EP0455761B1 - Hydraulic control device for valves of internal combustion engines - Google Patents

Hydraulic control device for valves of internal combustion engines Download PDF

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Publication number
EP0455761B1
EP0455761B1 EP90915296A EP90915296A EP0455761B1 EP 0455761 B1 EP0455761 B1 EP 0455761B1 EP 90915296 A EP90915296 A EP 90915296A EP 90915296 A EP90915296 A EP 90915296A EP 0455761 B1 EP0455761 B1 EP 0455761B1
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EP
European Patent Office
Prior art keywords
valve
pressure
reservoir
piston
control
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EP90915296A
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German (de)
French (fr)
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EP0455761A1 (en
Inventor
Helmut Rembold
Ernst Linder
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Robert Bosch GmbH
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Robert Bosch GmbH
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Publication of EP0455761A1 publication Critical patent/EP0455761A1/en
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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
    • F01L9/00—Valve-gear or valve arrangements actuated non-mechanically
    • F01L9/10—Valve-gear or valve arrangements actuated non-mechanically by fluid means, e.g. hydraulic
    • F01L9/11—Valve-gear or valve arrangements actuated non-mechanically by fluid means, e.g. hydraulic in which the action of a cam is being transmitted to a valve by a liquid column
    • F01L9/12—Valve-gear or valve arrangements actuated non-mechanically by fluid means, e.g. hydraulic in which the action of a cam is being transmitted to a valve by a liquid column with a liquid chamber between a piston actuated by a cam and a piston acting on a valve stem
    • F01L9/14—Valve-gear or valve arrangements actuated non-mechanically by fluid means, e.g. hydraulic in which the action of a cam is being transmitted to a valve by a liquid column with a liquid chamber between a piston actuated by a cam and a piston acting on a valve stem the volume of the chamber being variable, e.g. for varying the lift or the timing of a valve
    • 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
    • F01L1/344—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 changing the angular relationship between crankshaft and camshaft, e.g. using helicoidal gear
    • F01L1/3442—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 changing the angular relationship between crankshaft and camshaft, e.g. using helicoidal gear using hydraulic chambers with variable volume to transmit the rotating force
    • F01L2001/34423—Details relating to the hydraulic feeding circuit
    • F01L2001/34446—Fluid accumulators for the feeding circuit

Definitions

  • the invention relates to a hydraulic valve control device for an internal combustion engine according to the preamble of the main claim.
  • the pressure line is controlled via a 3/2-way valve by, according to a special exemplary embodiment (FIGS. 8 and 9), the directional valve in one switching position the pressure line with the pressure chamber of a valve lifter and in the other switching position with the pressure chamber of another valve lifter and this using only a single liquid reservoir for both pressure rooms.
  • One control position of the solenoid valve is used for two engine intake valves and only one accumulator is used for both intake valves.
  • the precision of the control i.e.
  • valve control device with the characterizing features of the main claim has the advantage that a low specific control pressure from the control line is sufficient to lift the accumulator piston from its valve seat. Since the control line is controlled by the solenoid valve, opening the solenoid valve in the delivery line, which is at a low pre-pressure, acts as a pressure surge of the control oil on the accumulator piston.
  • a pressure surface which acts against the force of the storage spring and is always acted upon by the pressure of the control oil in the pressure channel is present on the storage piston, the force of the storage spring being greater than the control force plus the pre-pressure force caused by this pressure surface.
  • This supporting actuating force, which acts on the accumulator piston from the pressure chamber on the pressure surface is then correspondingly large, when the associated valve lifter is being actuated by the drive cam and the high working pressure required for actuating the intake valve is thereby generated in the pressure chamber.
  • control pressure even with pressure surge alone is not sufficient here to lift the accumulator piston from the valve seat. It is made possible in a very simple manner that several control units are acted upon by the control pressure at the same time, and nevertheless only lift those accumulator pistons from the seat in which the valve tappet is also being actuated by the drive cam. Since this is a force-balanced system, a lower control pressure is sufficient, so that it is possible to work with simple low-pressure solenoid valves. As soon as the accumulator piston has lifted off its seat, its further displacement is brought about by the high pressure from the pressure chamber, as in the basic embodiment according to claim 1, since the control oil now acts on the entire end face of the accumulator piston beyond the pressure surface. In any case, however, the control pressure must be set very precisely in order to achieve the actual lifting of the accumulator piston from the seat at the desired point in time.
  • the valve control edge of the storage valve is preferably the bottom edge of the storage piston, which cooperates with a fixed seat, so that in the rest position or starting position of the storage piston the pressure channel is radially limited by the outer surface of the storage piston, while the storage space is limited by the end face of the storage piston.
  • an annular groove can be formed around the lateral surface in the area of the seat, so that the hydraulic fluid can flow evenly from all sides into the storage space after the storage piston has been lifted off the seat.
  • the pressure surface on the accumulator piston is formed by a step in its outer surface, so that the diameter of the valve seat is somewhat smaller than the diameter of the accumulator piston in its radially guided section and the resulting differential ring surface forms the pressure surface.
  • a slide control of this storage valve can of course also be provided, according to which the pressure channel is connected to the storage space only after a certain minimum path of the storage piston has been covered.
  • a relief line branches off from the storage space, in which a back-up throttle and possibly a pressure control valve are contained.
  • the relief line is preferably in the bottom of the storage piston is arranged and connects the storage space to the storage spring space, so that liquid quantities flowing out via the pressure holding valve can flow into the pressure-relieving storage spring space and from there into the oil container.
  • the switching precision is additionally increased by this pressure-maintaining valve, since this enables an exactly definable control pressure to be achieved in the storage space.
  • a pre-pressure accumulator is connected upstream of the solenoid valve to the delivery line.
  • This pre-pressure accumulator provides additional precision and maintenance of the control pressure, since the moment the solenoid valve opens, despite the rapid flow of a part of the quantity to the accumulator or control chamber, this pressure of the pre-pressure accumulator continues and causes a defined pressure surge there.
  • the solenoid valve is designed as a 2/2-way valve with the advantage of a high switching frequency and operational reliability with low manufacturing costs.
  • the force of the storage spring is less than the control force formed on the storage piston from the control pressure and the storage piston bottom, where appropriate, the control pressure set by the pressure-maintaining valve is lower than that lower pressure of the liquid source, and the filling line is controlled by the storage piston, the filling line being blocked after the connection between the pressure channel and the storage space has been established and opened again in the starting position of the storage piston.
  • this control can be such that a longitudinal groove is provided on the outer surface of the storage piston, which is in constant overlap with an annular groove in the bore receiving the storage piston and connected to the pressure channel in the rest position or starting position of the storage piston is, however, after the storage piston is displaced from its initial position against the force of the storage spring, is separated from the pressure channel.
  • a longitudinal groove is provided on the outer surface of the storage piston, which is in constant overlap with an annular groove in the bore receiving the storage piston and connected to the pressure channel in the rest position or starting position of the storage piston is, however, after the storage piston is displaced from its initial position against the force of the storage spring, is separated from the pressure channel.
  • a longitudinal groove is provided on the outer surface of the storage piston, which is in constant overlap with an annular groove in the bore receiving the storage piston and connected to the pressure channel in the rest position or starting position of the storage piston is, however, after the storage piston is displaced from its initial position against the force of the storage spring, is separated from the pressure channel.
  • the delivery line is connected upstream of the solenoid valve via a fill line to the pressure channel, with one in the direction in the fill line Pressure channel opening check valve is arranged. This compensates for any leakage losses that occur during operation and also maintains a constant upstream pressure in the pressure channel or pressure chamber in order to further specify the balance of forces.
  • the individual valve control units are each only up to a drive of 180 ° camshaft rotation angle (° NW) via the electronic control unit. controllable, so that several valve control units are controlled by only one solenoid valve, whereby overlaps of control times, that is to say switch-on times of the solenoid valve, are prevented above 180 ° NW per valve. Downstream of the solenoid valve, the control line is branched to the individual control units. The operating periods of these control units therefore have no overlaps above 180 ° angle of rotation of the crankshaft (° KW) from the start of the control process of the respective control unit.
  • a peculiarity of the hydraulic valve control devices is used, namely that with increasing speeds the final closing time is delayed in relation to the running angle of rotation of the crankshaft.
  • This delay in the closing process depends on the increasing Speed increasing mass acceleration forces together, as well as decreasing control periods with constant closing speed (determined by spring force), whereby the average pressure level in the pressure chamber of the ram drops.
  • the closing speed corresponds approximately to the cam speed.
  • the intake valve of the engine valve is designed so that it is reached about 60 - 80 ° KW after bottom dead center, ie after the turning point of the drive cam. This achieves maximum performance at high speeds. An increase in performance can no longer be achieved there via the engine valve control.
  • control units can be controlled via a valve control device according to the invention via only one solenoid valve, in which there is between there is no overlap of the valve strokes at these first 180 ° KW.
  • groups of valve control units can be controlled by a first division of the control line downstream of the solenoid valve by at least one pre-selection valve. This is particularly advantageous for engines with larger inlet closing angles.
  • the preselection valve is designed as a 2/2-way valve, in which case several such preselection valves are then connected in parallel.
  • the preselection valve is designed as a 3/2-way valve, two pressure spaces being controllable via a 3/2-way valve in connection with the control valve.
  • FIG. 1 shows a longitudinal section through the valve control device of an engine intake valve with the associated hydraulic circuit diagram
  • FIG. 2 shows a detail from FIG. 1 on an enlarged scale
  • Fig. 3 three stacked control diagrams of the opening movement of the valve
  • 4 and 5 show two variants of the hydraulic circuit diagram of FIG. 1
  • 6 shows a variant in the storage piston control with a corresponding and enlarged section from FIG. 1.
  • a hydraulic valve control device is shown in longitudinal section and as a hydraulic circuit diagram. This is arranged between a valve stem 2 carrying a valve plate and a drive cam 4 rotating with a camshaft 3.
  • the valve stem 2 is guided in an axially displaceable manner in a valve housing 5 and is loaded in the closing direction of the valve by valve closing springs 6 and 7, as a result of which the valve plate 1 is pressed onto a valve seat 8 in the valve housing 5.
  • the valve disk 1 controls a valve inlet opening 9 formed between it and the valve seat 8 when the valve is open.
  • the hydraulic valve control device has a control housing 11, which is inserted into a housing bore 10 of the engine valve housing 5 and in which a spring chamber 12 is arranged, the valve closing springs 6 and 7 being accommodated coaxially to one another in the spring chamber 12.
  • the control housing 11 is from a cup-shaped spring plate 13 anchored and axially displaceable with the valve stem 2 and loaded by the valve closing springs 6 and 7 is inserted below.
  • a valve piston 15 which interacts positively with the valve stem 2 of the inlet valve, and above it a working piston 16 of a cam piston 17 is arranged axially displaceably.
  • the working piston 16 is loaded by a return spring 18, which is supported on the one hand on a shoulder of the control housing 11 and on the other hand engages a flange of the working piston 16 and thereby presses the cam piston 17 against the valve control cams 4.
  • an oil-filled pressure chamber 19 is enclosed in the housing bore 14, the effective length of the entire valve tappet being determined by the amount of oil present in this pressure chamber 19.
  • the effective opening stroke of the intake valve is less; if the maximum filling is maintained, the stroke of the intake valve is at a maximum.
  • the pressure chamber 19 is connected via a pressure channel 21 to a storage valve 22 which has a radially sealing cup-shaped storage piston 23 which is loaded by a storage spring 24 its rest position shown rests on a valve seat 25.
  • the lower end face of the storage piston 23 delimits a storage space 26, while part of the outer surface of the storage piston 23 delimits an annular channel 27 surrounding the latter, into which the pressure channel 21 opens.
  • the valve control device works with a hydraulic circuit, with a feed pump 28, which sucks in the control oil from an oil tank 29 and supplies it to the valve control device via a feed line 31.
  • a pressure control valve 33 is arranged in a line 32 branching off from the delivery line 31 and returning to the oil container 29.
  • the delivery line 31 leads to a 2/2-way solenoid valve 34, which controls a control line 35, which leads to the storage space 26 via a check valve 36.
  • a supply pressure accumulator 37 is connected to the delivery line 31 shortly before the solenoid valve 34, the storage pressure of which is coordinated with the pressure control valve 33 and which is largely filled with control oil in the closed position of the solenoid valve 34 shown.
  • Additional control lines 38 branch off from the control line 35 and lead to further engine control valve units of the same engine, these units being designed in accordance with the one shown.
  • a filling line 39 branches off from the delivery line 31, which leads to the pressure channel 21 and in which a check valve 41 which opens towards the pressure channel 21 is arranged.
  • the storage valve 22 is shown on an enlarged scale.
  • the accumulator piston has a shoulder 42 on its outer surface, which creates a pressure shoulder 43 which acts in the opening direction of this valve. Accordingly, the diameter of the valve seat 25 is smaller than the diameter of the accumulator piston 23 in its radial guide area.
  • a spring plate 44 of a weak spring 45 is tensioned on the storage piston base within the cup-shaped storage piston 23, the spring 45 loading the movable valve member of a relief valve 46, which is arranged in a relief line 47, which connects the storage space 26 with the storage spring space 48 connects.
  • the relief line 47 is designed here as a throttle line, so that it acts as a throttle for an outflow of control oil from the storage space 26 to the storage spring space 48.
  • the relief valve 46 can also be designed as a pressure control valve, so as to maintain a specific admission pressure in the storage space 26.
  • the valve control device described in FIGS. 1 and 2 works as follows: When the camshaft 3 rotates, the cam piston 17 with the working piston 16 is moved downward against the return spring 18 via the drive cam 4 and displaces the hydraulic oil downward in the pressure chamber 19. The pressure generated thereby continues on the one hand via the pressure channel 21 to the accumulator valve 22, but acts primarily on the upper end face of the valve piston 15, which, including the valve stem 2 with the valve disk 1, is pushed downward against the force of the valve closing springs 6 and 7, wherein the valve plate 1 lifts off the valve seat 8 and releases the inlet opening 9, so that combustion air flows into the combustion chamber of the engine in accordance with the released cross section and the available opening time, that is to say in accordance with the opening time cross section.
  • the opening of the intake valve takes place synchronously with the suction strokes of the engine piston, whereby again the individual engine valves are opened one after the other in coordination with the firing order or the crank drive of the internal combustion engine, for example if the engine cylinders arranged next to one another are numbered I to IV, the opening or Firing order could be III, IV, II and finally I, after which the motor valve of cylinder III would open again in such a 4-cylinder internal combustion engine, etc.
  • the engine valve control is shown in a drive break, ie in a working position in which the base circle of the cam 4 interacts with the cam piston 17 and the valve disk 1 of the intake valve rests sealingly on its valve seat 8 driven by the valve closing springs 6 and 7.
  • Any leakage losses of hydraulic oil in the pressure chamber 19 that occur during operation are compensated for via the fill line 39, via the hydraulic oil under delivery pressure via the check valve 41 into the pressure channel 21 and so that it can flow into the pressure chamber 19.
  • a pre-pressure which is always the same during the drive breaks is generated in the pressure chamber 19 and cavities are also avoided, which could lead to control errors with respect to the opening time but also the opening stroke of the engine valve.
  • the delivery pressure prevailing in the pre-pressure accumulator 37 is transmitted from the delivery line 31 via the control line 35 and the check valve 36 into the storage space 26, so that the lower end face of the storage piston 23 is acted upon by a control pressure which is only slight is lower than the delivery pressure in the delivery line 31.
  • This control pressure generates, in relation to the end face acted upon, a force acting on the accumulator piston in the opening direction, which is less than the force of the accumulator spring 24.
  • the opening stroke of the engine valve is correspondingly reduced, as a result of which the opening time cross section is also reduced.
  • Such a change in the opening time cross-section has an effect on the intake air volume of the engine and thus directly on the speed of the engine.
  • the solenoid valve 34 is only reversed when the opening stroke of the engine valve has already started, that is, when the drive cam 4 has already started to displace the working piston 16.
  • the control lines 38 are also supplied with hydraulic oil under control pressure, so that, in addition to the storage piston 23 shown, a number of storage pistons belonging to other engine valve controls of the same engine with hydraulic oil under control pressure.
  • the store 37 serves the purpose of which the store volume is designed accordingly. While the accumulator is charging in the times in which the solenoid valve 34 is closed, so that its pre-pressure accumulator piston 49 assumes the position shown, when the solenoid valve 34 is open, this pre-pressure accumulator piston shifts further down, for example into the position shown in broken lines.
  • the maximum output of the feed pump 28 can be kept correspondingly lower and, in addition, a high delivery rate is made available for a short time, so that a kind of pressure surge takes place on the respective loaded piston 23.
  • the attacking forces of control pressure, admission pressure and springs are coordinated with one another in such a way that only the accumulator pistons 23 lift off their seat 25, which are additionally acted upon on their pressure shoulder 43 by the working pressure, which can only occur when the working cam 4 acts on the working piston 16.
  • a sufficient fill pressure is maintained in the storage space 26 via the relief valve 46 in connection with its spring 45.
  • the control pressure is also set in this storage space 26, which is, however, higher than the pilot pressure of the relief valve 46 so that it opens. Since the relief line 47 is formed in the accumulator piston bottom as a throttle line, this creates a jam so that the control pressure can be maintained in the accumulator chamber 26. In any case, the delivery rate of the delivery pump 28 is greater than the amount of hydraulic oil flowing out over all simultaneously connected storage spaces 26 and their relief lines 47. As soon as the working pressure from the pressure channel 21 is added to the pressure shoulder 43 of the accumulator piston 23, this accumulator piston 23 lifts off the seat 25 and the check valve 36 is blocked by the working pressure which is far higher than the control pressure in the control line 35.
  • the advantage is that the timing of opening the solenoid valve 34 initiates the closing of the engine valve, with the further closing movement of the engine valve caused by the valve closing springs 6 and 7 - apart from the pressures in the combustion chamber itself which act on the valve plate 1 - being determined by the evasion speed of the accumulator piston 23.
  • Fig. 3 the working stroke of the valve for three different speeds is shown on the basis of three diagrams.
  • the stroke of the engine valve h (ordinate) is shown above the degree of rotation of the crankshaft as ° KW (abscissa).
  • the first diagram a is for an engine speed of 1000 rpm; the second diagram b corresponds to a speed of 3000 rpm and the lowest diagram c applies to a speed of 5000 rpm.
  • the outer jacket curve in all three diagrams corresponds to the opening and closing of the inlet valve without the influence of the control via the solenoid valve 34.
  • the group of curves shown in dash-dotted lines in each diagram corresponds to a shortening of the opening stroke or the opening time due to the action of the solenoid valve 34, that is to say by opening the same and take effect of the storage valve 22. While the course of the opening portion of the curves is the same for all curves, the closing course is different.
  • the opening stroke section of the curve is determined solely by the drive cam 4, which always has the same opening effect on the engine valve. This also applies to the Closing action corresponding to the trajectory of the drive cam 4. However, as soon as the solenoid valve 34 has opened, the section of the curve corresponding to the closing of the motor valve is determined by the influences described above, above all by the action of the accumulator piston 23.
  • a valve control carried out via the solenoid valve 34 at 180 ° KW no longer has an effect, since at these high speeds the inlet closure with that at 240 ° KW would coincide with how it takes place without control anyway.
  • Time cross-section controls at maximum speed and with low load or power is controlled in that the solenoid valve 34 is switched on accordingly below 180 ° KW.
  • a control system above 180 ° KW could theoretically have an effect, only it is not necessary there.
  • the inlet valve In the range up to 3000 rpm, the inlet valve is normally closed at 180 ° KW in order to obtain the maximum power yield required there. In diagram a, this corresponds to switching the solenoid valve 34 at approximately 160 ° KW and at 3000 rpm according to diagram b at 130 ° KW.
  • the control is designed according to the invention in such a way that the solenoid valve 34 is opened only up to 180 ° KW.
  • This makes it theoretically possible to control all four motor cylinders with only one solenoid valve, in which the solenoid valve is turned on at least twice per revolution and only the storage valve 22 always lifts off its seat 25, the cam 4 of which is currently active in the associated motor valve. Since no control is required in the remaining 180 ° to 240 ° KW, there can be no overlap.
  • the solenoid valve 34 and corresponding branching of the control line 35 in each of the two control lines 38 are in turn arranged a 2/2 solenoid valve 51, the control lines 38 being branched further downstream of these solenoid valves 51 are to lead to the individual engine valve control units.
  • control line 35 of the solenoid valve 34 opens into the input of a 3/2-way solenoid valve 52, the outputs of which in turn lead to the control lines 38, which then branch out again to the individual engine valve control units.
  • FIG. 6 shows a variant for the control of the filling line 39, the opening of the filling line 39 taking place downstream of the check valve 42 through the storage piston 23.
  • the filling line 39 opens into an annular groove 53 in the bore wall, in which the storage piston 23 is guided in a radially sealing manner, this annular groove 53 being connected to the pressure channel 21 via a longitudinal groove 54 of limited length in the illustrated rest position of the storage piston 23.
  • the longitudinal groove 54 is separated from the pressure channel 21 by the displacement of the accumulator piston 23, so that in such a shifted position no hydraulic oil can get into the pressure channel 21 from the filling line 39.
  • the pressure balance in the control system can be refined, so that there are no incorrect controls even at high speed and a correspondingly lower working pressure take place in that the storage valve 22 opens unwanted.
  • the force which acts on the accumulator piston 23 by the accumulator spring 24 can then be less than that which acts on the accumulator piston 23 in the opening direction, which is caused by the admission pressure when it acts on the entire end face.

