 via an electrical control line 53.
  • the electrohydraulic actuating device is used to actuate a device for adjusting a camshaft of an internal combustion engine relative to its crankshaft. This adjustment results in a phase shift of the control curves of these shafts.
  • the camshaft for intake valve control is rotated relative to the crankshaft in a manner known per se. This requires rapid adjustment over a relatively large angular range (approximately 50 ° crankshaft angle).
  • This device for adjusting the camshaft is actuated by the differential cylinder or differential piston as an actuator.
  • a second camshaft of the internal combustion engine - namely that for exhaust valve control - can also be rotated relative to the crankshaft.
  • the angular range to be covered can be smaller than that of the first camshaft.
  • the second differential cylinder 41 or differential piston 42 is used for the relative rotation of the second camshaft (exhaust camshaft).
  • the first differential cylinder 10 is in its left end position.
  • the intake camshaft, not shown, would thus be adjusted to "late”, i.e. to a late turning position or valve actuation (relative to the crankshaft).
  • the second differential cylinder 41 is also in its left end position.
  • the exhaust camshaft, not shown, would be thus also in a "late" rotational position.
  • the 3/2-way valve 46 is in its spring-loaded neutral position I, so that the pressure chamber 44 of the second differential cylinder 41 is relieved via the pressure line 45 and the 3/2-way valve 46 to the third connection 48 or to the container 29.
  • the line connection 47 is closed at the same time, so that the annular space 40 is connected to the pump 16 via the connecting line 39 and the pressure line 15.
  • the second differential cylinder 41 thus remains in its left end position or is brought into this.
  • the annular space 14 of the first differential cylinder 10 is also connected to the pump 16 via the pressure line 15. Via the connecting line 31, the overflow valve 32 and the line 19 there is also a connection from the pressure line 15 to the pressure chamber 11 of the differential cylinder 10 and to the pressure control valve 21.
  • the pressure control valve 21 serves as an active control element for the movement of the first differential cylinder 10, in which the pressure of the pressure chamber 11 is influenced via controlled relief via the valve seat 22 of the pressure relief valve 21.
  • valve ball 23 In the non-energized switching position of the pressure control valve 21 shown, the valve ball 23 is brought into contact with the valve seat 22 by the tappet 24 of the proportional magnet.
  • the force on the valve member 23 in the closing direction is, however, small due to the only slightly preloaded compression spring (not shown) of the proportional magnet 25. Therefore, the valve member 23 lifts off the valve seat 22 even at low pressure at the inlet 20.
  • a predetermined, low back pressure is exceeded, therefore opens the pressure relief valve 21, so that the pressure chamber 11 of the differential cylinder 10 is relieved.
  • the annular space 14 is simultaneously subjected to pressure, as a result of which the differential piston 12 together with the piston rod 13 is moved to the left or is held in the left end position.
  • the overflow valve 32 or its compression spring 37 In order to be able to build up the pressure in the annular space 14 required for a movement of the differential piston 12 to the left, the overflow valve 32 or its compression spring 37 must be biased. By setting the compression spring 37 accordingly, the opening pressure of the overflow valve 32 is set to 30 bar, for example.
  • the opening cross-section of the overflow valve 32 which becomes free when this opening pressure is exceeded is dimensioned such that no appreciable throttle loss occurs when it flows through.
  • the bias of the compression spring (not shown) is increased by appropriate actuation of the proportional magnet 25, so that the opening pressure at the inlet 20 increases. Due to the higher opening pressure of the active pressure control valve 21, the pressure in also increases Pressure chamber 11 of the differential cylinder 10.
  • the opening pressure of the active pressure control valve 21 is correspondingly high, the overflow valve 32 opens completely, the released cross section being dimensioned in such a way that no appreciable throttle loss occurs. Due to the larger effective piston area on the pressure chamber 11, the differential piston 12 is adjusted to the right.
  • the pressure on the active pressure control valve 21 is set via appropriate excitation (lower current, control via control unit 27) of the proportional magnet 25 so that the resulting force on the differential piston 12 is just that due to the pressures in the two pressure chambers Restoring force from the device for adjusting the intake camshaft corresponds.