Abstract

A hydraulic control device for valves of multi-cylinder internal combustion engines comprises a storage chamber (26) associated with the pressure chamber (19) of the valve tappet. The storage chamber (26) has a storage piston (23) which acts at the same time as a valve and which enables the storage chamber (26) to be separated from the pressure chamber (19). The storage piston (23) is displaced from its rear position to assume its storage function by means of a hydraulic impact controlled by a solenoid valve (34) and transmitted to the storage chamber (26) via a control line (35), in which a nonreturn valve (36) is arranged. The storage piston (23) is displaced only when the valve is actuated by a drive cam (4) when the working pressure in the pressure chamber (19) is sufficient to displace the storage piston (23).

Description

Stand der TechnikState of the art

Die Erfindung geht aus von einer hydraulischen Ventilsteuervorrichtung für eine Brennkraftmaschine nach der Gattung des Hauptanspruchs.The invention relates to a hydraulic valve control device for an internal combustion engine according to the preamble of the main claim.

Bei einer bekannten hydraulischen Ventilsteuervorrichtung der gattungsgemäßen Art (DE-OS 3 511 820) wird über ein 3/2-Wegeventil die Druckleitung gesteuert, indem gemäß einem speziellen Ausführungsbeispiel (Fig. 8 und 9) das Wegeventil in der einen Schalt-Stellung die Druckleitung mit dem Druckraum eines Ventilstößels und in der anderen Schaltstellung mit dem Druckraum eines anderen Ventilstößels verbindet und dies unter Verwendung nur eines einzigen Flüssigkeitsspeichers für beide Druckräume. Es werden also für zwei Motoreinlaßventile je eine Steuerstellung des Magnetventils und für beide Einlaßventile nur ein Speicher verwendet. Die Präzision der Steuerung, d.h. wie genau der angestrebte Öffnungszeitquerschnitt des Motorventils erreichbar ist, hängt besonders bei hohen Drehzahlen davon ab, wie groß das gesamte Ölvolumen ist, das bei der Steuerung hin und her geschoben werden muß und wieviele Steuerkanäle mit entsprechenden Steuerquerschnitten durchströmt werden müssen. Für die Kosten und die Störanfälligkeit einer solchen hydraulischen Ventilsteuervorrichtung ist vor allem das Magnetventil beachtlich, wobei bei Motoren üblicher Maximaldrehzahl die mögliche Schaltfrequenz dieser Magnetventile bei weitem nicht ausgenutzt ist. Hinzu kommt die Belastung für die Kosten jedes extra Magnetventils.In a known hydraulic valve control device of the generic type (DE-OS 3 511 820), the pressure line is controlled via a 3/2-way valve by, according to a special exemplary embodiment (FIGS. 8 and 9), the directional valve in one switching position the pressure line with the pressure chamber of a valve lifter and in the other switching position with the pressure chamber of another valve lifter and this using only a single liquid reservoir for both pressure rooms. One control position of the solenoid valve is used for two engine intake valves and only one accumulator is used for both intake valves. The precision of the control, i.e. how exactly the desired opening time cross-section of the engine valve can be reached, depends, in particular at high speeds, on the size of the total oil volume that has to be pushed back and forth in the control and how many control channels with corresponding control cross-sections have to be flowed through . The solenoid valve is particularly noteworthy for the costs and the susceptibility to malfunction of such a hydraulic valve control device, the possible switching frequency of these solenoid valves being largely underutilized in the case of motors of the usual maximum speed. Added to this is the burden on the cost of each extra solenoid valve.

Es ist auch schon vorgeschlagen worden (DE-P 38 15 668.7), bei einer gattungsgemäßen hydraulischen Ventilsteuervorrichtung den Speicherkolben als bewegliches Ventilglied auszubilden, wobei die Stirnkante des Kolbens mit einem Ventilsitz zusammenwirkt, wodurch die Verbindung zwischen Druckleitung und Speicherraum steuerbar ist. Der Speicherkolben dient gleichzeitig als Anker eines stromlos offenen Magnetventils, so daß bei erregtem Magnet die Druckleitung vom Speicherraum getrennt ist. Zwar ist durch diese Lösung eine Kombination von Flüssigkeitsspeicher und Magnetventil erreicht, bei der das gleiche Teil einerseits als bewegliches Ventilglied des Magnetventils und andererseits als Speicherkolben dient, es muß jedoch für jede Ventilsteuereinheit eine solche "Magnetventilspeichereinheit" zur Verfügung stehen.It has also been proposed (DE-P 38 15 668.7) to design the accumulator piston as a movable valve member in a generic hydraulic valve control device, the end edge of the piston cooperating with a valve seat, whereby the connection between the pressure line and the accumulator chamber can be controlled. The storage piston also serves as the armature of a normally open solenoid valve, so that the pressure line is separated from the storage space when the magnet is energized. Although this solution is a combination of liquid storage and Solenoid valve reached, in which the same part serves on the one hand as a movable valve member of the solenoid valve and on the other hand as a storage piston, but such a "solenoid valve storage unit" must be available for each valve control unit.

Vorteile der ErfindungAdvantages of the invention

Die erfindungsgemäße Ventilsteuervorrichtung mit den kennzeichnenden Merkmalen des Hauptanspruchs hat demgegenüber den Vorteil, daß bereits ein niedriger bestimmter Steuerdruck von der Steuerleitung her genügt, um den Speicherkolben von seinem Ventilsitz abzuheben. Da die Steuerleitung durch das Magnetventil gesteuert wird, wirkt sich ein Öffnen des Magnetventils in der unter geringem Vordruck stehenden Förderleitung als Druckstoß des Steueröls auf den Speicherkolben aus.The valve control device according to the invention with the characterizing features of the main claim has the advantage that a low specific control pressure from the control line is sufficient to lift the accumulator piston from its valve seat. Since the control line is controlled by the solenoid valve, opening the solenoid valve in the delivery line, which is at a low pre-pressure, acts as a pressure surge of the control oil on the accumulator piston.

Nach einer vorteilhaften Ausgestaltung der Erfindung ist am Speicherkolben eine entgegen der Kraft der Speicherfeder wirkende und vom Druck des im Druckkanal vorhandenen Steueröls stets beaufschlagte Druckfläche vorhanden, wobei die Kraft der Speicherfeder größer ist als die Steuerkraft zuzüglich der durch diese Druckfläche bewirkten Vordruckkraft. Diese unterstützende Betätigungskraft, die am Speicherkolben vom Druckraum her an der Druckfläche angreift, ist dann entsprechend groß,
wenn der zugeordnete Ventilstößel gerade durch den Antriebsnocken betätigt wird und dadurch im Druckraum der hohe zur Einlaßventilbetätigung erforderliche Arbeitsdruck erzeugt wird. Bei der gleichzeitigen Beaufschlagung von mehreren Speicherkolben durch den Steuerdruck bleibt bei dieser Ausgestaltung der Erfindung somit der Druckstoß bei all den Speicherkolben unwirksam, bei denen auch der Antriebsnocken des Motorventils gerade unwirksam ist. Der Steuerdruck auch mit Druckstoß reicht hier alleine nicht aus, um den Speicherkolben vom Ventilsitz abzuheben. Durch sehr einfache Weise wird damit ermöglicht, daß mehrere Steuereinheiten gleichzeitig vom Steuerdruck beaufschlagt werden, und trotzdem nur jene Speicherkolben vom Sitz abheben, bei denen auch der Ventilstößel gerade vom Antriebsnocken betätigt wird. Da es sich hier um ein kräfteausgewogenes System handelt, genügt bereits ein geringerer Steuerdruck, so daß es möglich ist mit einfachen Niederdruckmagnetventilen zu arbeiten. Sobald der Speicherkolben einmal von seinem Sitz abgehoben hat, wird wie bei der Grundausführung nach Anspruch 1 seine weitere Verschiebung durch den hohen Druck vom Druckraum her bewirkt, da nunmehr das Steueröl über die Druckfläche hinaus die ganze Stirnfläche des Speicherkolbens beaufschlagt. Allerdings muß in jedem Fall der Steuerdruck recht genau eingestellt sein, um zu dem gewünschten Zeitpunkt auch das tatsächliche Abheben des Speicherkolbens vom Sitz zu erzielen.According to an advantageous embodiment of the invention, a pressure surface which acts against the force of the storage spring and is always acted upon by the pressure of the control oil in the pressure channel is present on the storage piston, the force of the storage spring being greater than the control force plus the pre-pressure force caused by this pressure surface. This supporting actuating force, which acts on the accumulator piston from the pressure chamber on the pressure surface, is then correspondingly large,
when the associated valve lifter is being actuated by the drive cam and the high working pressure required for actuating the intake valve is thereby generated in the pressure chamber. When the control pressure acts on a plurality of storage pistons at the same time, in this embodiment of the invention the pressure surge thus remains ineffective in all the storage pistons in which the drive cam of the engine valve is also currently ineffective. The control pressure even with pressure surge alone is not sufficient here to lift the accumulator piston from the valve seat. It is made possible in a very simple manner that several control units are acted upon by the control pressure at the same time, and nevertheless only lift those accumulator pistons from the seat in which the valve tappet is also being actuated by the drive cam. Since this is a force-balanced system, a lower control pressure is sufficient, so that it is possible to work with simple low-pressure solenoid valves. As soon as the accumulator piston has lifted off its seat, its further displacement is brought about by the high pressure from the pressure chamber, as in the basic embodiment according to claim 1, since the control oil now acts on the entire end face of the accumulator piston beyond the pressure surface. In any case, however, the control pressure must be set very precisely in order to achieve the actual lifting of the accumulator piston from the seat at the desired point in time.

Als Ventilsteuerkante des Speicherventils dient hierbei vorzugsweise die Bodenkante des Speicherkolbens, die mit einem feststehenden Sitz zusammenwirkt, so daß in der Ruhelage oder Ausgangslage des Speicherkolbens der Druckkanal radial durch die Mantelfläche des Speicherkolbens begrenzt wird, während der Speicherraum durch die Stirnfläche des Speicherkolbens begrenzt ist. Hierfür kann beispielsweise im Bereich des Sitzes eine Ringnut um die Mantelfläche gebildet sein, so daß die Hydraulikflüssigkeit nach Abheben des Speicherkolbens vom Sitz gleichmäßig von allen Seiten in den Speicherraum strömen kann. Vorzugsweise wird auch die Druckfläche am Speicherkolben durch eine Stufe, in dessen Mantelfläche gebildet, so daß der Durchmesser des Ventilsitzes etwas kleiner ist als der Durchmesser des Speicherkolbens in seinem radial geführten Abschnitt und wobei die sich dabei ergebende Differenzringfläche die Druckfläche bildet.The valve control edge of the storage valve is preferably the bottom edge of the storage piston, which cooperates with a fixed seat, so that in the rest position or starting position of the storage piston the pressure channel is radially limited by the outer surface of the storage piston, while the storage space is limited by the end face of the storage piston. For this purpose, for example, an annular groove can be formed around the lateral surface in the area of the seat, so that the hydraulic fluid can flow evenly from all sides into the storage space after the storage piston has been lifted off the seat. Preferably, the pressure surface on the accumulator piston is formed by a step in its outer surface, so that the diameter of the valve seat is somewhat smaller than the diameter of the accumulator piston in its radially guided section and the resulting differential ring surface forms the pressure surface.

Statt einer Sitzsteuerung kann natürlich auch eine Schiebersteuerung dieses Speicherventils vorgesehen sein, gemäß der erst nach Zurücklegung eines bestimmten Minimalweges des Speicherkolbens der Druckkanal mit dem Speicherraum verbunden wird.Instead of a seat control, a slide control of this storage valve can of course also be provided, according to which the pressure channel is connected to the storage space only after a certain minimum path of the storage piston has been covered.

Nach einer vorteilhaften Ausgestaltung der Erfindung zweigt vom Speicherraum eine Entlastungsleitung ab, in welcher eine Staudrossel und möglicherweise ein Druckhalteventil enthalten sind. Die Entlastungsleitung ist vorzugsweise im Boden des Speicherkolbens angeordnet ist und verbindet den Speicherraum mit dem Speicherfederraum, so daß über das Druckhalteventil abströmende Flüssigkeitsmengen in den grundsätzlich druckentlasteten Speicherfederraum und von dort in den Ölbehälter strömen können. Durch dieses Druckhalteventil wird die Schaltpräzision zusätzlich erhöht, da hierdurch im Speicherraum ein exakter definierbarer Steuerdruck erzielbar ist.According to an advantageous embodiment of the invention, a relief line branches off from the storage space, in which a back-up throttle and possibly a pressure control valve are contained. The relief line is preferably in the bottom of the storage piston is arranged and connects the storage space to the storage spring space, so that liquid quantities flowing out via the pressure holding valve can flow into the pressure-relieving storage spring space and from there into the oil container. The switching precision is additionally increased by this pressure-maintaining valve, since this enables an exactly definable control pressure to be achieved in the storage space.

Nach einer weiteren vorteilhaften Ausgestaltung der Erfindung ist stromauf des Magnetventils an die Förderleitung ein Vordruckspeicher angeschlossen. Durch diesen Vordruckspeicher wird eine zusätzliche Präzisierung sowie Aufrechterhaltung des Steuerdrucks erzielt, da in dem Moment, in dem das Magnetventil aufmacht, sich trotz schnellen Wegströmens einer Teilmenge zum Speicher- bzw. Steuerraum hin dieser Druck des Vordruckspeichers fortsetzt und dort einen definierten Druckstoß bewirkt.According to a further advantageous embodiment of the invention, a pre-pressure accumulator is connected upstream of the solenoid valve to the delivery line. This pre-pressure accumulator provides additional precision and maintenance of the control pressure, since the moment the solenoid valve opens, despite the rapid flow of a part of the quantity to the accumulator or control chamber, this pressure of the pre-pressure accumulator continues and causes a defined pressure surge there.

Nach einer weiteren vorteilhaften Ausgestaltung der Erfindung ist das Magnetventil als 2/2-Wegeventil ausgebildet mit dem Vorteil einer hohen Schaltfrequenz und Betriebssicherheit bei geringem Fertigungsaufwand.According to a further advantageous embodiment of the invention, the solenoid valve is designed as a 2/2-way valve with the advantage of a high switching frequency and operational reliability with low manufacturing costs.

Nach einer weiteren vorteilhaften Ausgestaltung der Erfindung ist die Kraft der Speicherfeder geringer als die aus Steuerdruck und Speicherkolbenboden gebildete, am Speicherkolben angreifende Aufsteuerkraft, wobei gegebenenfalls der durch das Druckhalteventil eingestellte Steuerdruck geringer ist als jener niedere Druck der Flüssigkeitsquelle, und es ist die Auffülleitung durch den Speicherkolben gesteuert, wobei die Auffülleitung nach Herstellen der Verbindung zwischen Druckkanal und Speicherraum gesperrt wird und in der Ausgangslage des Speicherkolbens wieder geöffnet ist.According to a further advantageous embodiment of the invention, the force of the storage spring is less than the control force formed on the storage piston from the control pressure and the storage piston bottom, where appropriate, the control pressure set by the pressure-maintaining valve is lower than that lower pressure of the liquid source, and the filling line is controlled by the storage piston, the filling line being blocked after the connection between the pressure channel and the storage space has been established and opened again in the starting position of the storage piston.

Nach einer weiteren Ausgestaltung der Erfindung kann diese Steuerung derart sein, daß auf der Mantelfläche des Speicherkolbens eine Längsnut vorhanden ist, die in stetiger Überdeckung mit einer in der den Speicherkolben aufnehmenden Bohrung vorhandenen Ringnut steht und in Ruhestellung bzw. Ausgangsstellung des Speicherkolbens mit dem Druckkanal verbunden ist, jedoch, nachdem der Speicherkolben aus seiner Ausgangsstellung entgegen der Kraft der Speicherfeder verschoben wird, von dem Druckkanal getrennt wird. Natürlich können statt einer Längsnut mehrere derartige Längsnuten oder auch eine Ringnut auf der Mantelfläche des Speicherkolbens angeordnet sein. Maßgebend ist, daß die Verbindung zwischen Auffülleitung und Druckkanal nach Abheben des Speicherkolbens von seinem Sitz unterbrochen wird.According to a further embodiment of the invention, this control can be such that a longitudinal groove is provided on the outer surface of the storage piston, which is in constant overlap with an annular groove in the bore receiving the storage piston and connected to the pressure channel in the rest position or starting position of the storage piston is, however, after the storage piston is displaced from its initial position against the force of the storage spring, is separated from the pressure channel. Of course, instead of one longitudinal groove, several such longitudinal grooves or an annular groove can be arranged on the outer surface of the storage piston. The decisive factor is that the connection between the filling line and the pressure channel is interrupted after the storage piston is lifted off its seat.