  • This restoring force or the restoring torque essentially results from the reaction forces when the camshaft is actuated by the camshaft and acts against the direction of rotation of the camshaft, that is to say in the direction of a late rotational position.
  • Appropriate control (control unit 27) of the proportional magnet 25 also ensures that these holding pressures are maintained at a level which changes just enough to absorb the restoring forces from the device for adjusting the intake camshaft even when the rotational speed of the camshaft changes.
  • the holding pressures required for this are significantly lower than the adjustment pressures required for (quick) adjustment of the differential piston or the intake camshaft.
  • the annular space 40 of the second differential cylinder 41 is always subjected to the pressure upstream of the overflow valve 32 via the connecting line 39.
  • the second differential cylinder 41 or differential piston 42 is brought or held in its left (late) end position.
  • the adjustment range of the exhaust camshaft required to optimize the exhaust gas quality is significantly smaller than that Intake camshaft. Therefore, the adjustment speed of the exhaust camshaft can be lower with the same adjustment time of the two camshafts.
  • the 3/2-way valve 46 is switched into its switching position II by actuating the electromagnet 51.
  • the pressure chamber 44 is thus also acted upon by the pressure in the connecting line 39 or the pressure upstream of the overflow valve 32. Due to the larger effective piston area, the differential piston 42 is moved into its right end position.
  • the pressures arising at the second differential cylinder 41 correspond at least to the opening pressure of the overflow valve 32 (30 bar) and are sufficient for the required adjustment speed. Should a higher pressure be present upstream of the overflow valve 32 due to appropriate control of the pressure control valve 21, the differential piston 42 or the exhaust camshaft is adjusted faster, which is not disadvantageous for the engine function.
  • a pure black and white adjustment of the exhaust camshaft is achieved with the electrohydraulic actuating device shown in FIG.
  • the exhaust camshaft is adjusted to its "late” end position in the neutral position I of the 3/2-way valve; when the 3/2-way valve 46 (switching position II) is switched, the exhaust camshaft is adjusted to its "early” final rotational position.
  • the pressure chamber 44 is thus relieved to the container 29, so that the pressure building up in the annular space 40 shifts the differential piston 42 to the left ("late" rotational position of the exhaust camshaft). If the hydraulic supply fails, the differential pistons 12 and 42 are moved to the left ("late") due to the mechanical restoring forces from the device for adjusting the camshaft. In both cases, an engine emergency running is ensured due to this reset to the "late" rotational position of the camshafts.
  • a 2/2-way valve can also be used as the active control element, which valve is actuated in clocked form by an electromagnet.
  • the pressure control for the first differential cylinder 10 then takes place via the control or regulation of the volume flow.
  • FIGS. 2 to 4 are suitable for a constant adjustment of the exhaust camshaft independently of the continuous adjustment of the intake camshaft or the first differential cylinder.
  • FIG Pump 16 applies.
  • the connecting line 39 branches off from the pressure line 15 and leads to the annular space 40 of the second differential cylinder 41.
  • a line 19 leads from the pressure chamber 11 of the first differential cylinder 10 to the pressure control valve 21, the structure of which corresponds to that described in FIG.
  • the pressure chamber 44 of the second differential cylinder 41 is connected via the pressure line 45 to a second active pressure control valve 60, the structure of which corresponds to that of the pressure control valve 21.
  • a first connecting line 61 branches off from line 19 and leads to a first output 62 of a current divider 63.
  • a second connecting line 65 leads from the second output 64 thereof to the pressure line 45.
  • a connecting line 67 extends from the input 66 of the flow divider 63 and is connected to the pressure line 15 or the connecting line 39 and thus the pump 16.
  • the overflow valve 32 is inserted, the structure and mode of operation of which corresponds to that described in Figure 1.
  • the overflow valve 32 is used so that it allows a pressure medium flow from the pump 16 to the flow divider 63.
  • a first throttle line 69 which is connected to the connecting line 67, branches off from the first connecting line 61, specifically between the overflow valve 32 and the flow divider 63.