Nach einer weiteren vorteilhaften Ausgestaltung der Erfindung ist die Förderleitung stromauf des Magnetventils über eine Auffülleitung mit dem Druckkanal verbunden, wobei in der Auffülleitung ein in Richtung Druckkanal öffnendes Rückschlagventil angeordnet ist. Hierdurch werden sich während des Betriebes einstellende Leckverluste ausgeglichen und es wird außerdem ein konstanter Vordruck im Druckkanal bzw. Druckraum eingestellt, um die Kräftebilanz zusätzlich zu präzisieren.According to a further advantageous embodiment of the invention, the delivery line is connected upstream of the solenoid valve via a fill line to the pressure channel, with one in the direction in the fill line Pressure channel opening check valve is arranged. This compensates for any leakage losses that occur during operation and also maintains a constant upstream pressure in the pressure channel or pressure chamber in order to further specify the balance of forces.

Nach einer weiteren vorteilhaften Ausgestaltung der Erfindung für die selbständiger Schutz beansprucht wird und die den Einsatz der Erfindung nur bei mehrzylindrigen Brennkraftmaschinen betrifft, sind über das elektronische Steuergerät die einzelnen Ventilsteuereinheiten (Magnetventile) jeweils nur bis zu einem Antrieb von 180° Nockenwellenverdrehwinkel (°NW) steuerbar, so daß mehrere Ventilsteuereinheiten durch nur ein Magnetventil gesteuert werden, wobei Überschneidungen von Steuerzeiten, also Einschaltzeiten des Magnetventils, oberhalb von 180° NW pro Ventil unterbunden sind. Stromab des Magnetventils ist die Steuerleitung zu den einzelnen Steuereinheiten verzweigt. Die Betätigungszeitabschnitte dieser Steuereinheiten weisen also keine Überschneidungen oberhalb von 180° Drehwinkel der Kurbelwelle (°KW) ab Beginn des Aufsteuervorgangs der jeweiligen Steuereinheit auf. Hierbei wird eine Eigenart der hydraulischen Ventilsteuereinrichtungen ausgenutzt, daß nämlich bei zunehmenden Drehzahlen der endgültige Schließzeitpunkt sich in Bezug auf den ablaufenden Drehwinkel der Kurbelwelle verspätet. Diese Verzögerung des Schließvorgangs hängt mit den mit zunehmender Drehzahl steigenden Massebeschleunigungskräften zusammen, sowie mit abnehmenden Steuerzeitabschnitten bei gleichbleibender Schließgeschwindigkeit (federkraftbestimmt), wobei das mittlere Druckniveau im Druckraum des Stößels absinkt. Bei hoher Drehzahl entspricht die Schließgeschwindigkeit in etwa der Nockengeschwindigkeit. Zudem ist bei hoher Drehzahl der Einlaßschluß des Motorventils so ausgelegt, daß er etwa 60 - 80° KW nach unterem Totpunkt, d. h. nach dem Wendepunkt der Antriebsnockenbahn erreicht wird. Hierdurch wird bei hoher Drehzahl eine maximale Leistung erzielt. Eine Steigerung der Leistung ist über die Motorventilsteuerung dort nicht mehr erreichbar. Anders ist es bei niedriger Motordrehzahl, bei der beispielsweise bei Einlaßschluß um 180° KW durch möglichst Frühlegen des endgültigen Schließens des Motorventils eine Leistungssteuerung erzielbar ist. In jedem Fall jedoch ist bei mittleren und niederen Drehzahlen in dem Bereich größer 180° KW kein durch die Ventilschließfedern bewirkter Hochdruck mehr im Druckraum oder Druckkanal vorhanden. Hierbei wird davon ausgegangen, daß der untere Totpunkt, also der Wendepunkt der Antriebsnockenbahn bei 120° NW liegt. Je früher nunmehr das Motorventil schließen soll, d. h. je früher das Speicherventil aufgesteuert wird, desto geringer sind diese Auswirkungen der Schließkräfte, so daß für eine vernünftige Steuerung vorteilhafterweise all die Steuereinheiten über eine erfindungsgemäße Ventilsteuervorrichtung über nur ein Magnetventil steuerbar sind, bei denen es zwischen den Ventilhüben bei diesen ersten 180° KW keine zeitlichen Überschneidungen gibt.According to a further advantageous embodiment of the invention for which independent protection is claimed and which relates to the use of the invention only in multi-cylinder internal combustion engines, the individual valve control units (solenoid valves) are each only up to a drive of 180 ° camshaft rotation angle (° NW) via the electronic control unit. controllable, so that several valve control units are controlled by only one solenoid valve, whereby overlaps of control times, that is to say switch-on times of the solenoid valve, are prevented above 180 ° NW per valve. Downstream of the solenoid valve, the control line is branched to the individual control units. The operating periods of these control units therefore have no overlaps above 180 ° angle of rotation of the crankshaft (° KW) from the start of the control process of the respective control unit. Here, a peculiarity of the hydraulic valve control devices is used, namely that with increasing speeds the final closing time is delayed in relation to the running angle of rotation of the crankshaft. This delay in the closing process depends on the increasing Speed increasing mass acceleration forces together, as well as decreasing control periods with constant closing speed (determined by spring force), whereby the average pressure level in the pressure chamber of the ram drops. At high speed, the closing speed corresponds approximately to the cam speed. In addition, at high speed, the intake valve of the engine valve is designed so that it is reached about 60 - 80 ° KW after bottom dead center, ie after the turning point of the drive cam. This achieves maximum performance at high speeds. An increase in performance can no longer be achieved there via the engine valve control. It is different at a low engine speed, in which, for example, a power control can be achieved by closing the engine valve as soon as possible by closing the engine valve as soon as possible. In any case, however, at medium and low speeds in the range greater than 180 ° KW there is no longer any high pressure caused by the valve closing springs in the pressure chamber or pressure channel. It is assumed that bottom dead center, i.e. the turning point of the drive cam track, is 120 ° NW. The earlier the motor valve is to close, i.e. the earlier the storage valve is opened, the less these effects of the closing forces are, so that for a sensible control, advantageously all the control units can be controlled via a valve control device according to the invention via only one solenoid valve, in which there is between there is no overlap of the valve strokes at these first 180 ° KW.

Nach einer weiteren vorteilhaften Ausgestaltung der Erfindung sind Gruppen von Ventilsteuereinheiten nach einer ersten Aufteilung der Steuerleitung stromab des Magnetventils durch mindestens ein Vorwahlventil für sich steuerbar. Dies ist besonders vorteilhaft bei Motoren mit größeren Einlaßschließwinkeln anwendbar.According to a further advantageous embodiment of the invention, groups of valve control units can be controlled by a first division of the control line downstream of the solenoid valve by at least one pre-selection valve. This is particularly advantageous for engines with larger inlet closing angles.

Nach einer weiteren vorteilhaften Ausgestaltung der Erfindung ist das Vorwahlventil als 2/2-Wegeventil ausgebildet, wobei dann entsprechend mehrere solche Vorwahlventile parallel geschaltet sind.According to a further advantageous embodiment of the invention, the preselection valve is designed as a 2/2-way valve, in which case several such preselection valves are then connected in parallel.

Nach einer anderen vorteilhaften Ausgestaltung der Erfindung ist das Vorwahlventil als 3/2-Wegeventil ausgebildet, wobei über ein 3/2-Wegeventil in Verbindung mit dem Steuerventil jeweils zwei Druckräume steuerbar sind.According to another advantageous embodiment of the invention, the preselection valve is designed as a 3/2-way valve, two pressure spaces being controllable via a 3/2-way valve in connection with the control valve.

Weitere Vorteile und vorteilhafte Ausgestaltungen der Erfindung sind der nachfolgenden Beschriebung, der Zeichnung und den Ansprüchen entnehmbar.Further advantages and advantageous embodiments of the invention can be found in the following description, the drawing and the claims.

Zeichnungdrawing

Es zeigen Fig. 1 einen Längsschnitt durch die Ventilsteuervorrichtung eines Motoreinlaßventils mit dem dazugehörigen Hydraulikschaltplan; Fig. 2 einen Ausschnitt aus Fig. 1 in vergrößertem Maßstab; Fig. 3 drei übereinander angeordnete Steuerdiagramme der Öffnungsbewegung des Ventils; Fig. 4 und Fig. 5 zwei Varianten des Hydraulikschaltplans von Fig. 1; Fig. 6 eine Variante in der Speicherkolbensteuerung mit einem entsprechenden sowie vergrößerten Ausschnitt aus Fig. 1.1 shows a longitudinal section through the valve control device of an engine intake valve with the associated hydraulic circuit diagram; FIG. 2 shows a detail from FIG. 1 on an enlarged scale; Fig. 3 three stacked control diagrams of the opening movement of the valve; 4 and 5 show two variants of the hydraulic circuit diagram of FIG. 1; 6 shows a variant in the storage piston control with a corresponding and enlarged section from FIG. 1.

Beschreibung des AusführungsbeispielsDescription of the embodiment

In Fig. 1 ist eine erfindungsgemäße hydraulische Ventilsteuervorrichtung im Längsschnitt sowie als Hydraulikschaltplan dargestellt. Diese ist zwischen einem einen Ventilteller tragendem Ventilschaft 2 und einem mit einer Nockenwelle 3 umlaufenden Antriebsnocken 4 angeordnet. Der Ventilschaft 2 ist in einem Ventilgehäuse 5 axial verschiebbar geführt und ist in Schließrichtung des Ventils durch Ventilschließfedern 6 und 7 belastet, wodurch der Ventilteller 1 auf einen Ventilsitz 8 im Ventilgehäuse 5 gepreßt wird. Der Ventilteller 1 steuert eine zwischen ihm und dem Ventilsitz 8 bei geöffnetem Ventil gebildete Ventileinlaßöffnung 9.In Fig. 1, a hydraulic valve control device according to the invention is shown in longitudinal section and as a hydraulic circuit diagram. This is arranged between a valve stem 2 carrying a valve plate and a drive cam 4 rotating with a camshaft 3. The valve stem 2 is guided in an axially displaceable manner in a valve housing 5 and is loaded in the closing direction of the valve by valve closing springs 6 and 7, as a result of which the valve plate 1 is pressed onto a valve seat 8 in the valve housing 5. The valve disk 1 controls a valve inlet opening 9 formed between it and the valve seat 8 when the valve is open.

Die hydraulische Ventilsteuervorrichtung weist ein in eine Gehäusebohrung 10 des Motorventilgehäuses 5 eingesetztes Steuergehäuse 11 auf, in welchem eine Federkammer 12 angeordnet ist, wobei in der Federkammer 12 die Ventilschließfedern 6 und 7 koaxial zueinander untergebracht sind. Im Steuergehäuse 11 ist von unten her ein mit dem Ventilschaft 2 verankerter und axial verschiebbarer sowie durch die Ventilschließfedern 6 und 7 belasteter topfförmiger Federteller 13 eingeschoben. In einer zentralen axial durchgehenden Bohrung 14 des Steuergehäuses 11 ist ein mit dem Ventilschaft 2 des Einlaßventils formschlüssig zusammenwirkender Ventilkolben 15 und über diesem ein Arbeitskolben 16 eines Nockenkolbens 17 axial verschiebbar angeordnet. Der Arbeitskolben 16 ist durch eine Rückstellfeder 18 belastet, die sich einerseits an einer Schulter des Steuergehäuses 11 abstützt und andererseits an einem Flansch des Arbeitskolbens 16 angreift und dadurch den Nockenkolben 17 an den Ventilsteuernocken 4 preßt.The hydraulic valve control device has a control housing 11, which is inserted into a housing bore 10 of the engine valve housing 5 and in which a spring chamber 12 is arranged, the valve closing springs 6 and 7 being accommodated coaxially to one another in the spring chamber 12. In the control housing 11 is from a cup-shaped spring plate 13 anchored and axially displaceable with the valve stem 2 and loaded by the valve closing springs 6 and 7 is inserted below. In a central axially continuous bore 14 of the control housing 11, a valve piston 15, which interacts positively with the valve stem 2 of the inlet valve, and above it a working piston 16 of a cam piston 17 is arranged axially displaceably. The working piston 16 is loaded by a return spring 18, which is supported on the one hand on a shoulder of the control housing 11 and on the other hand engages a flange of the working piston 16 and thereby presses the cam piston 17 against the valve control cams 4.

Zwischen den einander zugewandten Stirnflächen des Ventilkolbens 15 und des Arbeitskolbens 16 ist in der Gehäusebohrung 14 ein mit Öl gefüllter Druckraum 19 eingeschlossen, wobei die wirksame Länge des gesamten Ventilstößels durch die Ölmenge bestimmt wird, die in diesem Druckraum 19 vorhanden ist. Bei Verringern der eingeschlossenen Ölmenge ist der wirksame Öffnungshub des Einlaßventils geringer; bei Aufrechterhalten der maximalen Füllung ist der Hub des Einlaßventils maximal.Between the mutually facing end faces of the valve piston 15 and the working piston 16, an oil-filled pressure chamber 19 is enclosed in the housing bore 14, the effective length of the entire valve tappet being determined by the amount of oil present in this pressure chamber 19. When the amount of oil trapped is reduced, the effective opening stroke of the intake valve is less; if the maximum filling is maintained, the stroke of the intake valve is at a maximum.