  • a first throttle 70 is inserted into the first throttle line 69.
  • a second throttle line 71 with a second throttle 72 leads from the connecting line 67 (between the overflow valve 32 and the flow divider 63) to the second connecting line 65.
  • the flow divider 63 has an approximately cylindrical housing 74 in which a valve needle 75 is guided.
  • This valve needle 75 has in her In the middle a circumferential annular groove 76, so that a guide collar 77 and 78 respectively remain in the area of their two end faces.
  • Each of these guide collars 77 and 78 has throttle grooves 79 on its outer circumference, which connect the respective end face of the valve needle 75 to the annular groove 76.
  • the cross section of these throttle grooves 79 in the region of the guide collars 77 and 78 is designed such that a defined throttling effect is achieved.
  • the pressure drop across these throttle grooves here is approximately 3 bar in the exemplary embodiment.
  • the valve needle 75 is centered in the housing 74 by two compression springs 80 and 81, respectively.
  • the first spring 80 lies in the area of the first outlet 62 on the inside of the housing 74 and, on the other hand, is supported on the guide collar 79 of the valve needle.
  • the second compression spring 81 rests on the second guide collar 78 of the valve needle 75 and, on the other hand, is supported in the region of the second outlet 64 in the interior of the housing 74.
  • the first outlet 62 and the second outlet 64 are each designed as a valve seat which interacts with the guide collar 77 and 78 and at the same time serves as a stop.
  • the inlet 66 of the flow divider is arranged approximately in the middle of the housing 64 and opens in the region of the annular groove 76 of the valve needle 75.
  • valve needle 75 of the flow divider 63 When the control valves (pressure control valve 21 and second pressure control valve 60) are not activated, the valve needle 75 of the flow divider 63 is in its spring-centered central position.
  • the pressure built up by the pump 16 builds up via the overflow valve 32 and the connecting line 67 at the inlet 64 of the flow divider.
  • the volume flow generated by the pump is divided symmetrically to both outlets 62 and 64. Due to the non-activated pressure control valves 21 and 60, there can be in the connecting lines 61 or 65, however, do not build up any significant pressure.
  • the pressure in the line 19 or the pressure line 45 and thus in the pressure chamber 11 of the differential cylinder 10 or in the pressure chamber 44 of the second differential cylinder 41 is likewise correspondingly low.
  • the pressure which builds up in the annular spaces 14 and 40 due to the action of the overflow valve 32 Via the pressure line 15 or connecting line 39, the differential pistons of the differential cylinders move into their "late" end position or hold them in this.
  • the flow rate through the first connecting line 61 is reduced first. Because of the volume flow thereby increasing in the second connecting line 65, there is a greater pressure drop at Throttling grooves 79 of the right guide collar 78. The valve needle is deflected to the right due to this greater pressure drop on the right side until the flow rate at the second outlet 64 is throttled. Because of this throttling effect at the second outlet 64, a pressure can build up in the first connecting line 61 or at the first outlet 62, which presses the valve needle 75 against the right stop (second outlet 64) on its left side by pressurization.
  • valve needle 75 or the needle end faces, the cross section of the throttle grooves 79, the valve seat cross section of the outputs 62 and 64 and the bias of the centering springs 80 and 81 must be designed so that this switching state of the current divider 63 always occurs when only one of the two pressure control valves 21 and 60 is activated.
  • the restoring force resulting from the pressure drop at the throttle cross section together with the force of the compression springs must be smaller than the forces acting on the respective end face due to the pressure building up.
  • the second pressure control valve 60 is additionally activated, a pressure is also established there as a result of the bypass flow via the throttle 72, which moves the valve needle 75 back towards the left in the direction of the central position depending on the pressure level.
  • the throttle cross sections of the throttle 70 or 72 are selected so that the main volume flow takes place via the flow divider 63.
  • the valve needle 75 adjusts itself so that the larger volume cross section is always released to the pressure control valve with a higher opening pressure.
  • the pressure build-up and the volume distribution occur analogously when the second pressure control valve 60 is activated first.