Der Druckraum 19 steht über einen Druckkanal 21 mit einem Speicherventil 22 in Verbindung, das einen radial dichtenden topfförmigen Speicherkolben 23 aufweist, der durch eine Speicherfeder 24 belastet in seiner dargestellten Ruhestellung auf einem Ventilsitz 25 aufliegt. Die untere Stirnfläche des Speicherkolbens 23 begrenzt dabei einen Speicherraum 26, während ein Teil der Mantelfläche des Speicherkolbens 23 einen diesen umgebenden Ringkanal 27 abgrenzt, in den der Druckkanal 21 mündet.The pressure chamber 19 is connected via a pressure channel 21 to a storage valve 22 which has a radially sealing cup-shaped storage piston 23 which is loaded by a storage spring 24 its rest position shown rests on a valve seat 25. The lower end face of the storage piston 23 delimits a storage space 26, while part of the outer surface of the storage piston 23 delimits an annular channel 27 surrounding the latter, into which the pressure channel 21 opens.

Die Ventilsteuervorrichtung arbeitet mit einem Hydraulikkreislauf, mit einer Förderpumpe 28, die aus einem Ölbehälter 29 das Steueröl ansaugt und über eine Förderleitung 31 der Ventilsteuervorrichtung zuführt. Zur Erzielung eines bestimmten Förderdrucks ist in einer von der Förderleitung 31 abzweigenden und zum Ölbehälter 29 zurückführenden Leitung 32 ein Drucksteuerventil 33 angeordnet.The valve control device works with a hydraulic circuit, with a feed pump 28, which sucks in the control oil from an oil tank 29 and supplies it to the valve control device via a feed line 31. To achieve a certain delivery pressure, a pressure control valve 33 is arranged in a line 32 branching off from the delivery line 31 and returning to the oil container 29.

Die Förderleitung 31 führt zu einem 2/2-Magnetventil 34, welches eine Steuerleitung 35 steuert, die über ein Rückschlagventil 36 zum Speicherraum 26 führt. An die Förderleitung 31 ist kurz vor dem Magnetventil 34 ein Vordruckspeicher 37 angeschlossen, dessen Speicherdruck mit dem Drucksteuerventil 33 abgestimmt ist und der in der dargestellten Schließstellung des Magnetventils 34 weitgehend mit Steueröl aufgefüllt ist. Von der Steuerleitung 35 zweigen weitere Steuerleitungen 38 ab, die zu weiteren Motorsteuerventileinheiten des selben Motors führen, wobei diese Einheiten entsprechend der dargestellten ausgebildet sind.The delivery line 31 leads to a 2/2-way solenoid valve 34, which controls a control line 35, which leads to the storage space 26 via a check valve 36. A supply pressure accumulator 37 is connected to the delivery line 31 shortly before the solenoid valve 34, the storage pressure of which is coordinated with the pressure control valve 33 and which is largely filled with control oil in the closed position of the solenoid valve 34 shown. Additional control lines 38 branch off from the control line 35 and lead to further engine control valve units of the same engine, these units being designed in accordance with the one shown.

Von der Förderleitung 31 zweigt eine Auffülleitung 39 ab, die zum Druckkanal 21 führt und in der ein zum Druckkanal 21 hin öffnendes Rückschlagventil 41 angeordnet ist.A filling line 39 branches off from the delivery line 31, which leads to the pressure channel 21 and in which a check valve 41 which opens towards the pressure channel 21 is arranged.

In Fig. 2 ist das Speicherventil 22 in vergrößertem Maßstab dargestellt. Im Bereich des Ringkanals 27 weist der Speicherkolben auf seiner Mantelfläche einen Absatz 42 auf, durch den eine in Öffnungsrichtung dieses Ventils wirkende Druckschulter 43 entsteht. Entsprechend ist der Durchmesser des Ventilsitzes 25 kleiner als der Durchmesser des Speicherkolbens 23 in seinem radialen Führungsbereich.2, the storage valve 22 is shown on an enlarged scale. In the area of the annular channel 27, the accumulator piston has a shoulder 42 on its outer surface, which creates a pressure shoulder 43 which acts in the opening direction of this valve. Accordingly, the diameter of the valve seat 25 is smaller than the diameter of the accumulator piston 23 in its radial guide area.

Durch die Speicherfeder 24 wird innerhalb des topfförmig ausgebildeten Speicherkolbens 23 ein Federteller 44 einer schwachen Feder 45 auf den Speicherkolbenboden gespannt, wobei die Feder 45 das bewegliche Ventilglied eines Entlastungsventils 46 belastet, welches in einer Entlastungsleitung 47 angeordnet ist, die den Speicherraum 26 mit dem Speicherfederraum 48 verbindet. Die Entlastungsleitung 47 ist hier als Drosselleitung ausgebildet, so daß sie als Staudrossel für einen Abfluß von Steueröl aus dem Speicherraum 26 zum Speicherfederraum 48 wirkt. Das Entlastungsventil 46 kann zudem als Drucksteuerventil ausgebildet sein, um so im Speicherraum 26 einen bestimmten Vordruck aufrecht zu erhalten.By means of the storage spring 24, a spring plate 44 of a weak spring 45 is tensioned on the storage piston base within the cup-shaped storage piston 23, the spring 45 loading the movable valve member of a relief valve 46, which is arranged in a relief line 47, which connects the storage space 26 with the storage spring space 48 connects. The relief line 47 is designed here as a throttle line, so that it acts as a throttle for an outflow of control oil from the storage space 26 to the storage spring space 48. The relief valve 46 can also be designed as a pressure control valve, so as to maintain a specific admission pressure in the storage space 26.

Die in Fig. 1 und 2 beschriebene Ventilsteuervorrichtung arbeitet wie folgt: Beim Rotieren der Nockenwelle 3 wird über den Antriebsnocken 4 der Nockenkolben 17 mit Arbeitskolben 16 entgegen der Rückstellfeder 18 nach unten verschoben und verdrängt im Druckraum 19 das Hydrauliköl nach unten. Der dabei erzeugte Druck setzt sich einerseits über den Druckkanal 21 zum Speicherventil 22 hin fort wirkt aber vor allem auf die obere Stirnfläche des Ventilkolbens 15, wobei dieser einschließlich dem Ventilschaft 2 mit Ventilteller 1 entgegen der Kraft der Ventilschließfedern 6 und 7 nach unten geschoben wird, wobei der Ventilteller 1 vom Ventilsitz 8 abhebt und die Einlaßöffnung 9 freigibt, so daß entsprechend dem freigegebenen Querschnitt und der zur Verfügung stehenden Öffnungszeit, also entsprechend dem Öffnungszeitquerschnitt Verbrennungsluft in den Brennraum des Motors strömt. Die Öffnung des Einlaßventils erfolgt synchron mit den Saughüben des Motorkolbens, wobei wiederum in Abstimmung mit der Zündfolge bzw. des Kurbeltriebs der Brennkraftmaschine die einzelnen Motorventile nacheinander geöffnet werden, beispielsweise, wenn die nebeneinander angeordneten Motorzylinder mit I bis IV durchnummeriert sind, die Öffnungs- bzw. Zündreihenfolge sein könnte III, IV, II und zuletzt I wonach dann wieder bei einer solchen 4-Zylinder-Brennkraftmaschine das Motorventil von Zylinder III öffnen würde, usw.The valve control device described in FIGS. 1 and 2 works as follows: When the camshaft 3 rotates, the cam piston 17 with the working piston 16 is moved downward against the return spring 18 via the drive cam 4 and displaces the hydraulic oil downward in the pressure chamber 19. The pressure generated thereby continues on the one hand via the pressure channel 21 to the accumulator valve 22, but acts primarily on the upper end face of the valve piston 15, which, including the valve stem 2 with the valve disk 1, is pushed downward against the force of the valve closing springs 6 and 7, wherein the valve plate 1 lifts off the valve seat 8 and releases the inlet opening 9, so that combustion air flows into the combustion chamber of the engine in accordance with the released cross section and the available opening time, that is to say in accordance with the opening time cross section. The opening of the intake valve takes place synchronously with the suction strokes of the engine piston, whereby again the individual engine valves are opened one after the other in coordination with the firing order or the crank drive of the internal combustion engine, for example if the engine cylinders arranged next to one another are numbered I to IV, the opening or Firing order could be III, IV, II and finally I, after which the motor valve of cylinder III would open again in such a 4-cylinder internal combustion engine, etc.

Der vom Druckraum 19 während des Antriebs des Öffnungsventils vorhandene, verhältnismäßig hohe Druck überträgt sich über den Druckkanal 21 in den Ringkanal 27 des Speicherventils 22 und beaufschlagt dort die Druckschulter 43 am Speicherkolben 23 entgegen der Kraft der Speicherfeder 24. Die durch die Fläche der Druckschulter 43 und den Druck im Ringkanal 27 dabei entwickelte Kraft ist jedoch stets kleiner als die Kraft der Speicherfeder 24, so daß der Speicherkolben 23 auf dem Ventilsitz 25 verharrt. Solange das Magnetventil 34 die dargestellte Schließstellung einnimmt, und der Speicherkolben 23 auf seinem Sitz 25 verharrt, legt der Ventilteller 1 einen maximalen Öffnungshub zurück, da das vom Arbeitskolben 16 verdrängte Hydrauliköl mangels sonstiger Ausweichmöglichkeit den Ventilkolben 15 so weit nach unten verschiebt wie der Arbeitskolben 16 verschoben wird, wobei der dabei zurückgelegte Weg unmittelbar der Höhe des Antriebsnockens 4 entspricht.The existing from the pressure chamber 19 during the drive of the opening valve, relatively high Pressure is transferred via the pressure channel 21 into the ring channel 27 of the storage valve 22 and acts there on the pressure shoulder 43 on the storage piston 23 against the force of the storage spring 24. However, the force developed by the surface of the pressure shoulder 43 and the pressure in the ring channel 27 is always smaller than the force of the storage spring 24, so that the storage piston 23 remains on the valve seat 25. As long as the solenoid valve 34 assumes the closed position shown, and the accumulator piston 23 remains in its seat 25, the valve disk 1 travels a maximum opening stroke, since the hydraulic oil displaced by the working piston 16 displaces the valve piston 15 downward as far as the working piston 16 due to the lack of any other possibility of evasion is displaced, the distance covered directly corresponding to the height of the drive cam 4.

In der Zeichnung ist die Motorventilsteuerung gerade in einer Antriebspause dargestellt, d. h. in einer Arbeitsstellung, in der der Grundkreis des Nockens 4 mit dem Nockenkolben 17 zusammenwirkt und wobei der Ventilteller 1 des Einlaßventils auf seinem Ventilsitz 8 durch die Ventilschließfedern 6 und 7 angetrieben dichtend aufliegt. Irgendwelche während des Betriebs entstehende Leckverluste von Hydrauliköl im Druckraum 19 werden über die Auffülleitung 39 ausgeglichen, über die Hydrauliköl unter Förderdruck über das Rückschlagventil 41 in den Druckkanal 21 und damit in den Druckraum 19 strömen kann. Hierdurch wird im Druckraum 19 ein während der Antriebspausen stets gleicher Vordruck erzeugt und es werden außerdem Hohlräume vermieden, die zu Steuerfehlern in Bezug auf den Öffnungszeitpunkt aber auch den Öffnungshub des Motorventils führen könnten.In the drawing, the engine valve control is shown in a drive break, ie in a working position in which the base circle of the cam 4 interacts with the cam piston 17 and the valve disk 1 of the intake valve rests sealingly on its valve seat 8 driven by the valve closing springs 6 and 7. Any leakage losses of hydraulic oil in the pressure chamber 19 that occur during operation are compensated for via the fill line 39, via the hydraulic oil under delivery pressure via the check valve 41 into the pressure channel 21 and so that it can flow into the pressure chamber 19. As a result, a pre-pressure which is always the same during the drive breaks is generated in the pressure chamber 19 and cavities are also avoided, which could lead to control errors with respect to the opening time but also the opening stroke of the engine valve.

Sobald das Magnetventil 34 umgeschaltet wird, wird von der Förderleitung 31 her der im Vordruckspeicher 37 herrschende Förderdruck über die Steuerleitung 35 und das Rückschlagventil 36 in den Speicherraum 26 übertragen, so daß die untere Stirnfläche des Speicherkolbens 23 von einem Steuerdruck beaufschlagt ist, der nur geringfügig niedriger ist als der Förderdruck in der Förderleitung 31. Dieser Steuerdruck erzeugt in Bezug auf die beaufschlagte Stirnfläche eine am Speicherkolben in Öffnungsrichtung wirkende Kraft, die geringer ist als die Kraft der Speicherfeder 24. Auch wenn zu dieser Steuerkraft die Vordruckkraft hinzukommt, die von der Ringschulter 43 des Speicherkolbens 23 ausgeht und stets vorhanden ist, solange im Druckraum 19 der konstante Vordruck herrscht, reicht dieses nicht aus, um die Kraft der Speicherfeder 24 zu überwinden. Erst wenn der Antriebsnocken 4 wirksam wird und den Arbeitskolben 16 betätigt, entsteht im Druckraum 19 ein verhältnismäßig hoher Arbeitsdruck, wodurch auch die aufgrund der Schulter 43 am Kolben 23 angreifende Kraft entsprechend ansteigt, wird die Kraft der Speicherfeder 24 überwunden und der Speicherkolben 23 nach oben geschoben, wobei er aus seiner Ruhelage vom Ventilsitz 25 abhebt, so daß das Hydrauliköl vom Druckraum 19 her über den Druckkanal 21 in den Speicherraum 26 strömen kann, wobei der Arbeitsdruck des Hydrauliköls nach Abheben des Speicherkolbens 23 vom Ventilsitz 25 die ganze untere Stirnfläche beaufschlagt und damit eine hohe die Kraft der Speicherfeder 24 überwindende Verstellkraft bewirkt. Demnach ist es Voraussetzung, daß der Antriebsnocken 4 wirksam ist, d. h., daß eine Aufsteuerung des Motorventils stattfindet, wenn der Speicherkolben 23 verschoben werden soll. Dadurch jedoch, daß vom Druckraum 19 ein Teil der verdrängten Menge zum Speicherraum 26 strömt, wird der Öffnungshub des Motorventils entsprechend verkleinert, wodurch auch der Öffnungszeitquerschnitt verringert wird. Eine solche Änderung des Öffnungszeitquerschnitts wirkt sich auf das angesaugte Luftvolumen des Motors aus und damit unmittelbar auf die Drehzahl des Motors. Um ein gewisses Öffnen des Motorventils in jedem Fall zu gewährleisten, wird das Magnetventil 34 erst dann umgesteuert, wenn bereits der Öffnungshub des Motorventils begonnen hat, d. h., wenn durch den Antriebsnocken 4 bereits eine Verschiebung des Arbeitskolbens 16 begonnen hat.As soon as the solenoid valve 34 is switched over, the delivery pressure prevailing in the pre-pressure accumulator 37 is transmitted from the delivery line 31 via the control line 35 and the check valve 36 into the storage space 26, so that the lower end face of the storage piston 23 is acted upon by a control pressure which is only slight is lower than the delivery pressure in the delivery line 31. This control pressure generates, in relation to the end face acted upon, a force acting on the accumulator piston in the opening direction, which is less than the force of the accumulator spring 24. Even if the admission pressure force added by the annular shoulder is added to this control force 43 of the accumulator piston 23 goes out and is always present, as long as the constant pre-pressure prevails in the pressure chamber 19, this is not sufficient to overcome the force of the accumulator spring 24. Only when the drive cam 4 takes effect and actuates the working piston 16 does a relatively high working pressure arise in the pressure chamber 19, as a result of which the force acting on the piston 23 due to the shoulder 43 also increases accordingly, the force of the storage spring 24 is overcome and the storage piston 23 moves upward pushed, whereby it lifts from its rest position from the valve seat 25 so that the hydraulic oil can flow from the pressure chamber 19 via the pressure channel 21 into the storage chamber 26, the working pressure of the hydraulic oil acting on the entire lower end face after lifting the storage piston 23 from the valve seat 25 and thus one high adjusting force overcoming the force of the storage spring 24. Accordingly, it is a prerequisite that the drive cam 4 is effective, ie that the motor valve is opened when the storage piston 23 is to be displaced. However, because part of the displaced quantity flows from the pressure chamber 19 to the storage chamber 26, the opening stroke of the engine valve is correspondingly reduced, as a result of which the opening time cross section is also reduced. Such a change in the opening time cross-section has an effect on the intake air volume of the engine and thus directly on the speed of the engine. In order to ensure a certain opening of the engine valve in any case, the solenoid valve 34 is only reversed when the opening stroke of the engine valve has already started, that is, when the drive cam 4 has already started to displace the working piston 16.

Gleichzeitig mit der Steuerleitung 35 werden auch die Steuerleitungen 38 mit Hydrauliköl unter Steuerdruck versorgt, so daß außer dem dargestellten Speicherkolben 23 auch eine Reihe zu anderen Motorventilsteuerungen des gleichen Motors gehörende Speicherkolben mit Hydrauliköl unter Steuerdruck beaufschlagt werden. Damit im Falle dieses Umschalten des Magnetventils 34 ein ausreichender Steuerdruck in all den Steuerleitungen erhalten bleibt, dient der Speicher 37 dessen Speichervolumen entsprechend ausgelegt ist. Während sich der Speicher in den Zeiten, in denen das Magnetventil 34 geschlossen ist, auflädt, so daß sein Vordruckspeicherkolben 49 die dargestellte Stellung einnimmt, verschiebt sich bei geöffnetem Magnetventil 34 dieser Vordruckspeicherkolben weiter nach unten, beispielsweise in die gestrichelt dargestellte Stellung. Hierdurch kann die Maximalleistung der Förderpumpe 28 entsprechend geringer gehalten werden und es wird außerdem kurzfristig eine hohe Fördermenge zur Verfügung gestellt, so daß eine Art Druckstoß auf die jeweils beaufschlagten Speicherkolben 23 stattfindet. Wie oben beschrieben sind die dabei angreifenden Kräfte von Steuerdruck, Vordruck und Federn so auf einander abgestimmt, daß nur die Speicherkolben 23 von ihrem Sitz 25 abheben, die zusätzlich auf ihrer Druckschulter 43 vom Arbeitsdruck beaufschlagt werden, der nur dann auftreten kann, wenn der Arbeitsnocken 4 auf den Arbeitskolben 16 wirkt. So wie über die Auffülleitung 39 im Druckkanal 21 ein konstanter Vordruck in den Arbeitspausen erzeugt wird, in denen der Antriebsnocken nicht wirksam ist, so wird über das Entlastungsventil 46 in Verbindung mit deren Feder 45 im Speicherraum 26 ein ausreichender Auffülldruck aufrechterhalten. Sobald über die Steuerleitung 35 nach Öffnen des Magnetventils 34 Hydrauliköl unter Steuerdruck in den Speicherraum 26 strömt, wird auch in diesem Speicherraum 26 der Steuerdruck eingestellt, der jedoch höher ist als der Aufsteuerdruck des Entlastungsventils 46, so daß dieses öffnet. Da die Entlastungsleitung 47 im Speicherkolbenboden als Drosselleitung ausgebildet ist, entsteht dadurch ein Stau, so daß sich im Speicherraum 26 der Steuerdruck aufrecht erhalten kann. In jedem Fall ist die Förderleistung der Förderpumpe 28 größer als die über alle gleichzeitig angeschlossenen Speicherräume 26 und deren Entlastungsleitungen 47 abströmende Menge an Hydrauliköl. Sobald dann der Arbeitsdruck vom Druckkanal 21 auf die Druckschulter 43 des Speicherkolbens 23 hinzukommt, hebt dieser Speicherkolben 23 vom Sitz 25 ab und das Rückschlagventil 36 wird durch den Arbeitsdruck gesperrt, der weit höher ist als der Steuerdruck in der Steuerleitung 35.Simultaneously with the control line 35, the control lines 38 are also supplied with hydraulic oil under control pressure, so that, in addition to the storage piston 23 shown, a number of storage pistons belonging to other engine valve controls of the same engine with hydraulic oil under control pressure. In order that a sufficient control pressure is maintained in all the control lines in the event of this switching of the solenoid valve 34, the store 37 serves the purpose of which the store volume is designed accordingly. While the accumulator is charging in the times in which the solenoid valve 34 is closed, so that its pre-pressure accumulator piston 49 assumes the position shown, when the solenoid valve 34 is open, this pre-pressure accumulator piston shifts further down, for example into the position shown in broken lines. As a result, the maximum output of the feed pump 28 can be kept correspondingly lower and, in addition, a high delivery rate is made available for a short time, so that a kind of pressure surge takes place on the respective loaded piston 23. As described above, the attacking forces of control pressure, admission pressure and springs are coordinated with one another in such a way that only the accumulator pistons 23 lift off their seat 25, which are additionally acted upon on their pressure shoulder 43 by the working pressure, which can only occur when the working cam 4 acts on the working piston 16. Just as a constant pre-pressure is generated in the work breaks during which the drive cam is not effective via the fill line 39 in the pressure channel 21, a sufficient fill pressure is maintained in the storage space 26 via the relief valve 46 in connection with its spring 45. As soon as via the control line 35 after opening the solenoid valve 34 Hydraulic oil flows under control pressure in the storage space 26, the control pressure is also set in this storage space 26, which is, however, higher than the pilot pressure of the relief valve 46 so that it opens. Since the relief line 47 is formed in the accumulator piston bottom as a throttle line, this creates a jam so that the control pressure can be maintained in the accumulator chamber 26. In any case, the delivery rate of the delivery pump 28 is greater than the amount of hydraulic oil flowing out over all simultaneously connected storage spaces 26 and their relief lines 47. As soon as the working pressure from the pressure channel 21 is added to the pressure shoulder 43 of the accumulator piston 23, this accumulator piston 23 lifts off the seat 25 and the check valve 36 is blocked by the working pressure which is far higher than the control pressure in the control line 35.

Der Speicherkolben 23 wird, wenn er einmal durch den Arbeitsdruck vom Druckkanal her beaufschlagt ist, entgegen der Kraft der Speicherfeder 24 verschoben. Ab dem Augenblick, ab dem der Speicherraum 26 zum Druckkanal 21 hin geöffnet ist, wird das unter Arbeitsdruck vom Arbeitskolben 16 verdrängte Hydrauliköl in diesen Speicherraum 26 verdrängt, so daß das Einlaßventil wieder zu schließen beginnt, wobei der Ventilschaft 2 mit dem Ventilkolben 15 nach oben geschoben wird und trotz der fortgesetzten Förderwirkung des Arbeitskolbens 16 Hydrauliköl aus dem Druckraum 19 in den Speicherraum 26 fördert. Hierdurch wird der Öffnungshub des Ventiltellers 1 verkürzt und damit auch der Öffnungszeitquerschnitt dieses Einlaßventils, wobei der Öffnungszeitquerschnitt vom Hub aber auch von der Drehzahl bestimmt wird. Während dieses Schließvorgangs des Ventiltellers 1 strömt eine gewisse Menge über die Entlastungsleitung 47 und das Entlastungsventil 46 ab, wobei jedoch diese Menge äußerst gering ist und damit nach vorheriger Berücksichtigung einer solchen Abflußmenge keine nachteilige Wirkung auf die Steuerung hat.The accumulator piston 23, once it is acted upon by the working pressure from the pressure channel, is displaced against the force of the accumulator spring 24. From the moment the storage space 26 is opened to the pressure channel 21, the hydraulic oil displaced by the working piston 16 under working pressure is displaced into this storage space 26, so that the inlet valve begins to close again, the valve stem 2 with the valve piston 15 upwards is pushed and despite the continued delivery of the working piston 16 hydraulic oil from the pressure chamber 19 promotes in the storage space 26. This shortens the opening stroke of the valve disk 1 and thus also the opening time cross section of this inlet valve, the opening time cross section being determined by the stroke but also by the speed. During this closing process of the valve plate 1, a certain amount flows out via the relief line 47 and the relief valve 46, but this amount is extremely small and therefore, after taking into account such a flow rate, has no adverse effect on the control.

Sobald der Antriebsnocken 4 mit seiner Ablaufflanke wirksam wird und der Arbeitskolben 16 wieder nach oben bewegt wird, kann, sobald auch der Ventilteller 1 wieder auf dem Ventilsitz 8 aufliegt, der Speicherkolben 23 mit seinem Rückhub beginnen, wobei er das vorher aufgenommene Hydrauliköl wieder zurück in den sich nunmehr vergrößernden Druckraum 19 fördert. Diese Rückförderung findet auch dann statt, wenn das Magnetventil 34 noch geöffnet ist, da der durch die Speicherfeder 24 bewirkte Verdrängungsdruck des Speicherkolbens 23 weit höher ist als der Steuerdruck von der Förderleitung 31 her. Erst wenn der Speicherkolben 23 auf seinem Sitz 25 aufliegt, kann über das Rückschlagventil 36 und das Entlastungsventil 46 eine Strömung über die Steuerleitung 35 durch den Speicherraum 26 stattfinden, ohne einen Einfluß auf die Steuerung zu haben. Der Vorteil jedoch besteht darin, daß der Zeitpunkt des Öffnens des Magnetventils 34 das Schließen des Motorventils einleitet, wobei die weitere Schließbewegung des Motorventils durch die Ventilschließfedern 6 und 7 bewirkt - abgesehen von den Drücken in der Brennkammer selbst, die den Ventilteller 1 beaufschlagen - von der Ausweichgeschwindigkeit des Speicherkolbens 23 bestimmt wird.As soon as the drive cam 4 takes effect with its trailing edge and the working piston 16 is moved upwards again, as soon as the valve plate 1 again rests on the valve seat 8, the storage piston 23 can begin its return stroke, in which case the hydraulic oil previously taken in again promotes the now expanding pressure chamber 19. This return flow also takes place when the solenoid valve 34 is still open, since the displacement pressure of the storage piston 23 caused by the storage spring 24 is far higher than the control pressure from the delivery line 31. Only when the accumulator piston 23 rests on its seat 25 can a flow via the control line 35 through the accumulator chamber 26 take place via the check valve 36 and the relief valve 46, without having any influence on the control. However, the advantage is that the timing of opening the solenoid valve 34 initiates the closing of the engine valve, with the further closing movement of the engine valve caused by the valve closing springs 6 and 7 - apart from the pressures in the combustion chamber itself which act on the valve plate 1 - being determined by the evasion speed of the accumulator piston 23.