  • the pressure profiles in the two pressure circuits can vary. influence each other due to the common overflow valve 32.
  • a pressure increase in one of the two pressure circuits (input pressure increase on one of the two pressure control valves) for adjusting one of the two differential cylinders can also act on the other circuit, in which, for example, an intermediate position of the differential cylinder has just been assumed. The resulting deviation or repercussions must be compensated for by the control unit. To avoid this, a further separation of the two pressure circuits can be carried out, as described in FIG. 3.
  • the electrohydraulic actuating device described in FIG. 3 differs from that described in FIG. 2 in that the overflow valve 32 is missing in the connecting line 67. Instead, each of the pressure control valves 21 and 60 and thus each of the two differential cylinders 10 and 41 is assigned its own overflow valve.
  • a first overflow valve 85 is arranged in the first connecting line 61, specifically between the mouth into the line 19 and the connection to the first throttle line 69.
  • a second overflow valve 86 is inserted between the second throttle line 71 and the pressure line 45 in the second connecting line 65 .
  • the structure and mode of operation of both overflow valves 85 and 86 correspond to those in FIGS. 1 and 2.
  • Both overflow valves 85 and 86 are used in such a way that they enable a pressure medium flow from the flow divider to the pressure control valve 21 or 60. Through these two overflow valves 85 and 86, the two pressure circuits (control of the first differential cylinder 10 and the second differential cylinder 41) are better decoupled, so that there is less pressure increase in one of the two circuits (activation of one of the two pressure control valves 21 or 60) strongly affects the other differential cylinder.
  • the exemplary embodiment of the electrohydraulic actuating device shown in FIG. 4 differs from the two previously described primarily by the design of the current divider.
  • the flow divider 85 has a cylindrical housing 86 in which a piston 87 is guided. This has a shoulder 88 or 89 of smaller diameter on its two end faces.
  • One end of a compression spring 90 is supported on the end faces of the piston 87 or 91, each of which comprises the corresponding paragraph 88 or 89.
  • the other end of the compression spring 90 or 91 abuts the respectively adjacent end of the housing 86.
  • a first outlet 92 is arranged in these end faces and a second outlet 93 is arranged on the opposite (right) side.
  • Two outputs 92 and 93 are designed as valve seats, and at the same time serve as a stop for the piston or the shoulders 88 and 89.
  • Two radial inputs 94 and 95 continue to open into the housing 86 of the flow divider 85, the first of which Entrance 94 is arranged in the area of the paragraph 88 and the second entrance 95 in the area of the paragraph 89.
  • Two further outputs are arranged approximately radially opposite one another, of which the third output 96 lies opposite the first input 94.
  • the fourth output 97 is opposite the second input 95.
  • the two inputs 94 and 95 are connected to one another by an input line 98, which in turn is connected to a pump 100 via a pressure line 99.
  • the first outlet 92 is connected to the annular space 14 of the first differential cylinder 10 via a first pressure line 101.
  • the second outlet 93 is connected to the annular space 40 of the second differential cylinder 41 via a second pressure line 102.
  • the pressure chamber 11 of the first differential cylinder 10 is connected to the pressure control valve 21 via the line 19.
  • the pressure chamber 44 of the second differential cylinder 41 is connected to the second pressure control valve 60 via the pressure line 45.
  • a first line connection 103 in which an aperture 104 is arranged, leads to the line 19 and is connected to it.
  • a second line connection 105 with an orifice 106 leads from the fourth outlet 97 to the pressure line 45 and is likewise connected to the latter.
  • a first branch line 107 emerges from the first connecting line 103, which on the other hand is connected to the first pressure line 101.
  • a first spring-loaded check valve 108 is arranged in this first branch line 107 and enables a pressure medium flow from the first line connection 103 to the first pressure line 101.
  • the second line connection 105 is connected to the second pressure line 102 via a second branch line 109.
  • a corresponding second check valve 110 is also arranged in this second branch line 109, which enables a pressure medium flow from the second line connection 105 to the second pressure line 102.
  • Another branch leads from the first pressure line 101 to the first line connection 103 and opens between the line 19 and the junction of the first branch line 107.