In Fig. 3 ist anhand von drei übereinander dargestellten Diagrammen der Arbeitshubverlauf des Ventils für drei verschiedene Drehzahlen dargestellt. In den Diagrammen ist über dem Drehwinkelgrad der Kurbelwelle als °KW (Abzisse) der Hub des Motorventils h (Ordinate) dargesteltt. Das erste Diagramm a ist für eine Motordrehzahl von 1000 U/min; das zweite Diagramm b entspricht einer Drehzahl 3000 U/min und das untereste Diagramm c gilt für eine Drehzahl von 5000 U/min. Die äußere Mantelkurve in allen drei Diagrammen entspricht dem Einlaßventil Öffnungs- und Schließvorgang ohne Einfluß der Steuerung über das Magnetventil 34. Die strichpunktiert dargestellte Kurvenschar in jedem Diagramm entspricht wiederum einer Verkürzuung des Öffnungshubs oder der Öffnungszeit durch die Wirkung des Magnetventils 34, d. h. durch Öffnen desselben und wirksamwerden des Speicherventils 22. Während der Verlauf des Öffnungsabschnitts der Kurven bei allen Kurven gleich ist, ist der Schließverlauf unterschiedlich. Der Öffnungshubabschnitt der Kurve wird allein durch den Antriebsnocken 4 bestimmt, der immer die gleiche Öffnungswirkung auf das Motorventil hat. Dies gilt auch für die Schließwirkung entsprechend der Ablaufbahn des Antriebsnockens 4. Sobald jedoch das Magnetventil 34 geöffnet hat, wird der den Motorventil schließen entsprechende Abschnitt der Kurve durch die oben beschriebenen Einflüsse bestimmt, vor allem durch die Wirkung des Speicherkolbens 23.In Fig. 3, the working stroke of the valve for three different speeds is shown on the basis of three diagrams. In the diagrams, the stroke of the engine valve h (ordinate) is shown above the degree of rotation of the crankshaft as ° KW (abscissa). The first diagram a is for an engine speed of 1000 rpm; the second diagram b corresponds to a speed of 3000 rpm and the lowest diagram c applies to a speed of 5000 rpm. The outer jacket curve in all three diagrams corresponds to the opening and closing of the inlet valve without the influence of the control via the solenoid valve 34. The group of curves shown in dash-dotted lines in each diagram in turn corresponds to a shortening of the opening stroke or the opening time due to the action of the solenoid valve 34, that is to say by opening the same and take effect of the storage valve 22. While the course of the opening portion of the curves is the same for all curves, the closing course is different. The opening stroke section of the curve is determined solely by the drive cam 4, which always has the same opening effect on the engine valve. This also applies to the Closing action corresponding to the trajectory of the drive cam 4. However, as soon as the solenoid valve 34 has opened, the section of the curve corresponding to the closing of the motor valve is determined by the influences described above, above all by the action of the accumulator piston 23.

Ein Vergleich der drei Diagramme ergibt, daß je höher die Motordrehzahl ist desto flacher verlaufen diese strichpunktiert dargestellten Schließkurven. Je höher die Drehzahl ist desto mehr nähert sich die Steigung der strichpunktiert dargestellten Schließkurven jener Steigung der Mantelkurve an, deren Verlauf durch den Antriebsnocken bestimmt wird. Dieser Effekt beruht darauf, daß aufgrund der Zunahme der Massenkräfte mit steigender Drehzahl das mittlere Druckniveau im Druckraum 19 absinkt, wodurch die Ausgleichsbewegung des Speicherkolbens 23 langsamer erfolgt. Diese Eigenschaft von dieser Art hydraulischer Motorventilsteuerung beeinflußt nicht unerheblich den tatsächlichen Aufsteuerzeitquerschnitt des Einlaßventils. Die durchgezogene Mantelkurve in den Diagrammen, die in ihrem Verlauf jenem des Antriebsnockens 4 entspricht, ist für eine maximale Leistung bei hoher Motordrehzahl angepaßt.A comparison of the three diagrams shows that the higher the engine speed, the flatter the closing curves shown in dash-dot lines. The higher the speed, the more the slope of the closing curves shown in dash-dot lines approaches the slope of the jacket curve, the course of which is determined by the drive cam. This effect is based on the fact that, due to the increase in inertial forces, the mean pressure level in the pressure chamber 19 drops with increasing speed, as a result of which the compensating movement of the accumulator piston 23 takes place more slowly. This property of this type of hydraulic engine valve control has a significant influence on the actual opening time cross section of the intake valve. The solid jacket curve in the diagrams, which corresponds in its course to that of the drive cam 4, is adapted for maximum performance at high engine speed.

Wie aus dem Diagramm c erkennbar ist, wirkt sich eine über das Magnetventil 34 bei 180° KW durchgeführte Ventilsteuerung nicht mehr aus, da bei diesen hohen Drehzahlen der Einlaßschluß mit jenem bei 240° KW zusammenfallen würde, wie er ohne Steuerung ohnehin stattfindet. Das heißt also, daß bei maximaler Drehzahl und Vollast, d. h. bei der Forderung nach einem maximalen Zeitquerschnitt ab 180° KW eine elektrische Steuerung des Motorventils über das Magnetventil 34 uninteressant ist und demnach nicht erforderlich ist. Zeitquerschnittssteuerungen bei Höchstdrehzahl und bei nierdrigerer Last bzw. Leistung wird dadurch gesteuert, daß unterhalb von 180° KW das Magnetventil 34 entsprechend eingeschaltet wird. Bei niedrigeren Drehzahlen hingegen könnte sich theoretisch eine Steuerung oberhalb von 180° KW noch auswirken, nur ist sie gerade dort nicht erforderlich. Im Bereich bis zu 3000 U/min findet im Normalfall das Schließen des Einlaßventils bei 180° KW statt, um die dort erforderliche maximale Leistungsausbeute zu erhalten. Beim Diagramm a entspricht das einem Umschalten des Magnetventils 34 bei etwa 160° KW und bei 3000 U/min entsprechend Diagramm b bei 130 ° KW.As can be seen from the diagram c, a valve control carried out via the solenoid valve 34 at 180 ° KW no longer has an effect, since at these high speeds the inlet closure with that at 240 ° KW would coincide with how it takes place without control anyway. This means that at maximum speed and full load, ie when a maximum time cross section from 180 ° KW is required, electrical control of the engine valve via the solenoid valve 34 is uninteresting and is therefore not necessary. Time cross-section controls at maximum speed and with low load or power is controlled in that the solenoid valve 34 is switched on accordingly below 180 ° KW. At lower speeds, on the other hand, a control system above 180 ° KW could theoretically have an effect, only it is not necessary there. In the range up to 3000 rpm, the inlet valve is normally closed at 180 ° KW in order to obtain the maximum power yield required there. In diagram a, this corresponds to switching the solenoid valve 34 at approximately 160 ° KW and at 3000 rpm according to diagram b at 130 ° KW.

Bei Mehrzylinder-Brennkraftmaschinen finden bekanntlich pro Umdrehung der Kurbelwelle mehrere Explosionshübe statt, beispielsweise bei einer Vier-Zylinder-Brennkraftmaschine zwei derartige Explosionshübe. Bei einer 4-Zylinder-Brennkraftmaschine finden somit die vier Explosionshübe der vier Zylinder innerhalb von zwei Umdrehungen statt. Eine übliche Zündfolge ist beispielsweise die gemäß der nebeneinander angeordneten Motorzylinder drei - vier - zwei - eins. Da wie in Fig. 3 dargestellt die Aufsteuerung eines Motorventils über 240° KW gehen kann, ergibt sich dadurch eine Überschneidung jeweils ab 180° KW. Solange die einzelnen Motorventile unabhängig von einander gesteuert sind, spielt dieses keine Rolle. Gemäß einer Ausgestaltung der Erfindung sollen jedoch mehrere Motorventile mit nur einem Magnetventil 34 gesteuert werden. Dadurch, daß wie oben erläutert, der Bereich oberhalb von 180° KW für die Steuerung uninteressant ist, wird erfindungsgemäß die Steuerung so ausgelegt, daß nur bis 180° KW das Magnetventil 34 aufgesteuert wird. Hierdurch ist es theoretisch möglich, mit nur einem Magnetventil alle vier Motorzylinder zu steuern, in dem pro Umdrehung mindestens zweimal das Magnetvnetil aufgesteuert wird und wobei nur immer das Speicherventil 22 von seinem Sitz 25 abhebt, bei dessen zugeordneten Motorventil gerade der Nocken 4 wirksam ist. Da in den verbleibenden 180° bis 240° KW keine Steuerung erforderlich ist, kann eine Überschneidung gar nicht stattfinden. Bei manchen Motoren, die eine höhere Zylinderzahl haben oder auch einen größeren Einlaßschlußwinkel nach dem unteren Totpunkt des Motorventils ist es sinnvoll, bei einem 4-Zylinder-Motor zwei Motorventilsteuereinheiten über je ein Magnetventil zu schalten. So sind nach der Variante in Fig. 4 dem Magnetventil 34 und entsprechender Aufzweigung der Steuerleitung 35 in jeder der zwei Steuerleitungen 38 wiederum ein 2/2-Magnetventil 51 angeordnet, wobei stromab dieser Magnetventile 51, die Steuerleitungen 38 weiter aufgezweigt sind, um zu den einzelnen Motorventilsteuereinheiten zu führen.In multi-cylinder internal combustion engines, as is known, several explosion strokes take place per revolution of the crankshaft, for example two such explosion strokes in a four-cylinder internal combustion engine. In a 4-cylinder internal combustion engine, the four explosion strokes of the four cylinders take place within two revolutions. A common ignition sequence is, for example, three - four - two - one according to the engine cylinders arranged next to one another. Since, as shown in Fig. 3, the opening of a Motor valve can go over 240 ° KW, this results in an overlap from 180 ° KW. As long as the individual engine valves are controlled independently of each other, this does not matter. According to an embodiment of the invention, however, several engine valves are to be controlled with only one solenoid valve 34. Because, as explained above, the area above 180 ° KW is of no interest for the control, the control is designed according to the invention in such a way that the solenoid valve 34 is opened only up to 180 ° KW. This makes it theoretically possible to control all four motor cylinders with only one solenoid valve, in which the solenoid valve is turned on at least twice per revolution and only the storage valve 22 always lifts off its seat 25, the cam 4 of which is currently active in the associated motor valve. Since no control is required in the remaining 180 ° to 240 ° KW, there can be no overlap. With some engines that have a higher number of cylinders or also a larger intake closing angle after the bottom dead center of the engine valve, it makes sense to switch two engine valve control units via a solenoid valve in each case with a 4-cylinder engine. Thus, according to the variant in FIG. 4, the solenoid valve 34 and corresponding branching of the control line 35 in each of the two control lines 38 are in turn arranged a 2/2 solenoid valve 51, the control lines 38 being branched further downstream of these solenoid valves 51 are to lead to the individual engine valve control units.

In Fig. 5 mündet die Steuerleitung 35 des Magnetventils 34 in den Eingang eines 3/2-Magnetventils 52, dessen Ausgänge wiederum zu den Steuerleitungen 38 führen, die in sich dann wieder aufgezweigt zu den einzelnen Motorventilsteuereinheiten führen.In Fig. 5, the control line 35 of the solenoid valve 34 opens into the input of a 3/2-way solenoid valve 52, the outputs of which in turn lead to the control lines 38, which then branch out again to the individual engine valve control units.

In Fig. 6 ist eine Variante für die Steuerung der Auffülleitung 39 dargestellt, wobei die Mündung der Auffülleitung 39 stromab des Rückschlagventils 42 durch den Speicherkolben 23 erfolgt. Die Auffülleitung 39 mündet hierfür in eine Ringnut 53 in der Bohrungswand, in der der Speicherkolben 23 radial dichtend geführt ist, wobei diese Ringnut 53 über eine in der Länge begrenzte Längsnut 54 in der dargestellten Ruhelage des Speicherkolbens 23 mit dem Druckkanal 21 verbunden ist. Hierdurch kann in dieser Ruhelage des Speicherkolbens 23 eine ungehinderte Auffüllung des Druckkanals 21 und damit des Druckraums erfolgen. Sobald dann der Speicherkolben 23 von seinem Sitz abgehoben hat, wird die Längsnut 54 durch das Verschieben des Speicherkolbens 23 von dem Druckkanal 21 getrennt, so daß in einer solchen Verschiebestellung kein Hydrauliköl von der Auffülleitung 39 her in den Druckkanal 21 gelangen kann. Hierdurch kann die Druckbilanz im Steuersystem verfeinert werden, so daß auch bei hoher Drehzahl und einem entsprechend geringeren Arbeitsdruck keine Fehlsteuerungen dadurch stattfinden, daß das Speicherventil 22 ungewünscht öffnet. Die Kraft, die durch die Speicherfeder 24 am Speicherkolben 23 angreift, kann dann geringer sein als die an dem Speicherkolben 23 in Öffnungsrichtung angreifende, die durch den Vordruck bewirkt wird, wenn er die gesamte Stirnfläche beaufschlagt. Sobald jedoch der Speicherkolben 23 auf dem Sitz 25 aufliegt, kann sich in dem Speicherraum 26 nur noch ein Druck einstellen, wie er durch das Entlastungsventil 46 bestimmt wird und der in jedem Fall wesentlich niedriger ist als der Steuerdruck bzw. konstante Vordruck im Druckkanal 21. Diese Situation ändert sich natürlich, wenn das Magnetventil 34 umschaltet und über die Steuerleitung 35 Hydrauliköl unter Steuerdruck in den Speicherraum 26 strömt und den Speicherkolben 23 vom Sitz 25 abhebt, so daß wiederum die ganze Stirnfläche vom Arbeitsdruck beaufschlagbar ist.6 shows a variant for the control of the filling line 39, the opening of the filling line 39 taking place downstream of the check valve 42 through the storage piston 23. For this purpose, the filling line 39 opens into an annular groove 53 in the bore wall, in which the storage piston 23 is guided in a radially sealing manner, this annular groove 53 being connected to the pressure channel 21 via a longitudinal groove 54 of limited length in the illustrated rest position of the storage piston 23. As a result, in this rest position of the accumulator piston 23, the pressure channel 21 and thus the pressure chamber can be filled without hindrance. As soon as the accumulator piston 23 has lifted from its seat, the longitudinal groove 54 is separated from the pressure channel 21 by the displacement of the accumulator piston 23, so that in such a shifted position no hydraulic oil can get into the pressure channel 21 from the filling line 39. As a result, the pressure balance in the control system can be refined, so that there are no incorrect controls even at high speed and a correspondingly lower working pressure take place in that the storage valve 22 opens unwanted. The force which acts on the accumulator piston 23 by the accumulator spring 24 can then be less than that which acts on the accumulator piston 23 in the opening direction, which is caused by the admission pressure when it acts on the entire end face. However, as soon as the accumulator piston 23 rests on the seat 25, only a pressure can be set in the accumulator chamber 26 as determined by the relief valve 46 and which is in any case significantly lower than the control pressure or constant admission pressure in the pressure channel 21. Of course, this situation changes when the solenoid valve 34 switches over and hydraulic oil flows under control pressure into the storage space 26 via the control line 35 and lifts the storage piston 23 from the seat 25, so that the entire end face can again be acted upon by the working pressure.

Claims (13)

  1. Hydraulic valve control appliance for internal combustion engines
    - having an engine valve (1, 2) axially driven by the drive cam (4) of an engine camshaft (3) by means of a cam piston (16), a pressure space (19), which is adjacent to the cam piston (16) and is filled with hydraulic fluid via a filling conduit (39) containing a non-return valve (36) [sic], and a valve piston (15),
    - having a reservoir, which is configured as a reservoir valve (22) whose valve element is a reservoir piston (23), which bounds a reservoir space (26), is loaded by a reservoir spring (24) and by means of which the pressure space (19) can be connected to the reservoir space (26) via a pressure conduit (21)
    - and having a magnetic valve (34), which is activated by means of an electronic control unit which processes engine parameters, for controlling the reservoir piston (23) and therefore for changing the effective length of the pressure space (19), which determines the stroke of the engine valve (1, 2), by withdrawing hydraulic fluid quantities into the reservoir,
    characterized
    - in that a control conduit (35) for hydraulic fluid brought to a certain control pressure opens into the reservoir space (26), which is separated from the pressure space (19) by the reservoir piston (23) in its closed position, in which control conduit are arranged the magnetic valve (34) and a non-return valve opening in the direction of the reservoir space (26),
    - in that there is a pressure surface (43) on the reservoir piston (23), which pressure surface (43) acts against the force of the reservoir spring (24) and is continually acted on by the pressure in the pressure passage (21)
    - and in that the spring force of the reservoir spring (24) acting on the reservoir piston is greater than the control force acting on the reservoir piston (23) due to the pressure present in the reservoir space (26) when the latter is separated from the pressure space (19) with the magnetic valve (34) closed and the force, which results from the action of pressure on the pressure surface (43) due to the pressure (admission pressure) present in the pressure space (19) before the beginning of the motion of the cam piston (16) in the opening direction of the engine valve (1, 2), but is, however, smaller than the force which is generated when the pressure surface of the reservoir piston (23) is acted on by the pressure in the pressure space (19) when the cam piston (16) is actuated by the drive cam (4) in the opening direction of the engine valve against a valve closing spring (6, 7) acting on the engine valve.
  2. Valve control appliance according to Claim 1, characterized in that a relief conduit (47), which contains a back-pressure throttle, branches off from the reservoir space (26).
  3. Valve control appliance according to Claim 2, characterized in that the relief conduit (47) is controlled by a pressure retention valve (46).
  4. Valve control appliance according to Claim 2 or 3, characterized in that the relief conduit (47) is arranged in the bottom of the reservoir piston (23) and connects the reservoir space (26) to the reservoir spring space (48).
  5. Valve control appliance according to one of the preceding claims, characterized in that the control conduit (35) is connected to a control pressure reservoir (37) upstream of the magnetic valve (34).
  6. Valve control appliance according to one of the preceding claims, characterized in that the magnetic valve (34) is configured as a 2/2-way magnetic valve.
  7. Valve control appliance according to one of Claims 3 to 6, characterized in that the force of the reservoir spring (24) is less than the opening force formed by the control pressure and the bottom of the reservoir piston with the magnetic valve (34) open and in that the filling conduit (39) is controlled by the reservoir piston (23), the filling conduit (39) being shut off after the connection is made between the pressure passage (21) and the reservoir space (26) (Fig. 6).
  8. Valve control appliance according to Claim 7, characterized in that there is a longitudinal groove (54) on the outer surface of the reservoir piston (23), which longitudinal groove (54) has a continual overlap with an annular groove (53) present in the hole accommodating the reservoir piston (23) and is connected, in the rest position, to the pressure passage (21) but is separated from the pressure passage (21) after the reservoir piston (23) is displaced against the force of the reservoir spring (24).
  9. Valve control appliance according to one of the preceding claims, characterized in that the filling conduit (39) is connected to a supply conduit (31) under control pressure.
  10. Valve control appliance according to one of the preceding claims for a multi-cylinder internal combustion engine, characterized in that by means of an electronic control unit, the individual valve control units (magnetic valves (34)) are respectively controllable only up to an actuation of 180° of crankshaft rotational angle (°KW), so that a plurality of valve control units are controlled by only one magnetic valve (34), overlaps of control periods (switch-on times of the magnetic valve (34)) being prevented above 180° KW per valve.
  11. Valve control appliance according to Claim 10, characterized in that groups of valve control units are controllable by at least one preselection valve 51, 52 after a first division of the control conduit downstream of the magnetic valve (34) (Fig. 4 or 5).
  12. Valve control appliance according to Claim 11, characterized in that the preselection valve is configured as a 2/2-way valve 51.
  13. Valve control appliance according to Claim 11, characterized in that a 3/2-way valve simultaneously effecting a division is used as the preselection valve.
EP90915296A 1989-11-25 1990-10-26 Hydraulic control device for valves of internal combustion engines Expired - Lifetime EP0455761B1 (en)

Applications Claiming Priority (2)

Application Number Priority Date Filing Date Title
DE3939065 1989-11-25
DE3939065A DE3939065A1 (en) 1989-11-25 1989-11-25 HYDRAULIC VALVE CONTROL DEVICE FOR INTERNAL COMBUSTION ENGINES

Publications (2)

Publication Number Publication Date
EP0455761A1 EP0455761A1 (en) 1991-11-13
EP0455761B1 true EP0455761B1 (en) 1993-12-29

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EP90915296A Expired - Lifetime EP0455761B1 (en) 1989-11-25 1990-10-26 Hydraulic control device for valves of internal combustion engines

Country Status (5)

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US (1) US5263441A (en)
EP (1) EP0455761B1 (en)
JP (1) JPH04502660A (en)
DE (2) DE3939065A1 (en)
WO (1) WO1991008385A1 (en)

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DE19949514A1 (en) * 1999-10-14 2001-04-19 Bosch Gmbh Robert Device for rapid pressure build-up in a motor vehicle device supplied with a pressure medium by a feed pump
WO2003018965A1 (en) 2001-08-21 2003-03-06 Robert Bosch Gmbh Valve mechanism with a variable valve opening diameter
WO2003018967A1 (en) 2001-08-21 2003-03-06 Robert Bosch Gmbh Valve mechanism comprising a variable cross-section of a valve opening

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US5829397A (en) * 1995-08-08 1998-11-03 Diesel Engine Retarders, Inc. System and method for controlling the amount of lost motion between an engine valve and a valve actuation means
AUPN678395A0 (en) * 1995-11-23 1995-12-14 Mitchell, William Richard Hydraulically or pneumatically actuated electronically controlled automotive valve system
ITTO980060A1 (en) * 1998-01-23 1999-07-23 Fiat Ricerche IMPROVEMENTS ON INTENRE COMBUSTION ENGINES WITH VARIABLE ACTING VALVES.
GB9906504D0 (en) * 1999-03-23 1999-05-12 Csa Performance Ltd Valve actuation means
US6135073A (en) * 1999-04-23 2000-10-24 Caterpillar Inc. Hydraulic check valve recuperation
ITTO20010269A1 (en) * 2001-03-23 2002-09-23 Fiat Ricerche INTERNAL COMBUSTION ENGINE, WITH HYDRAULIC VARIABLE VALVE OPERATION SYSTEM, AND MEANS OF COMPENSATION OF VOLUME VARIATIONS
ITTO20010270A1 (en) * 2001-03-23 2002-09-23 Fiat Ricerche INTERNAL COMBUSTION ENGINE WITH HYDRAULIC VARIABLE VALVE OPERATION SYSTEM AND DOUBLE PISTON PUNTER.
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DE10231143B4 (en) * 2002-07-10 2004-08-12 Siemens Ag Method for controlling the valve lift of discretely adjustable intake valves of a multi-cylinder internal combustion engine
JP2004197588A (en) * 2002-12-17 2004-07-15 Mitsubishi Motors Corp Valve system for internal combustion engine
FI117348B (en) * 2004-02-24 2006-09-15 Taimo Tapio Stenman Hydraulic device arrangement for controlling the operation of internal combustion engine valves
DE102010018209A1 (en) * 2010-04-26 2011-10-27 Schaeffler Technologies Gmbh & Co. Kg Hydraulic unit for a cylinder head of an internal combustion engine with hydraulically variable gas exchange valve drive
KR20120017982A (en) * 2010-08-20 2012-02-29 현대자동차주식회사 Electro-hydraulic variable vavlve lift apparatus
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Publication number Priority date Publication date Assignee Title
DE19949514A1 (en) * 1999-10-14 2001-04-19 Bosch Gmbh Robert Device for rapid pressure build-up in a motor vehicle device supplied with a pressure medium by a feed pump
DE19949514C2 (en) * 1999-10-14 2001-10-18 Bosch Gmbh Robert Device for rapid pressure build-up in a motor vehicle device supplied with a pressure medium by a feed pump
WO2003018965A1 (en) 2001-08-21 2003-03-06 Robert Bosch Gmbh Valve mechanism with a variable valve opening diameter
WO2003018967A1 (en) 2001-08-21 2003-03-06 Robert Bosch Gmbh Valve mechanism comprising a variable cross-section of a valve opening

Also Published As

Publication number Publication date
WO1991008385A1 (en) 1991-06-13
DE59004044D1 (en) 1994-02-10
JPH04502660A (en) 1992-05-14
EP0455761A1 (en) 1991-11-13
DE3939065A1 (en) 1991-05-29
US5263441A (en) 1993-11-23

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