  • the overflow valve 82 is arranged, the inlet 33 of which faces the first pressure line 101.
  • the second pressure line 102 is connected to the second line connection 105 by a second branch 112.
  • the second overflow valve 83 is accordingly arranged in this second branch 112.
  • the first line connection 103 can also branch off from the pressure line 101 in the form of a line connection 103A (dashed illustration), the third outlet 96 of the flow divider 85 and the first branch line 107 with the first check valve 108 then being omitted.
  • a second branch line 105A can originate from the pressure line 102, the fourth outlet 97 of the flow divider and the second branch line 109 with the second check valve 110 then being omitted.
  • the two differential cylinders 10 and 41 are in their respective "late" end position, the flow divider 85 is in its spring-centered central position.
  • the annular space 14 of the first differential cylinder 10 is via the first pressure line 101, the second outlet 92, the interior of the flow divider 85 and the first inlet 94, the inlet line 98 and the pressure line 99 are acted upon by the pump 100.
  • the annular space 40 of the second differential cylinder 41 is acted upon by the pump 100 via the second pressure line 102, the second outlet 93, the interior of the flow divider 85, the first inlet 89 and the inlet line 98 and the pressure line 99.
  • the two pressure control valves 21 and 60 are not activated, so that - as described in the exemplary embodiments above - the pressure cylinders 11 and 44 of the differential cylinders 10 and 41 cannot build up a higher pressure.
  • the overflow valves 82 and 83 and the first orifice plate 104 and the second orifice plate 106 sufficient pressure can build up in the annular spaces 14 and 40. If, for example, the first pressure control valve 21 is now activated such that the opening pressure previously set at the overflow valve 82 is exceeded, the piston 87 of the flow divider 85 is moved to the right due to the pressure building up at the first outlet 92.
  • the piston 87 moves to the right until the second outlet 93 or the corresponding valve seat is closed by the shoulder 89 of the piston 87.
  • the main volume flow is thus available for the control of the first differential cylinder 10 which, analogously to the previously described exemplary embodiments, is moved in the respective actuating direction by actuation of the first pressure control valve 21 or is held in a stationary position.
  • the pressure medium volume in the annular space 40 is the second Differential cylinder 41 due to the action of the second check valve 110 and the action of the second overflow valve 83 included, so that movement of the differential piston in the direction of "early” is prevented until the overflow valve 83 opens.
  • the second differential cylinder 41 is accordingly moved into its "late” end position or held in this. If the differential cylinder is to be held in any intermediate position, a corresponding counterpressure must be built up in the pressure chamber 44 by correspondingly controlling the second pressure control valve 60. If the differential piston of the second differential cylinder 41 is to be moved from any intermediate position into a "late” position, the opening pressure of the second pressure control valve 60 is reduced accordingly.
  • the annular space 40 can be pressurized accordingly. This takes place on the one hand via a volume exchange between the pressure chamber 44 and the check valve 110 and on the other hand via a pressure medium inflow from the fourth outlet 97 of the flow divider via the orifice 106.
  • the restoring effect of the device for adjusting the camshaft also has a supporting effect in the case of such an adjustment of the differential piston.
  • the opening pressure is increased by appropriate actuation of the second pressure control valve 60, so that the inflow from the fourth outlet 97 via the second Line connection 105 and the second aperture 106 jams and the piston 87 of the flow divider 85 is moved to the left.
  • the pressure chamber 44 is simultaneously subjected to the corresponding pump pressure, so that an adjustment is made due to the larger effective piston area.
  • both differential cylinders 10 and 41 are to be moved from a "late" position to an “early” position, the larger volume flow is allocated to the differential cylinder, which exerts a larger adjusting force, via the flow divider 85. Due to the greater retroactive pressure at the first or second outlet 92 or 93 resulting from the greater adjusting force, the piston 87 is deflected, so that the outlet is throttled accordingly on the connection side of lower pressure. By controlling one of the two pressure limiting valves 21 and 60, this volume distribution function can be changed accordingly.

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Abstract

Le dispositif électrohydraulique de positionnement, destiné à actionner un dispositif afin de déplacer au moins un arbre à cames d'un moteur à combustion interne par rapport à son vilebrequin, comporte deux pistons différentiels (42, 12) qui servent chacun à déplacer un arbre à cames (arbre à cames d'admission et arbre à cames d'échappement). Les deux pistons différentiels (42, 12) sont mis chacun sous pression au niveau de leur petite surface (de piston) active, par une pompe. Une soupape de commande de pression autonome (21, 46; 60) est associée aux chambres de pression (11, 44) sur la plus grande surface active de piston et sert à moduler la pression dans cette chambre de pression, indépendamment de l'autre piston différentiel.

Claims (6)

  1. Installation de réglage des arbres à cames d'un moteur à combustion interne, comprenant une installation pour régler au moins un arbre à cames d'un moteur à combustion interne par rapport au vilebrequin avec au moins un piston différentiel (12, 42) dont la plus petite surface de piston (14, 40) est alimentée en liquide sous pression par une pompe (16, 100), alors que la pression dans la chambre de pression (11, 44) correspondant à sa plus grande surface de piston est variable par une soupape de commande électromagnétique (21, 60) ainsi qu'un second piston différentiel (42), indépendant, dont la petite surface de piston (40) est alimentée en liquide sous pression par la pompe (16, 100), alors que la pression dans la chambre de pression (11, 44) correspondant à la surface de piston plus grande est modifiée par une seconde soupape de commande électromagnétique (21, 60), ce second piston différentiel indépendant (42) pouvant régler un second arbre à cames du moteur à combustion interne par rapport au vilebrequin, caractérisée en ce que les deux soupapes de commande (21, 60) sont reliées l'une à l'autre par l'intermédiaire d'un répartiteur de débit (63, 85), et sont reliées à la pompe (16, 100).
  2. Installation de réglage d'arbres à cames selon la revendication 1, caractérisée en ce que le répartiteur de débit (63, 85) comporte en parallèle au moins une installation d'étranglement (70, 72).
  3. Installation de réglage d'arbres à cames selon l'une des revendications 1 à 2, caractérisée en ce qu'entre la pompe (16, 100) et la soupape de commande de pression (21, 60) il est prévu au moins une soupape de débordement réglable (32, 82, 83, 85, 86).
  4. Installation de réglage d'arbres à cames selon la revendication 3, caractérisée en ce que la soupape de débordement (82, 83, 85, 86) est prévue entre le répartiteur de débit (63, 85) et la soupape de commande de pression (21, 60).
  5. Installation de réglage d'arbres à cames selon l'une des revendications 1 à 4, caractérisée en ce qu'une soupape anti-retour (108, 110) est branchée en parallèle sur la soupape de débordement (82, 83).
  6. Installation de réglage d'arbres à cames selon l'une des revendications 1 à 5, caractérisée en ce qu'une installation d'étranglement (104, 106) est branchée en parallèle sur la soupape de débordement (82, 83).
EP93914613A 1992-07-25 1993-07-08 Dispositif electrohydraulique de positionnement Expired - Lifetime EP0653014B1 (fr)

Applications Claiming Priority (3)

Application Number Priority Date Filing Date Title
DE4224653 1992-07-25
DE4224653A DE4224653A1 (de) 1992-07-25 1992-07-25 Elektrohydraulische Stelleinrichtung
PCT/DE1993/000608 WO1994002715A1 (fr) 1992-07-25 1993-07-08 Dispositif electrohydraulique de positionnement

Publications (2)

Publication Number Publication Date
EP0653014A1 EP0653014A1 (fr) 1995-05-17
EP0653014B1 true EP0653014B1 (fr) 1996-05-15

Family

ID=6464123

Family Applications (1)

Application Number Title Priority Date Filing Date
EP93914613A Expired - Lifetime EP0653014B1 (fr) 1992-07-25 1993-07-08 Dispositif electrohydraulique de positionnement

Country Status (4)

Country Link
EP (1) EP0653014B1 (fr)
JP (1) JPH08500165A (fr)
DE (2) DE4224653A1 (fr)
WO (1) WO1994002715A1 (fr)

Families Citing this family (2)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE4128656C2 (de) * 1991-08-29 1994-10-20 Bosch Gmbh Robert Hydraulische Stelleinrichtung und deren Verwendung
DE19756017A1 (de) * 1997-12-17 1999-06-24 Porsche Ag Einrichtung zur relativen Drehlagenänderung einer Welle zum Antriebsrad

Family Cites Families (4)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE1262883C2 (de) * 1965-12-18 1975-02-13 Fa. Louis Schierholz, 2800 Bremen Vorrichtung zum synchronen antrieb von foerderstraengen in foerderanlagen zum foerdern von dung und dergleichen faserfoermigem gut
DE2356414A1 (de) * 1973-11-12 1975-05-15 Peiner Masch Schrauben Stromteiler
DE3616234A1 (de) * 1986-05-14 1987-11-19 Bayerische Motoren Werke Ag Vorrichtung zur relativen drehlagenaenderung zweier in antriebsverbindung stehender wellen, insbesondere zwischen in einem maschinengehaeuse einer brennkraftmaschine gelagerten kurbelwelle und nockenwelle
DE4037824A1 (de) * 1990-01-16 1991-07-18 Bosch Gmbh Robert Hydraulische stelleinrichtung

Also Published As

Publication number Publication date
DE4224653A1 (de) 1994-01-27
EP0653014A1 (fr) 1995-05-17
WO1994002715A1 (fr) 1994-02-03
DE59302622D1 (de) 1996-06-20
JPH08500165A (ja) 1996-01-09

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